JP4979653B2 - Variable compression ratio internal combustion engine - Google Patents

Description

  The present invention includes a piston inner connected to a connecting rod via a piston pin, and a piston outer supported by the piston inner so as to be movable in a cylinder axis direction and facing a combustion chamber, and the piston is operated during operation of the internal combustion engine. A variable compression ratio internal combustion engine that can change the position of the piston outer between a low compression ratio position near the piston inner and a high compression ratio position near the combustion chamber by the resultant force of the in-cylinder pressure and inertial force acting on the outer Related to institutions.

  A variable compression ratio internal combustion engine provided with a double piston composed of such a piston inner and a piston outer is known from Patent Document 1 below.

Also, in a variable compression ratio internal combustion engine that changes the compression ratio by moving the cylinder block and cylinder head up and down with respect to the crankshaft and piston by the cam mechanism and increasing or decreasing the volume of the combustion chamber, the driving force of the cam mechanism is reduced. A technique for changing the valve timings of the intake valve and the exhaust valve is known from Japanese Patent Application Laid-Open No. 2004-228561.
JP 2003-65090 A JP 2005-83361 A

  By the way, what is described in the above-mentioned patent document 1 is that the piston outer rises with respect to the piston inner when the low compression ratio → high compression ratio is switched, and the piston outer with respect to the piston inner when the high compression ratio → low compression ratio is switched. Since the outer descends, there is a problem that an impact is generated at the end of the movement, causing vibration and noise, and the impact becomes more significant as the internal combustion engine rotates at a higher speed.

  Further, in Patent Document 2, the driving force of the cam mechanism that changes the compression ratio is reduced by changing the valve timings of the intake valve and the exhaust valve. Since this acts in a direction that causes a sudden change, there is a possibility that a strong impact may occur when the change of the compression ratio is completed.

  The present invention has been made in view of the above-described circumstances, and an object thereof is to reduce an impact generated when a compression ratio of a variable compression ratio internal combustion engine provided with a piston inner and a piston outer is changed.

To achieve the above object, according to the first aspect of the present invention, a piston inner coupled to a connecting rod via a piston pin, and a combustion chamber supported by the piston inner so as to be movable in the cylinder axial direction. The piston outer is positioned near the low pressure ratio position near the piston inner and the combustion chamber by the resultant force of the in-cylinder pressure and the inertial force acting on the piston outer during operation of the internal combustion engine. In the variable compression ratio internal combustion engine that can be changed between the high compression ratio position, the cylinder pressure control means for controlling the cylinder pressure when the compression ratio is changed is provided , and the cylinder pressure control means when in the vicinity of the point is moved to the high compression ratio position said piston outer from the low compression ratio position, this to advance the closing timing of the exhaust valve as compared with the non-switching of the compression ratio Variable compression ratio internal combustion engine is proposed, wherein.

According to a second aspect of the invention, there is provided a piston inner coupled to the connecting rod via a piston pin, and a piston outer that is supported by the piston inner so as to be movable in the cylinder axial direction and faces the combustion chamber. The position of the piston outer is set between the low compression ratio position near the piston inner and the high compression ratio position near the combustion chamber by the resultant force of the in-cylinder pressure and the inertial force acting on the piston outer during operation of the internal combustion engine. In a variable compression ratio internal combustion engine that can be changed between, a cylinder pressure control means for controlling the cylinder pressure when the compression ratio is changed is provided, and the cylinder pressure control means controls the piston outer in the vicinity of the intake bottom dead center. The variable compression is characterized in that, when moving from the high compression ratio position to the low compression ratio position, the valve opening timing of the intake valve is retarded compared to when the compression ratio is not switched. Internal combustion engine is proposed.

  The variable valve timing mechanism 63 of the embodiment corresponds to the in-cylinder pressure control means of the present invention.

According to the present invention , the position of the piston outer is set to a low compression ratio position near the piston inner and a high compression ratio position near the combustion chamber by the resultant force of the in-cylinder pressure and the inertial force acting on the piston outer during operation of the internal combustion engine. When changing between the two, the in-cylinder pressure control means changes the in-cylinder pressure to control the resultant force of the in-cylinder pressure and the inertial force so that the piston outer does not move suddenly relative to the piston inner. Impact can be reduced, vibration and noise can be reduced, and the durability of the piston can be increased.

In particular , according to the invention of claim 1 , when the piston outer is moved from the low compression ratio position to the high compression ratio position to the combustion chamber side in the vicinity of the exhaust top dead center, the valve closing timing of the exhaust valve is advanced. Thus, the in-cylinder pressure can be kept high, and the moving speed of the piston outer to the combustion chamber side can be kept low to reduce the impact at the time of switching the compression ratio.

In particular , according to the invention of claim 2 , when the piston outer is moved from the high compression ratio position to the low compression ratio position toward the piston inner side in the vicinity of the intake bottom dead center, the opening timing of the intake valve is retarded. Thus, the in-cylinder pressure can be kept low, and the moving speed of the piston outer to the piston inner side can be suppressed to reduce the impact at the time of switching the compression ratio.

