JPH03213630A - Variable compression ratio device of internal combustion engine - Google Patents

Abstract

PURPOSE:To carry out a highly reliable compression ratio variable operation by providing an eccentric sleeve to make the bearing hole of a connecting rod and a supporting shaft eccentric each other at the pivoting positions of the connecting rod of a piston, and fixing the eccentric sleeve at the plural positions with a pin member reciprocated by a hydraulic driving mechanism. CONSTITUTION:In a connecting rod 6 whose both ends are pivoted with the piston pin 7 of a piston 8 and the crankpin 2 of a crankshaft, an eccentric sleeve 5 to make the bearing hole of the connecting rod 6 and the crankpin 2 eccentric each other is provided rotatable, at the crankpin 2 and its pivoting position. And a stopper pin 12 movable in the axial direction of the shaft center sleeve 5 is provided, and the pin 12 is moved laterally by a hydraulic driving mechanism 11A made by inserting a piston member 12a in an oil pressure chamber 13 to which the pressure oil is fed and exhausted through oil pressure passages 17 and 18. And by fitting the stopper pin 12 to a fitting part 5a or 5b formed to the eccentric sleeve 5, the eccentric sleeve 5 can be fixed at the maximum eccentric position or at the minimum eccentric position.

Description

DETAILED DESCRIPTION OF THE INVENTION [Field of Industrial Application] The present invention relates to a variable compression ratio device for an internal combustion engine (hereinafter referred to as "engine" as necessary).

[Prior Art] Conventionally, in order to prevent knocking in a high load range or high engine speed range that is larger than the engine load range, and to improve fuel efficiency by increasing thermal efficiency below the middle load range, Various engines with variable compression ratios have been proposed.

Such a variable compression ratio mechanism is disclosed in Japanese Patent Publication No. 63-3297, for example.
It is disclosed in Publication No. 2. The variable compression ratio mechanism disclosed in this publication includes an eccentric bearing on one of the shaft supports at both ends of the connecting rod of the engine, which makes the bearing hole of the connecting rod and the support shaft inserted through the bearing hole eccentric to each other. The eccentric bearing is designed to rotate freely by the rotational force generated by the eccentric load and the reaction force from the support shaft, and furthermore, by driving a lock pin that is movable in the radial direction of the bearing, the eccentric bearing can be rotated freely. Provided with hydraulically actuated locking means for switching between fixed and
The pressure of the operating oil supplied to this locking means is determined by a signal from a computer that receives signals from the piston position detection means and the operating condition detection means, and hydraulic pressure is constantly applied to the locking means during locking, and during unlocking. The locking mechanism is controlled under conditions where no hydraulic pressure is applied to the locking means.

[Problems to be Solved by the Invention] However, in such a conventional variable compression ratio mechanism of an internal combustion engine, the pressure of the hydraulic oil supplied to the locking means is constantly applied to the locking means during locking. However, in this case, since the hydraulic oil acts only on one side of the hydraulic chamber that is provided between the pistons, the hydraulic pressure in the hydraulic chamber fluctuates greatly due to centrifugal force and the acceleration of the reciprocating movement of the connecting rod. This poses a problem in that the locking operation may become uncertain, and as a result, the reliability at the time of locking decreases.

The present invention is an attempt to solve these problems, and aims to provide a variable internal combustion engine that prevents the locking/unlocking operation of the locking means from becoming uncertain due to centrifugal force or acceleration of the reciprocating motion of the connecting rod. The purpose is to provide a compression ratio device.