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

  1 to 23 show an embodiment of the present invention. FIG. 1 is a longitudinal sectional front view of an essential part of an internal combustion engine provided with a compression ratio variable device, and FIG. 2 is an exploded perspective view from above of the compression ratio variable device. 3 is an exploded perspective view of the variable compression ratio device from below, FIG. 4 is an enlarged view of the main part of FIG. 1 (low compression ratio state), FIG. 5 is a sectional view taken along line 5-5 in FIG. 5 is a sectional view taken along line 6-6 in FIG. 5, FIG. 7 is a sectional view taken along line 7-7 in FIG. 5, FIG. 8 is a sectional view taken along line 8-8 in FIG. 10 is a sectional view taken along line 10-10 of FIG. 9, FIG. 11 is a sectional view taken along line 11-11 of FIG. 10, FIG. 12 is a sectional view taken along line 12-12 of FIG. Fig. 14 is a cross-sectional view taken along line -13 (low compression ratio state), Fig. 14 is a diagram corresponding to Fig. 13 showing the high compression ratio state, and Fig. 15 is a diagram showing changes in the load acting on the piston (without valve timing control). 16 is a view showing a change in load acting on the piston (with valve timing control), FIG. 17 is a detailed view of the main part of FIG. 16, showing the switching from the low compression ratio to the high compression ratio, FIG. FIG. 19 is a diagram corresponding to FIG. 17 showing the switching of the high compression ratio to the low compression ratio, FIG. 19 is an explanatory diagram of the switching of the valve timing according to the rotational speed of the internal combustion engine, and FIG. FIG. 21 is a map for searching for the target compression ratio, FIG. 22 is an explanatory diagram of the vibration reduction effect when switching from the low compression ratio to the high compression ratio, and FIG. 23 is the high compression ratio → It is explanatory drawing of the vibration reduction effect at the time of switching of a low compression ratio.

  1 and 5, an engine body 1 of an internal combustion engine E includes a cylinder block 2 having a cylinder bore 2a, a crankcase 3 coupled to the lower end of the cylinder block 2, and a pent roof type connected to the upper end of the cylinder bore 2a. The cylinder head 4 has a combustion chamber 4a and is coupled to the upper end of the cylinder block 2. The cylinder head 4 is provided with an intake port 30i and an exhaust port 30e that open to the ceiling surface of the combustion chamber 4a. An intake valve 31i and an exhaust valve 31e that are opened and closed, and a spark plug 32 that faces an electrode at the center of the combustion chamber 4a are provided.

  A small end 7a of a connecting rod 7 is connected via a piston pin 6 to a piston 5 slidably fitted in the cylinder bore 2a, and a large end 7b of the connecting rod 7 is connected via a pair of left and right bearings 8 and 8. And connected to a crank pin 9a of a crank shaft 9 rotatably supported by the crank case 3.

  As shown in FIGS. 2 to 5, the piston 5 is slidably fitted to a piston inner 5 a connected to the small end portion 7 a of the connecting rod 7 through a piston pin 6 and an outer peripheral surface of the piston inner 5 a. The piston outer 5b has a top surface facing the combustion chamber 4a, and a plurality of piston rings 10a to 10c that are slidably in close contact with the inner peripheral surface of the cylinder bore 2a are mounted on the outer periphery of the piston outer 5b. Is done.

  The piston inner 5a slides on the inner peripheral surface of the cylinder bore 2a except for a pair of pin boss portions 11, 11 supporting both ends of the piston pin 6 and a portion corresponding to the outer ends of the pin boss portions 11, 11. A pair of arc-shaped skirt portions 12 and 12 that are freely fitted are integrally formed. The piston pin 6 is hollow.

  On the other hand, the piston outer 5b ends the peripheral wall on which the piston rings 10a to 10c are mounted at a position facing the upper end surface 12a of the skirt portions 12 and 12. The piston outer 5b is integrally formed with a pair of ear portions 13, 13 that are opposed to the outer ends of the pin boss portions 11, 11. These are provided with long holes 14 and 14 having a long diameter in the axial direction of the piston 5, and the long holes 14 and 14 have both end portions of the extension shaft 15 penetrating the hollow portion of the piston pin 6. The extension shaft 15 is slidably fitted along the axial direction of the piston 5 and is fixed to the piston pin 6 by press fitting or the like. Thus, by fitting the elongated holes 14 and 14 and the extension shaft 15, the piston inner 5 a and the piston outer 5 b are allowed to slide in the axial direction while being prevented from rotating relative to each other. , 14, the lower sliding limit of the piston inner 5 a with respect to the piston outer 5 b is regulated.

  A pair of inner sliding flat surfaces 23, 23 extending in the axial direction of the piston pin 5 are formed on both sides of the outer peripheral surface of the piston inner 5 a facing both end surfaces of the piston pin 6. , 23 are formed on the inner side surfaces of the ear portions 13, 13 of the piston outer 5b. The sliding flat surfaces 23, 24 are also formed on the piston inner 5a and the piston outer 5b. Relative sliding in the axial direction is allowed while preventing relative rotation. Therefore, the relative rotation of the piston inner 5a and the piston outer 5b is made strong by the fitting between the long holes 14, 14 and the extension shaft 15 and the contact between the inner and outer sliding flat surfaces 23, 24. Can be blocked.

  2, 3 and 5, the piston inner 5 a and the piston outer 5 b are configured such that the extension shaft 15 and the long holes 14 and 14 are slidably fitted, and a pair of arcuate surfaces 33 on the outer periphery of the piston inner 5 a. 33 and a slidable fitting of the inner peripheral surface 42a of the female spline 42 of the piston outer 5b, a sufficient axial relative sliding support length can be obtained, and stable axial relative sliding can be ensured. . The circular arc surfaces 33 and 33 are vertically formed so as to connect the upper end surfaces 12 a of the pair of skirt portions 12 and 12 and the first support surface 17.

  As clearly shown in FIG. 3 to FIG. 5, the upper part of the piston inner 5 a has an upward first annular support surface 17 in order from the outer peripheral side, and a first pivot that stands from the inner peripheral edge of the first support surface 17. 18, an annular second support surface 19 formed at the upper end of the first pivot 18, a second pivot 20 rising from the inner peripheral edge of the second support surface 19, and an upper end surface of the second pivot 20. An annular third support surface 21 is formed coaxially with the piston inner 5a. The second pivot 20 is divided into a plurality of blocks along the circumferential direction in order to reduce the weight, and an opening 22 is provided at the center of the second pivot 20 so that the small end 7a of the connecting rod 7 faces. The scattered lubricating oil generated in the crankcase 3, that is, in the crank chamber 3 a passes through the opening 22.

  An annular lock plate 25 mounted on the first support surface 17 is rotatably fitted to the first pivot 18 and is fitted to the second pivot 20 so as to face the upper surface of the lock plate 25. The annular first pressing plate 26 to be joined is fixed to the second support surface 19 by a plurality of screws 27, 27. Further, an annular lift member 28 placed on the first presser plate 26 is rotatably fitted to the second pivot 20, and a second presser plate 29 facing the upper surface of the inner peripheral edge of the lift member 28. Are fixed to the third support surface 21 by a plurality of screws 34, 34.

  The lift member 28 can reciprocate between a lift position B and a lift release position A set around the second pivot 20, and the piston outer 5 b is moved toward the piston inner 5 a with the reciprocal rotation. It constitutes a main part of a cam mechanism 37 that holds alternately at a position L (see FIGS. 4 and 5) and a high compression ratio position H (see FIGS. 9 and 10) near the combustion chamber 4a.

  That is, as shown in FIGS. 4, 5, and 8, the cam mechanism 37 includes the lift member 28 and a plurality of first cam peaks 38 in an annular arrangement that protrude integrally with the upper surface of the lift member 28. 38 ... and second cam peaks 39, 39 ... arranged in an annular arrangement projecting from the lower surface of the head portion of the piston outer 5b. Each of the cam peaks 38, 39 has a flat top surface. At the same time, both side surfaces aligned in the arrangement direction of the cam peaks 38 and 39 are formed in a rectangular cross-sectional shape that is a vertical surface with respect to the top surface.

  Thus, when the lift member 28 is in the lift release position A, the upper second cam peaks 39, 39... Can enter and leave the valley between the first cam peaks 38, 38. 13), the transition of the piston outer 5b to the low compression ratio position L or the high compression ratio position H is permitted. If the first and second cam peaks 38 and 39 mesh with each other and the top surface of at least one cam peak contacts the valley bottom between the other cam peaks, the cam mechanism 37 is in an axially contracted state, and the piston outer 5b is lowered. A compression ratio position L is given.

  When the lift member 28 is at the lift position B, the first and second cam peaks 38 and 39 abut each other on a flat top surface (see FIG. 14), so that the cam mechanism 37 is in the axially expanded state. Thus, the high compression ratio position H is given to the piston outer 5b. At this time, the extension shaft 15 fixed to the piston pin 6 as described above comes into contact with the lower side surfaces of the long holes 14 and 14 of the ear portions 13 and 13 in the piston outer 5b, whereby the piston outer 5b is compressed at a predetermined high compression. Moving beyond the specific position H to the combustion chamber 4a side is prevented.

  As shown in FIGS. 4, 5, and 7, the lock plate 25 reciprocates between an unlock position C (see FIG. 12) and a lock position D (see FIG. 7) set around the first pivot 18. The main part of the lock mechanism 40 that holds the axially contracted state of the cam mechanism 37 at the lock position D is formed.

  That is, the lock mechanism 40 is formed on the inner periphery of the piston outer 5b so that the lock plate 25, the male spline 41 formed on the outer periphery of the lock plate 25, and the male spline 41 are slidably fitted. The female spline 42 and an annular lock groove 43 that allows the upper end of the groove portion of the female spline 42 to communicate with each other and allow the teeth of the male spline 41 to rotate and enter, and the piston outer 5b has a low compression. When the position is switched between the specific position L and the high compression ratio position H, the lock plate 25 is set to the unlock position C, the male spline 41 is placed in a sliding relationship with the female spline 42, and the piston outer When 5b comes to the low compression ratio position L, the lock plate 25 is rotated to the lock position D so that the tooth portion of the male spline 41 enters the lock groove 43, and the tooth portion of the male spline 41 and the female spline 41 By abutting the end faces to each other with the teeth portion of the in-42, a low compression ratio position L of the piston outer 5b is locked.

  As shown in FIGS. 2 and 10, in order to strengthen the pressing of the lock plate 25 by the first pressing plate 26, it is arranged in a plurality of grooves of the male spline 41 and supports the lower surface of the outer peripheral portion of the first pressing plate 26. Are formed integrally with the piston inner 5a, and the outer periphery of the first pressing plate 26 is fixed to the bosses 35, 35 with a plurality of screws 27 ', 27'. Of course, the bosses 35, 35... Are formed so as not to prevent the male spline 41 from turning to the unlocked position C and the locked position D.

The piston inner 5a is provided with first and second actuators 45 ₁ and 45 _{2 for} driving the lift member 28 and the lock plate 25, respectively, with reference to FIGS. 5, 6, 13 and 14. explain.

First, a description from the first actuator 45 _1. The piston inner 5 a has a bottomed cylinder hole 46 ₁ extending in parallel to one side of the piston pin 6, and a series of lengths penetrating the upper wall of the intermediate part of the cylinder hole 46 ₁ and the first pressing plate 26. A hole 47 ₁ is provided, and a pressure receiving pin 48 ₁ protruding from the lower surface of the lift member 28 faces the cylinder hole 46 ₁ through the long hole 47 ₁ .