Means for Solving Problem C] Therefore, the variable compression ratio device for an internal combustion engine according to the present invention according to claim 1 has one end pivotally supported by a piston that reciprocates within the cylinder of the internal combustion engine, and the other end. A connecting rod is provided whose end is pivotally supported by a crankshaft, and a bearing hole of the connecting rod and a support shaft passing through the bearing hole are eccentrically arranged in one of the pivot parts at both ends of the connecting rod. An eccentric sleeve is rotatably provided, and eccentric sleeve locking means capable of fixing rotation of the eccentric sleeve at a predetermined position is provided, and the eccentric sleeve locking means engages with an engaging portion formed on the eccentric sleeve. and a piston-type fluid pressure drive mechanism that can drive the pin member by moving a piston portion connected to the pin member, and the piston-type fluid pressure drive mechanism Fluid pressure chambers are provided on both sides of the piston portion, and one fluid pressure chamber is provided with a return spring that biases the piston portion toward the other fluid pressure chamber, and both fluid pressure chambers are provided with a return spring that biases the piston portion toward the other fluid pressure chamber side. means for applying a required fluid pressure; It is also characterized by having means for applying high fluid pressure.

Further, the variable compression ratio device for an internal combustion engine according to the present invention according to claim 2 has one end pivotally supported by a piston that reciprocates within a cylinder of the internal combustion engine, and the other end pivotally supported by a crankshaft. A connecting rod is provided, and an eccentric sleeve is rotatably provided on one of the pivot portions at both ends of the connecting rod to make a bearing hole of the connecting rod and a support shaft inserted through the bearing hole eccentric with respect to each other. Eccentric sleeve locking means capable of fixing rotation of the eccentric sleeve at a predetermined position is provided, the eccentric sleeve locking means comprising a pin member engageable with an engagement portion formed on the eccentric sleeve, and a pin member engageable with an engagement portion formed on the eccentric sleeve; A piston-type fluid pressure drive mechanism capable of driving the pin member by moving a connected piston part, and the piston-type fluid pressure drive mechanism has fluid pressure chambers on both sides of the piston part. At the same time, a pressure difference is generated between the two fluid pressure chambers, and fluid pressure is applied to both fluid pressure chambers so that the required fluid pressure is maintained even on the low pressure side, and the fluid pressure in one fluid pressure chamber is higher. The fluid pressure chamber is characterized by having a means for switching between a state in which the fluid pressure in the other fluid pressure chamber is higher.

[Function] First, in the variable compression ratio device for an internal combustion engine according to the present invention according to claim 1, a required fluid pressure is applied in advance to both fluid pressure chambers formed on both sides of the piston portion of the piston type fluid pressure drive mechanism. Apply it.

For example, when moving the piston toward one fluid pressure chamber to drive the pin member in a certain direction, a fluid pressure higher than the required fluid pressure is applied to the other fluid pressure chamber.

As a result, the piston portion moves toward one of the fluid pressure chambers against the biasing force of the return spring.

Furthermore, when the piston portion is moved toward the other fluid pressure chamber and the pin member is driven in the opposite direction, the other fluid pressure chamber is returned to the original required fluid pressure state. As a result, the piston portion moves toward the other fluid pressure chamber by the urging force of the return spring.

Next, in the variable compression ratio device for an internal combustion engine according to the second aspect of the present invention, a pressure difference occurs between both fluid pressure chambers formed on both sides of the piston portion in the piston type fluid pressure drive mechanism, even on the low pressure side. Fluid pressure is applied to both fluid pressure chambers to have the required fluid pressure.

For example, when moving the piston toward one fluid pressure chamber to drive the pin member in a certain direction, the fluid pressure in the other fluid pressure chamber is set to be higher. As a result, the piston portion moves toward one of the fluid pressure chambers due to the differential pressure between the two fluid chambers.

Further, when moving the piston portion toward the other fluid pressure chamber and driving the pin member in the opposite direction, the fluid pressure in one fluid pressure chamber is switched to a higher state.

As a result, the piston portion moves toward the other fluid pressure chamber due to the differential pressure between the two fluid chambers.