The pressure receiving pin 48 _1, disc-shaped slider 49 ₁ to obtain play in the radial direction in the cylinder bore 46 within ₁ loosely fitted in the cylinder bore 46 ₁ is mounted for relative swinging. An operating plunger 50 ₁ and a bottomed cylindrical return plunger 51 ₁ are slidably fitted into the cylinder hole 46 ₁ with the slider 49 ₁ interposed therebetween. Accordingly, the slider 49 ₁ is interposed between the pressure receiving pin 48 ₁ and the operating plunger 50 ₁ and the return plunger 51 ₁ , but the pressure receiving pin 48 ₁ around the rotation center of the lift member 28. The arc motion is allowed by the slider 49 ₁ moving in the cylinder hole 46 ₁ while sliding between the actuating plunger 50 ₁ and the return plunger 51 ₁ . Moreover, since the contact between the pressure receiving pin 48 ₁ and the operation plunger 50 ₁ and the return plunger 51 ₁ is always surface contact, the wear resistance of the contact portion is ensured.

A hydraulic chamber 52 ₁ facing the inner end of the operating plunger 50 ₁ is defined in the cylinder hole 46 ₁ , and when hydraulic pressure is supplied to the hydraulic chamber 52 ₁ , the operating plunger 50 ₁ receives the hydraulic pressure and the slider 50 ₁ receives the hydraulic pressure. The lift member 28 is rotated to the lift position B via 49 ₁ and the pressure receiving pin 48 ₁ , and the elongated hole 47 ₁ has a size that does not hinder the movement of the pressure receiving pin 48 ₁ at that time. ing.

Also the open end of the cylinder bore 46 _1, the cylindrical spring holding tube 53 ₁ is engaged through the retaining ring 54 _1, between the plunger 51 ₁ returning said this spring holding cylinder 53 _1, its return spring 55 ₁ return for biasing the plunger 51 ₁ on the pressure receiving pin 48 ₁ side is provided in a compressed state.

Thus, the lift release position A of the lift member 28 is defined by the pressure receiving pin 48 ₁ coming into contact with the inner end wall of the elongated hole 47 _{1 on} the side of the actuating plunger 51 ₁ (see FIG. 13). position B is defined by the pressure receiving pin 48 ₁ comes into contact with the spring holding tube 53 ₁ through a slider 49 ₁ and return the plunger 51 ₁ (see FIG. 14).

Second actuator 45 _2, first actuator 45 ₁ and is axially symmetrical or point-symmetrically disposed across the piston pin 6, the pressure receiving pin 48 ₂ is protruded from the lower surface of the lock plate 25. Since the other configuration is the same as that of the first actuator 45 ₁ , in the drawing, the parts corresponding to the first actuator 45 ₁ are denoted by the same reference numerals with only the subscript “ ₂ ”, and the details thereof are described. Description is omitted.

Thus, unlocking position C of the lock plate 25, the pressure receiving pin 48 ₂ elongated hole 47 ₂ is defined by abutting the actuating plunger 50 ₂ side inner end wall, the locking position D of the locking plate 25, the pressure receiving is defined by the pin 48 ₂ is brought into contact with the spring holding cylinder 53 ₂ via the slider 49 ₂ and the return plunger 51 _2.

By the way, if the operation strokes of the pressure receiving pins 48 ₁ and 48 ₂ are defined by the inner end walls of the long holes 47 ₁ and 47 ₂ , the operation strokes of the pressure receiving pins 48 ₁ and 48 ₂ can be defined with high accuracy. If the operation strokes of the pressure receiving pins 48 ₁ and 48 ₂ are restricted by bringing the operation plungers 50 ₁ and 50 ₂ and the return plungers 51 ₁ and 51 ₂ into contact with the inner end walls of the cylinder holes 46 ₁ and 46 ₂ , respectively. The load can be removed from the pressure receiving pins 48 ₁ and 48 _{2 at} the operation limit of the pressure receiving pins 48 ₁ and 48 ₂ .

Thus, the first and second actuators 45 ₁ and 45 ₂ are configured to have substantially the same structure, and below the lift member 28 and the lock plate 25 that are stacked one above the other with the first pressing plate 26 interposed therebetween. It arrange | positions so that the axis line of piston inner 5a may be pinched | interposed. In addition, compatibility is given to parts corresponding to each other of the first and second actuators 45 ₁ and 45 ₂ . As a result, the components of the first and second actuators 45 ₁ and 45 ₂ can be shared, and the cost can be greatly reduced.

As shown in FIGS. 1 and 6, a cylindrical oil chamber 57 is defined between the piston pin 6 and the extension shaft 15 fitted in the hollow portion thereof. First and second distribution oil passages 58 ₁ , 58 ₂ connected to the hydraulic chambers 52 ₁ , 52 ₂ of the second actuators 45 ₁ , 45 ₂ are provided across the piston pin 6 and the piston inner 5a. The oil chamber 57 is connected to an oil passage 59 provided across the piston pin 6, the connecting rod 7 and the crankshaft 9. The oil passage 59 is connected to an oil pump 61 serving as a hydraulic pressure source via an electromagnetic switching valve 60, and The oil sump 62 is switchably connected. The oil sump 62 is an oil pan attached to the bottom of the crankcase 3, and therefore the lubricating oil of the internal combustion engine E is used as the operating oil for the first and second actuators 45 ₁ and 45 ₂ .

  In FIG. 4, the extension shaft 15 has a hollow portion 15b whose open surfaces at both ends are closed by end plates 15a and 15a. The hollow portion 15b is a piston through a through hole 16a in the central portion of the extension shaft 15. The hollow portion 15 b communicates with a cylindrical oil chamber 57 in the pin 6, and the hollow portion 15 b is inserted into the long holes 14, 14 of the ear portions 13, 13 through the injection holes 16 b, 16 b at both ends of the extension shaft 15. Communicated. At that time, it is desirable that the nozzle holes 16b at the respective ends of the extension shaft 15 are arranged so as to open toward the lower end surface of the corresponding long holes 14, and in the illustrated example, the nozzle holes 16b are arranged at the ends of the extension shaft 15. Even if the piston pin 6 rotates, at least one injection hole 16b is directed to the lower end surface of the long hole 14 even if the piston pin 6 rotates.