[Examples] Examples of the present invention will be described below with reference to the drawings.
Figures 1 to 8 show a variable compression ratio device for an internal combustion engine as a first embodiment of the present invention. Figure 1 is an overall configuration diagram showing the situation when it is in a low compression ratio state, and Figure 2 is a Figure 3 is a front view of the connecting rod showing how it looks when it is in a low compression ratio state, Figure 3 is an overall configuration diagram showing how it looks when it is in a high compression ratio state, and Figure 4 shows how it looks when it is in a high compression ratio state. The front view of the connecting rod shown in Figure 5 is V in Figures 1 and 3.
FIG. 6 is a cross-sectional view of the hydraulic drive mechanism showing the state when it is in a low compression ratio state. FIG. 7 is a cross-sectional view of the hydraulic drive mechanism showing the state when it is in a high compression ratio state. Fig. 8 is a hydraulic circuit diagram thereof, and Figs. 9 to 13 show a variable compression ratio device for an internal combustion engine as a second embodiment of the present invention.
Figure 9 is an overall configuration diagram showing the situation when the compression ratio is low, Figure 10 is an overall configuration diagram showing the situation when the compression ratio is high, and Figure 11 is the overall configuration diagram showing the situation when the compression ratio is low. Figure 12 is a cross-sectional view of the hydraulic drive mechanism showing the state in a high compression ratio state, Figure 13 is its hydraulic circuit diagram, and Figures 14 and 15 are This shows a variable compression ratio device for an internal combustion engine as a third embodiment of the invention, and FIG. 14 is an overall configuration diagram showing its main parts broken away, and FIG. 15 is a sectional view of its hydraulic drive mechanism. Fig. 16 is a diagram explaining the dimensions of each part of the crankshaft, and Fig. 17 is a characteristic diagram of the hydraulic pressure applied to the hydraulic drive mechanism and its hydraulic supply system. In each figure, the same reference numerals indicate almost similar parts. There is.

First, a first example will be described.

Now, as shown in FIGS. 1 to 4, the connecting rod 6 has its small end pivotally supported by a piston pin 7 of a piston 8 that reciprocates within the cylinder of a gasoline engine (internal combustion engine), and its large end. The end portion is pivotally supported by a crank pin 2 of a crankshaft 1.

In addition, an eccentric sleeve 5 is rotatably provided at the pivot point at the large end of the connecting rod 6 to make the bearing hole of the connecting rod 6 and the crank pin 2, which is a support shaft inserted through the bearing hole, eccentric with respect to each other. It is being

That is, the eccentric sleeve 5 is eccentric between the center of its inner circumferential circle and the center of its outer circumferential circle, and the minimum eccentric position can be obtained by rotating the outer circumference of the crank pin 2 by 180° from the maximum eccentric position.

As shown in detail in FIG. 5, a metal bearing 9 with the inner circumferential surface of the eccentric sleeve 5 is interposed between the inner circumferential surface of the eccentric sleeve 5 and the outer circumferential surface of the crank pin 2. A metal bearing 10 with an inner circumferential surface of the bearing hole of the connecting rod 6 is interposed between the outer circumferential surface of the eccentric sleeve 5 and the inner circumferential surface of the bearing hole of the connecting rod 6. Thereby, it is possible to slide between the eccentric sleeve 5 and the crank pin 2, and also between the eccentric sleeve 5 and the bearing hole of the connecting rod 6.

Incidentally, an eccentric sleeve locking means 11 is provided, and this eccentric sleeve locking means 11 is provided with a stopper pin 12 as a pin member movable in the axial direction of the eccentric sleeve 5, that is, in the axial direction of the crankshaft 1. By operating this stopper pin 12 with a hydraulic drive mechanism 11A serving as a piston-type fluid pressure drive mechanism, two engaging portions 5a and 5b formed on the eccentric sleeve 5 are activated.
The stopper pin 12 is engaged with the eccentric sleeve 5.
can be fixed at two positions (the maximum eccentricity position and the minimum eccentricity position described above).

Furthermore, this eccentric sleeve locking means 11 will be explained in detail. As shown in Figures 6 and 7, first, stopper pin 1
A piston part 12a is integrally provided in the middle part of the connecting rod 6 with an enlarged diameter in the form of a flange, and the stopper pin 12 with the piston part 12a is fitted into a through hole formed in the large end of the connecting rod 6. are combined. This through hole passes through the large end of the connecting rod 6 in the crankshaft axial direction, and is configured as a three-step hole having three diameters, and the small diameter hole located at one end is for the stopper pin 12. The diameter of the piston 12a is approximately the same as that of the piston 12a, and the medium diameter hole located in the middle is approximately the same as the diameter of the piston 12a.

Therefore, when the stopper pin 12 with the piston portion 12a is inserted into this through hole, the stopper pin 12 is inserted into the small diameter hole of the through hole.
2 is fluid-tightly inserted thereinto, and the piston portion 12a is fluid-tightly inserted into the medium-diameter hole portion of the through hole. Then, insert the return spring 15, and then fit the cap 16, which has a through hole in the center that is approximately the same diameter as the large diameter part of the through hole and approximately the same diameter as the stopper pin 12, and connect this cap 16 to the connecting rod 6. When the stopper pin 12 is fixed with a bolt or the like, one end of the stopper pin 12 is fitted into the small diameter part of the through hole in a liquid-tight manner, and the other end is fitted into the through hole of the cap 16 in a liquid-tight manner. The middle diameter part is the piston part 12
It is divided into two chambers 13.14 at a. Hydraulic passages 17.1 and 17.1 are provided in these chambers 13.14, respectively.
8 are connected in series. Thereby, these two chambers are configured as hydraulic chambers (fluid pressure chambers) 13 and 14 formed on both sides of the piston portion 12a. Note that the return spring 15 is loaded into the hydraulic chamber 13 and urges the stopper pin 12 with the piston portion 12a toward the hydraulic chamber 14 side. Note that the pressure receiving areas on both sides of the piston portion 12a are set to be equal.

As a result, the piston portion 12a attached to this stopper pin 12, the hydraulic chambers 13, 14 . Return spring 15.
The stopper pin 1 is moved by moving the piston portion 12a connected to the stopper pin 12 using the cap 16 or the like.
A piston-type hydraulic drive mechanism 11A capable of driving 2 is constructed.

Further, as shown in FIGS. 1 to 4, the eccentric sleeve 5 has flanges spaced apart in the axial direction so as to sandwich the large end of the connecting rod 6. A notch-like engagement portion 5a is formed in the portion where the eccentricity is at the minimum position, and a notch-like engagement portion is formed at the portion where the eccentric sleeve 5 is at the maximum eccentricity position in the other flange portion. When the joint 5b is formed and the stopper pin 12 moves to the right as shown in Figs. 3 and 7 and assumes the first position, the stopper pin 12 moves to the right as shown in Figs. 12 and the engaging portion 5b, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 12 is moved to the left as shown in FIGS. In the state where it has moved to the second position, the stopper pin 12 and the engaging portion 5a engage with each other, as shown in FIGS. 1 and 2, and the eccentric sleeve 5 moves the connecting rod 6 at the minimum eccentric position. It is designed to be fixed at the big end.

When the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position, the connecting rod 6 appears to be in the most extended state, and a high compression ratio state can be achieved. When the eccentric sleeve 5 is fixed to the big end of the connecting rod 6 in the minimum eccentric position,
The connecting rod 6 appears to be in the most compressed state, thereby realizing a low compression ratio state.

Note that the compression ratio in this low compression ratio state is selected to a value that does not cause the engine to knock, and this value is approximately the same as the value set in a normal engine. Therefore, the compression ratio in the high compression ratio state is set to a higher value than the value set for a normal engine.

Furthermore, means for previously applying a required hydraulic pressure (standard hydraulic pressure) to both hydraulic chambers 13 and 14 through hydraulic passages 17 and 18, and a means for applying a required hydraulic pressure (standard hydraulic pressure) to both hydraulic chambers 13 and 14 in advance, and a means for applying a required hydraulic pressure (standard hydraulic pressure) to both hydraulic chambers 13 and 14 in advance, and a stopper pin 12 with a piston portion 12a are provided to apply hydraulic pressure against the biasing force of the return spring 15. In order to move the oil to the chamber 13 side, a hydraulic pressure higher than the standard oil pressure (
Means for applying standard oil pressure +α) is provided.