  As shown in FIGS. 5 and 10, a variable valve timing mechanism 63 capable of changing the valve timing is connected to the intake valve 31i and the exhaust valve 31e. For example, as described in JP-A-7-332051, a plurality of rocker arms driven by a plurality of valve cams having different shapes are connected to the variable valve timing mechanism 63 by hydraulically operated pins. By releasing the connection, the valve timing of the intake valve 31i and the exhaust valve 31e is changed, or as described in JP 2000-227033 A, provided at the shaft end of the camshaft and driven by a timing chain There are some which are provided between the sprocket and the camshaft, and change the valve timing of the intake valve 31i and the exhaust valve 31e by hydraulically changing the phase of the camshaft with respect to the phase of the sprocket.

  Next, the operation of the embodiment of the present invention having the above configuration will be described.

  First, changes in the load acting on the piston outer 5b will be described with reference to FIG. The load acting on the piston outer 5b includes an in-cylinder pressure load based on the pressure of the combustion chamber 4a and an inertial force load based on the raising and lowering of the piston 5, and the load (hereinafter referred to as the load) acting on the piston outer 5b. , Referred to as piston load).

  The in-cylinder pressure suddenly increases during the compression stroke, reaches a peak value at the compression top dead center, then decreases rapidly during the expansion stroke, and reaches a bottom value of approximately atmospheric pressure at the bottom expansion dead point. It is maintained at a value near the bottom value in the intake stroke. Therefore, when the upward piston load is positive and the downward piston load is negative, the in-cylinder pressure load decreases rapidly from a substantially zero value in the compression stroke and reaches the bottom value at the compression top dead center. It rapidly increases in the expansion stroke, returns to a substantially zero value at the expansion bottom dead center, and is maintained at a substantially zero value in the subsequent exhaust stroke and intake stroke. On the other hand, the inertial load changes to a sine curve, increases from a negative value to a positive value in the compression stroke, decreases from a positive value to a negative value in the expansion stroke, increases from a negative value to a positive value in the exhaust stroke, and is inhaled. Decrease from positive value to negative value in the process.

  As a result, the piston load acting on the piston outer 5b becomes a positive value (upward) in the period from the latter half of the exhaust stroke to the first half of the intake stroke through the exhaust top dead center, and becomes negative (downward) in the period before and after that. Become.

  As shown in FIGS. 3 to 8 and 13, the lift member 28 of the cam mechanism 37 is in the lift release position A, and the lock plate 25 is engaged with the lock groove 43. Is assumed to be held at the low compression ratio position L on the piston inner 5a side. Therefore, the compression ratio of the internal combustion engine E operated in this state is controlled to be relatively low.

In order to obtain a high compression ratio state in order to improve the output, for example, at high speed operation of the internal combustion engine E from such a state, the switching valve 60 is energized, that is, the on state, and the oil passage 59 is connected to the oil pump 61. Connecting. In this way, the hydraulic oil discharged from the oil pump 61 passes through the upstream oil passage 59a of the crankshaft 9, the downstream oil passage 59b of the connecting rod 7, the first and second distribution oil passages 58 ₁ and 58 ₂ , and the first and second distribution oil passages 58 ₁ and 58 _2. It is supplied to the hydraulic chambers 52 ₁ , 52 ₂ of the second actuators 45 ₁ , 45 ₂ .

Then, first, the pressure receiving pin 48 ₂ receive in conjunction with slider 49 ₂ hydraulic plunger 50 of the _second actuator 45 ₂ hydraulic chamber 52 ₂ of the hydraulic, so to press against the biasing force of the return spring 55 _2, receiving pin 48 ₂ rotates the lock plate 25 from the locking position D (see FIG. 7) to the unlocked position C (see FIG. 9), the female spline 42 of the male spline 41 and the piston outer 5b of the lock plate 25 The sliding fitting is possible.

  Therefore, the piston outer 5b moves to the high compression ratio position H by the action of the following piston load. That is, as described with reference to FIG. 15, when the piston load that has been acting downward until then changes near the exhaust top dead center, the piston outer 5b moves away from the piston inner 5a toward the combustion chamber 4a. Accordingly, the extension shaft 15 that is displaced and supported by the piston inner 5a relatively descends the long holes 14 and 14 of the ear portions 13 and 13 of the piston outer 5b, and the lower end walls of the long holes 14 and 14 are moved downward. The piston outer 5b is prevented from being displaced at a predetermined high compression ratio position H.

Therefore, the movement limit of the piston outer 5b toward the high compression ratio position can be restricted without using a special stopper member, which can contribute to simplification of the structure of the device. Moreover, the impact at the time of restricting the movement limit of the piston outer 5b to the high compression ratio position side is directly transmitted from the piston outer 5b to the piston pin 6 via the long holes 14 and 14 and the extension shaft 15 that are in contact with each other. Since it is not transmitted to the piston inner 5a, it is possible to prevent the influence of the impact on the cam mechanism 37, the lock mechanism 40, the first and second actuators 45 ₁ , 45 _{2 and the} like provided in the piston inner 5a. Durability and operational stability can be ensured.