That is, as shown in FIGS. 1 and 3, the hydraulic passages 17, 18 extend from the crank journal 3 of the crankshaft 1, through the crank arm 4, and from the crank pin 2 to the metal bearing 9. Eccentric sleeve5. The metal bearing 10 and the large end of the connecting rod 6 are connected to hydraulic chambers 13 and 14, respectively.

Since the metal bearing 9 and the crank pin 2 and the metal bearing 10 and the eccentric sleeve 5 slide, as shown in FIG.
Two endless grooves are formed around this inner peripheral surface and connected to hydraulic passages 17 and 18, respectively, and the hydraulic passages 17 and 17 formed in the eccentric sleeve 5 are located in the grooves formed in the metal bearing 9.
．． A through hole is formed that matches the section 18, and
A through hole is also formed in the groove formed in the metal bearing 10 and is aligned with the hydraulic passage 17 and 18 formed in the large end of the connecting rod 6.

Further, as shown in FIG. 8, the part of the hydraulic passage 17 outside the crankshaft is connected to the main gear rally 23 side, and the part of the hydraulic passage 18 outside the crankshaft is connected to the sub oil pump 24 or the main gear rally 23 side. There is. That is, oil (lubricating oil) from the oil tank or oil pan 20 is supplied to the main gear rally 23 via the oil filter 22 as oil at the required oil pressure (hydraulic pressure that supplies standard oil pressure) by the oil pump 19 with the relief valve 21. , this main gear rally 2
3 supplies standard oil pressure oil through a hydraulic passage 17. Furthermore, the oil from the main gear rally 23 is supplied to the sub-oil pump 24 to further increase the oil pressure (II!
The hydraulic pressure from the sub oil pump 24 is supplied to the hydraulic passage 18 selectively with the hydraulic pressure from the main gear rally 23 via a switching valve 25. ing. That is, when the switching valve 25 is placed in position a as shown in FIG. A high hydraulic pressure (iR semi-hydraulic pressure +α) is supplied from the pump 24.

Therefore, when the switching valve 25 is placed in the b position, the high oil pressure (standard oil pressure + α) is supplied from the sub-oil pump 24 to the hydraulic passage 18, and this high oil pressure is supplied to the hydraulic chamber 14. At this time, the standard hydraulic pressure from the main gear rally 23 is supplied into the hydraulic chamber 13 through the hydraulic passage 17, so the stopper pin 12 with the piston portion 12a moves against the biasing force of the return spring 15.
When it moves to the right as shown in Fig. 7 and assumes the first position, the stopper pin 12 and the engaging portion 5b engage with each other as shown in Figs. 3 and 4, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position. As a result, the connecting rod 6 appears to be in the most extended state, and a high compression ratio state can be achieved.

Further, when the switching valve 25 is set to the a position, the standard hydraulic pressure is supplied from the main gear rally 23 to the hydraulic passage 18, and this standard hydraulic pressure is also supplied to the hydraulic chamber 14. At this time, standard hydraulic pressure is supplied from the main gear rally 23 into the hydraulic chamber 13 via the hydraulic passage 17, so the biasing force of the return spring 15 causes the piston portion 12 to
When the stopper pin 12 with a moves to the left as shown in FIGS. 2 and 6 and assumes the second position, the stopper pin 12 and the engaging portion 5a engage as shown in FIGS. is engaged,
An eccentric sleeve 5 is fixed to the big end of the connecting rod 6 in a minimum eccentric position. As a result, the connecting rod 6 appears to be in the most compressed state, and a low compression ratio state can be realized.

In addition, oil pump 19. Both sub oil pumps 24 are driven by the engine.