When the piston outer 5b comes to the high compression ratio position H, the first cam peaks 38, 38... Of the lift member 28 disengage from the valleys between the second cam peaks 39, 39. in 45 _1, already pushed against the urging force of the hydraulic chamber 52 actuating plunger 50 ₁ had received the oil pressure of ₁ returns the pressure receiving pin 48 ₁ with the slider 49 _first spring 55 _1, the lift member 28 It rotates from the lift release position A to the lift position B. Therefore, as shown in FIG. 14, the first cam peaks 38, 38... And the second cam peaks 39, 39. That is, the cam mechanism 37 is in the axially expanded state.

  In this way, the piston outer 5b is held at the high compression ratio position H by the axially expanded state of the cam mechanism 37 and the contact between the extension shaft 15 and the lower end walls of the elongated holes 14 and 14. Therefore, the piston inner and the outer 5a, 5b can move up and down in the cylinder bore 2a integrally while increasing the compression ratio, and can contribute to the improvement of the output performance of the internal combustion engine E. Moreover, in the cam mechanism 37, the contact surfaces of the top surfaces of the first and second cam peaks 38 and 39 in the annular arrangement that contact each other are evenly distributed over the entire circumference of the piston 5, and the total area thereof is wide. The cam mechanism 37 can sufficiently withstand a large in-cylinder pressure in the expansion stroke and compression stroke of the internal combustion engine E.

Next, in order to switch the internal combustion engine E from the high compression ratio state to the low compression ratio state again, the switching valve 60 is turned off, that is, a non-energized state, and the oil passage 59 is opened to the oil sump 62. Then, the hydraulic chambers 52 ₁ and 52 ₂ of the first and second actuators 45 ₁ and 45 ₂ connected to the downstream oil passage 59b are depressurized, and the pressure receiving pins 48 ₁ and 48 of the first and second actuators 45 ₁ and 45 ₂ are reduced. ₂ is placed under the control of return plungers 51 ₁ and 51 ₂ that receive the urging forces of the return springs 55 ₁ and 55 ₂ , respectively.

Thus, when the hydraulic pressure chamber 52 _1, 52 ₂ is depressurized, first, the first actuator 45 _1, returns the plunger 51 ₁ and pushed into the hydraulic chamber 52 ₁ side pressure receiving pin 48 ₁ with slider 49 _1, lift The member 28 is rotated to the lift release position A, and the first cam peaks 38, 38... And the second cam peaks 39, 39. When the piston load that has been acting upward until then changes in the vicinity of the suction bottom dead center, as shown in FIG. 13, the piston outer 5b has first cam peaks 38, 38... And second cam peaks 39. , 39..., 39 are meshed with each other and displaced so as to be close to the piston inner 5 a, and the top of one cam crest 39 hits the bottom of the valley between the other cam crests 38. Position L is determined

When the piston outer 5b reaches the low compression ratio position L, the male splines 41 of the lock plate 25, since it is possible enter the locking groove 43 of the piston outer 5b, the spring plunger 51 ₂ return of the second actuator 45 ₂ returns 55 the pressure receiving pin 48 ₂ with slider 49 _{2 2} with a force and pushes the hydraulic chamber 52 ₂ side, rotates the lock plate 25 to the locking position D, and the locking mechanism 40 in a locked state. That is, the male spline 41 of the lock plate 25 is opposed to the upper end surface of the female spline 42 of the piston outer 5b, thereby preventing the two splines 41 and 42 from sliding relative to each other.

  Next, the principle of reducing the switching vibration by advancing the valve timing of the exhaust valve 31e when switching from the low compression ratio to the high compression ratio will be described with reference to FIGS. In FIG. 17, among the characteristic curves, the thin line corresponds to the case without valve timing control, and the thick line corresponds to the case with valve timing control.

  Normally, the exhaust valve 31e is closed near the exhaust top dead center, but the valve closing timing of the exhaust valve 31e is advanced and closed earlier than usual so that the exhaust valve 31e enters the combustion chamber 4a at the end of the exhaust stroke. The exhaust pressure is confined and the in-cylinder pressure can be kept high. As a result, the timing at which the combined load of the in-cylinder pressure load and the inertial force load becomes a positive value (upward) is delayed compared to the case where the valve timing of the exhaust valve 31e is not controlled (see FIG. 15), and exhaust top dead Move to just before the point.

  Switching from the low compression ratio to the high compression ratio is performed when the combined load of the in-cylinder pressure load and the inertial force load becomes a positive value (upward), but the valve timing of the exhaust valve 31e is not controlled (see FIG. 15), the combined load becomes a positive value early, so that when the low compression ratio is switched to the high compression ratio in the vicinity of the exhaust top dead center, the upward combined load becomes excessive and the piston outer 5b May suddenly rise toward the high compression ratio position and generate a large impact.

  On the other hand, according to the present embodiment, since the upward combined load in the vicinity of the exhaust top dead center is kept small, when the low compression ratio → the high compression ratio is switched in the vicinity of the exhaust top dead center, the upward load Is maintained at an appropriate magnitude, and the speed at which the piston outer 5b rises toward the high compression ratio position can be suppressed to prevent the occurrence of a large impact.

  If the valve timing of the exhaust valve 31e is excessively advanced, the timing at which the combined load becomes a positive value (upward) will be after the exhaust top dead center, and the piston outer 5b will rise after the piston 5 starts to descend. The compression ratio is switched, and an extremely large impact may occur. Therefore, it is desirable to set the advance amount of the valve timing of the exhaust valve 31e so that the timing at which the combined load becomes a positive value (upward) is immediately before the exhaust top dead center.

  Next, the principle of reducing the switching vibration by retarding the valve timing of the intake valve 31i when switching from the high compression ratio to the low compression ratio will be described with reference to FIG. In FIG. 18, among the characteristic curves, the thin line corresponds to the case without valve timing control, and the thick line corresponds to the case with valve timing control.

  Normally, the intake valve 31i opens near the exhaust top dead center, but by opening the intake valve 31i later than usual by retarding the valve opening timing, the combustion chamber 4a is opened in the first half of the intake stroke. Can be sealed to keep the in-cylinder pressure low. As a result, the timing at which the combined load of the in-cylinder pressure load and the inertial force load becomes negative (downward) is delayed than usual, and when the high compression ratio → low compression ratio is switched near the suction bottom dead center Further, it is possible to prevent the downward composite load from becoming excessive and to suppress the speed at which the piston outer 5b descends toward the low compression ratio position, thereby preventing the occurrence of a large impact.

  As shown in FIG. 19, the variable valve timing mechanism 63 according to the present embodiment is configured such that the valve timing of the exhaust valve 31e is the normal valve timing, the first valve timing that is advanced, and the first valve timing that is further advanced. The valve timing of the intake valve 31i can be changed in two stages of the normal valve timing and the third valve timing retarded.

  In FIG. 19A, the horizontal axis represents the rotational speed Ne of the internal combustion engine E, and the vertical axis represents the crank angle CA in which the piston load increases, and the piston load increases upward as the rotational speed Ne increases. The crank angle CA is advanced. The set angle is the limit crank angle CA for preventing the impact at the time of switching from the low compression ratio to the high compression ratio from exceeding the design value. When the crank angle CA exceeds the set angle, the impact at the time of switching exceeds the design value. It will be.

  When the rotational speed Ne of the internal combustion engine E is equal to or less than the threshold value A, the crank angle CA at which the piston load is upward does not exceed the set angle even at the normal valve timing, and the impact at the time of switching does not exceed the design value. When the rotational speed Ne exceeds the threshold value A, the impact at the time of switching exceeds the design value at the normal valve timing. Therefore, by advancing to the first valve timing, the crank angle CA that causes the piston load to rise upward is less than the set angle. Can be suppressed.

  If the rotational speed Ne exceeds the threshold value B, the crank angle CA at which the piston load increases upwards exceeds the set angle, and the impact at the time of switching exceeds the design value at the first valve timing, so the angle is advanced to the second valve timing. Thus, the crank angle CA at which the piston load is upward can be suppressed to less than the set angle. When the rotational speed Ne exceeds the threshold value C, the crank angle CA at which the piston load becomes upward exceeds the set angle, and the impact at the time of switching exceeds the design value at the second valve timing, so switching from the low compression ratio to the high compression ratio. By prohibiting the crank angle CA, the crank angle CA at which the piston load is upward is suppressed to less than the set angle.

  In FIG. 19B, the horizontal axis is the rotational speed Ne of the internal combustion engine E, and the vertical axis is the crank angle CA at which the piston load is downward. The piston load is downward as the rotational speed Ne increases. The crank angle CA is advanced. The set angle is the limit crank angle CA for preventing the impact at the time of switching from the high compression ratio to the low compression ratio from exceeding the design value. When the crank angle CA exceeds the set angle, the impact at the time of switching exceeds the design value. It will be.

  When the rotational speed Ne of the internal combustion engine E is equal to or less than the threshold value D, the crank angle CA at which the piston load is downward does not exceed the set angle even at normal valve timing, and the impact at the time of switching does not exceed the design value. When the rotational speed Ne exceeds the threshold value D, the crank angle CA at which the piston load becomes downward exceeds the set angle. Since the impact at the time of switching exceeds the design value at the normal valve timing, the crank angle CA is retarded to the third valve timing. The crank angle CA at which the piston load is downward can be suppressed to less than the set angle.

  When the rotational speed Ne exceeds the threshold E, the crank angle CA at which the piston load becomes downward exceeds the set angle, and the impact at the time of switching exceeds the design value at normal valve timing, so switching from a high compression ratio to a low compression ratio. By prohibiting, the crank angle CA at which the piston load is downward is suppressed to less than the set angle.

  In addition, since there is a tendency that knocking is likely to occur during high-load operation, there is a case where the high compression ratio is changed to the low compression ratio in order to prevent the knocking. In other words, there is a possibility that a high compression ratio → low compression ratio switching prohibition request and a high compression ratio → low compression ratio switching request may antagonize in a high rotation / high load state. Therefore, when the rotational speed Ne is predicted to be equal to or greater than the threshold value E, it is possible to take measures against the conflicting requirements by switching to the low compression ratio in advance as feedforward.

  Next, the contents of the valve timing change control for reducing the impact at the time of switching the compression ratio will be described again based on the flowchart of FIG.

  First, in step X1, operating conditions such as the rotational speed and load of the internal combustion engine E are detected, and in step X2, the rotational speed and load are applied to the map of FIG. As a result, when the high compression ratio command is output in step X3, if the switching to the high compression ratio is completed in step X4, the operation at the normal valve timing is performed in the high compression ratio state in step X5. If the switching to the high compression ratio is not completed in step X4 and the rotational speed is equal to or less than the threshold value A in step X6, the operation at the normal valve timing is performed in the high compression ratio state in step X5.

  If the rotation speed Ne exceeds the threshold value A and is equal to or less than the threshold value B in step X7, the operation at the first valve timing is performed in the high compression ratio state in step X8. If the rotational speed exceeds the threshold value B and is equal to or less than the threshold value C in step X9, the operation at the second valve timing is performed in the high compression ratio state in step X10. If the rotational speed exceeds the threshold value C in step X9, switching from the low compression ratio to the high compression ratio is prohibited in step X11, and operation at the normal valve timing is performed in the low compression ratio state in step X12.