Further, reference numeral 26 in FIG. 8 is a relief valve, and this relief valve 26 is used to adjust the differential pressure α between (standard oil pressure + α) and the standard oil pressure to be constant.

Furthermore, 27 is an oil control valve for switching control of the switching valve 25, and when this oil control valve 27 is set to the a position, the switching valve 25
The pilot oil pressure of the oil control valve 2 can be lowered and the switching valve 25 can be moved to the a position.
7 to the b position, the pilot oil pressure of the switching valve 25 increases and the switching valve 25 can be moved to the b position.

A switching control signal from a controller 40 is input to this oil control valve 27.
1 and the engine rotation speed sensor 42, and when it detects an engine high load range or engine high rotation range that is larger than the engine medium load range, it outputs a control signal to set the oil control valve 27 to position a. When detecting a region where the engine load is medium or lower, a control signal is issued to set the oil control valve 27 to position b.

With the above-mentioned configuration, when a region below the engine load is detected, a control signal is issued to set the oil control valve 27 to the b position, so the switching valve 25 is also set to the b position.
position, high hydraulic pressure from the oil pump 24 is supplied to the hydraulic passage 18, and this high hydraulic pressure (
Standard oil pressure + α) is supplied. At this time, the standard oil pressure from the main gear rally 23 is supplied into the oil pressure chamber 13 via the oil pressure passage 17, so as a result, the differential pressure α between the above-mentioned high oil pressure (It semi-hydraulic pressure + α) and the standard oil pressure is applied to the piston. As a result, this differential pressure α moves the stopper pin 12 with the piston portion 12a to the right, as shown in FIGS. 3 and 7, against the biasing force of the return spring 15. As a result, the stopper pin 12 assumes the first position, and as shown in FIGS. 3 and 4, the stopper pin 12 and the engaging portion 5b engage with each other, and the eccentric sleeve 5 is placed in the maximum eccentric position with the connecting rod 6. is fixed to the big end of the As a result, the connecting rod 6 appears to be in the most extended state, resulting in a high compression ratio state. Setting the compression ratio to such a high state improves thermal efficiency and can be expected to improve fuel efficiency.

Furthermore, when a high engine load range or a high engine speed range that is larger than the engine medium load range is detected, a control signal is issued to set the oil control valve 27 to the a position, so the switching valve 25 also moves to the a position, and the oil pressure is Standard oil pressure is supplied to both the passages 17 and 1.8 from the main gear rally 23, and the standard oil pressure is supplied to the hydraulic chambers 13.14.

As a result, the stopper pin 12 with the piston portion 12a is moved by the urging force of the return spring 15.
If you move to the left and assume the second position as shown in the figure,
As shown in FIGS. 1 and 2, the stopper pin 12 and the engaging portion 5
a are engaged with each other, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. the result,
The connecting rod 6 appears to be in the most compressed state and is in a low compression ratio state. By creating a low compression ratio state in this way, knocking can be reliably avoided.

In this way, in this first embodiment, the control oil pressure of the standard oil pressure is applied to both oil pressure chambers 13 and 14 as a base, so that changes in the engine speed and crank angle can cause
The pressure difference between the two hydraulic chambers does not change, and as a result, centrifugal force, acceleration of the reciprocating movement of the connecting rod 6, etc.
The operation of the stopper pin 12 is not uncertain.

Further, since the stopper pin 12 is configured to be movable in the axial direction of the crankshaft 1, the stopper pin 12 can be operated reliably even under the influence of inertia generated due to reciprocating movement of the connecting rod, etc. I can do it.

Further, when the stopper pin 12 assumes the first position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 12 moves in the opposite direction to the second position. Once in this position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the minimum eccentric position, so there is no need to consider the drive timing of the stopper pin 12, and this allows control for changing the compression ratio. can be simplified.

On the other hand, in the device described in Japanese Patent Publication No. 63-32972, it is necessary to detect the piston position and control the lock pin protrusion timing, making the control complicated.