  If the high compression ratio command is not output in step X3 (that is, the low compression ratio command is output) and the switching to the low compression ratio is completed in step X13, the low compression ratio is determined in step X12. The normal valve timing operation is performed in this state. If the switching to the low compression ratio is not completed in step X13 and the rotational speed is equal to or less than the threshold value D in step X14, the operation at the normal valve timing is performed in the low compression ratio state in step X12.

  If the rotational speed exceeds the threshold value D and is equal to or less than the threshold value E in step X15, the operation at the third valve timing is performed in the low compression ratio state in step X16. If the rotational speed Ne exceeds the threshold value E in the step X15, switching from the high compression ratio to the low compression ratio is prohibited in the step X17, and the operation at the normal valve timing is performed in the high compression ratio state in the step X18.

  FIG. 22 shows the effect at the time of switching from the low compression ratio to the high compression ratio.

  When there is no control for advancing the valve timing at which the exhaust valve 31e closes, the switching vibration increases from small to medium to large as the rotational speed of the internal combustion engine E increases.

  When the valve timing can be advanced by one step, the switching vibration is small even at the normal valve timing when the rotation speed is low, and the switching vibration can be suppressed small by retarding to the first valve timing when the rotation speed is medium. it can. However, when the rotational speed is high, it is not sufficient to advance to the first valve timing, and the switching vibration increases to a medium level.

  When the valve timing can be advanced in two stages, the switching vibration is small even at the normal valve timing when the rotational speed is low, and the switching vibration is suppressed by advancing to the first valve timing when the rotational speed is medium. When the rotational speed is high, the switching vibration can be suppressed small by advancing to the second valve timing.

  FIG. 23 shows the effect at the time of switching from the high compression ratio to the low compression ratio.

  When the switching vibration increases as the rotational speed of the internal combustion engine E increases, the switching vibration can be reduced by retarding the valve timing at which the intake valve 31i closes.

  The embodiments of the present invention have been described above, but various design changes can be made without departing from the scope of the present invention.

  For example, the variable compression ratio device of the embodiment is configured to switch the compression ratio between two levels of high and low, but may be one that switches the compression ratio to three or more levels.

Longitudinal front view of main part of internal combustion engine provided with variable compression ratio device Exploded perspective view of the compression ratio variable device from above An exploded perspective view from below of the variable compression ratio device 1 is an enlarged view of the main part (low compression ratio state). Sectional view along line 5-5 in FIG. 6-6 sectional view of FIG. Sectional view along line 7-7 in FIG. 8-8 sectional view of FIG. Corresponding figure with Fig. 4 showing high compression ratio state Sectional view taken along line 10-10 in FIG. 11-11 sectional view of FIG. 12-12 sectional view of FIG. Sectional view taken along line 13-13 in FIG. 5 (low compression ratio state) Corresponding figure with FIG. 13 showing the high compression ratio state Diagram showing change in load acting on piston (without valve timing control) Diagram showing change in load acting on piston (with valve timing control) FIG. 16 is a detailed view of the main part of FIG. 16, showing the switching from the low compression ratio to the high compression ratio. Correspondence diagram with FIG. 17 showing the switching from the high compression ratio to the low compression ratio Explanatory drawing of switching of valve timing according to the rotation speed of the internal combustion engine Flow chart of valve timing change control for reducing impact at compression ratio switching Map to search for target compression ratio Illustration of vibration reduction effect when switching from low compression ratio to high compression ratio Illustration of vibration reduction effect when switching from high compression ratio to low compression ratio

Explanation of symbols

E Internal combustion engine 4a Combustion chamber 5a Piston inner 5b Piston outer 6 Piston pin 7 Connecting rod 31e Exhaust valve 31i Intake valve 63 Variable valve timing mechanism (in-cylinder pressure control means)

Claims (2)

A piston inner (5a) connected to the connecting rod (7) via a piston pin (6), and a piston outer (5a) supported by the piston inner (5a) so as to be movable in the cylinder axis direction and facing the combustion chamber (4a) 5b), and the position of the piston outer (5b) is moved closer to the piston inner (5a) by the resultant force of the cylinder pressure and the inertial force acting on the piston outer (5b) during operation of the internal combustion engine (E). In a variable compression ratio internal combustion engine that can be changed between a low compression ratio position of the engine and a high compression ratio position near the combustion chamber (4a),
In-cylinder pressure control means (63) for controlling the in-cylinder pressure when changing the compression ratio ,
When the piston outer (5b) is moved from the low compression ratio position to the high compression ratio position near the exhaust top dead center, the in-cylinder pressure control means (63) is more exhausted than when the compression ratio is not switched. variable compression ratio internal combustion institutions, characterized by advancing the closing timing of the valve (31e). A piston inner (5a) connected to the connecting rod (7) via a piston pin (6), and a piston outer (5a) supported by the piston inner (5a) so as to be movable in the cylinder axis direction and facing the combustion chamber (4a) 5b), and the position of the piston outer (5b) is moved closer to the piston inner (5a) by the resultant force of the cylinder pressure and the inertial force acting on the piston outer (5b) during operation of the internal combustion engine (E). In a variable compression ratio internal combustion engine that can be changed between a low compression ratio position of the engine and a high compression ratio position near the combustion chamber (4a),
In-cylinder pressure control means (63) for controlling the in-cylinder pressure when changing the compression ratio,
The in-cylinder pressure control means (63) when said moving to the low compression ratio position the piston outer (5b) in the vicinity of the intake bottom dead center from the high compression ratio position, as compared to the non-switching of the compression ratio variable compression ratio internal combustion engine, characterized in that retarding the opening timing of the intake valve (31i).

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