As mentioned above, the control oil pressure supplied to the hydraulic chambers 13 and 14 is supplied through the path from the cylinder block journal to the crank pin 2 of the crankshaft 1 to the connecting rod 6, so this control oil pressure is caused by the centrifugal force of the crankshaft. It fluctuates greatly due to the centrifugal force of the connecting rod.

That is, as shown in FIG. 16, the crankshaft angular velocity is ω, the crank radius is R, and the crank journal radius is r.
j, the crank pin radius is rp, the oil path length from the crank pin center to the stopper pin center is Q, and the oil pressure of the journal portion is PA, then the oil pressure PPC at the crank pin center at an angular velocity ω is as follows.

PPc=PA+(1/A)IΔmrli2=PA+(1
/2) u2(R"-rj2)" (1) Here, A is the cross-sectional area of the oil passage, Δm is the mass of oil over a minute distance, and ρ is the density of oil.

Furthermore, the connecting rod supply oil pressure PCR is as follows.

PcR=Ppc” (1/A)/Δmri2”PA”(
1/2) pi”(R2-rj2)+Pi2NR・rp・
cosgt...(2) Similarly, connecting rod/stopper pin hydraulic pressure P
SP becomes as follows.

Psp"PA"(1/2)-&2(R2-rj2)+-
112R(rp+1)cosht”PA+lm2((1
/2)(R2-rj”)+R(rp+1)coshlJ
...(3) Here, PB=(1/2)u2(R"-rj")...(4) PC
= pi"R(rp+1)cosi+t...1(5), pA is a function of engine speed and oil viscosity coefficient, and PB is a function of engine speed and crank phase, so The control oil pressure of the stopper pin varies greatly as shown in Fig. 17. In Fig. 17, the connecting rod supply oil pressure PCR is equal to the crank pin part oil pressure (max, m1n).
(Characteristics B and C), and the connecting rod/stopper pin hydraulic pressure PSP is the stopper pin hydraulic pressure (max.

m1n) (characteristic, E).

However, as in this embodiment, if the pressure difference α is made to act between the pressure chambers 13 and 14 on both sides, the pressure that fluctuates due to the crankshaft centrifugal force and the connecting rod centrifugal force is canceled out. Therefore, it is possible to form a hydraulic pressure supply system that is not affected by engine speed or crank phase.

Next, a second embodiment will be described using FIGS. 9 to 13.

Now, in this second embodiment, as shown in FIGS. 9 and 10,
An eccentric sleeve 5 is attached to a pivot portion at the large end of the connecting rod 6 to make the bearing hole of the connecting rod 6 and the crank pin 2, which serves as a support shaft inserted through the bearing hole, eccentric from each other. It is rotatable so that it can be adjusted.

Further, an eccentric sleeve locking means 11 is also provided,
This eccentric sleeve locking means 11 also operates a stopper pin 12 as a pin member movable in the axial direction of the eccentric sleeve 5, that is, in the axial direction of the crankshaft 1, by a hydraulic drive mechanism 11A as a piston type fluid pressure parking mechanism. By this, two engaging portions 5a are formed on the eccentric sleeve 5.

5b is engaged with the stopper pin 12 to fix the rotation of the eccentric sleeve 5 at two positions (the maximum eccentricity position and the minimum eccentricity position described above).

In this embodiment as well, the eccentric sleeve 5 has flanges spaced apart in the axial direction so as to sandwich the large end of the connecting rod 6, and the eccentric sleeve 5 at the - side flange is at the minimum eccentric position. A notch-like engaging portion 5a is formed in the portion where the eccentric sleeve 5 takes the maximum eccentricity position, and a notch-like engaging portion 5b is formed in the portion where the eccentric sleeve 5 takes the maximum eccentric position in the other flange portion. is formed, and when the stopper pin 12 moves to the right and assumes the first position as shown in FIG. 10.12, as shown in FIG. 5b, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 1
2 moves to the left and assumes the second position as shown in FIGS. 9 and 11, the stopper pin 12 and the engaging portion 5a engage with each other as shown in FIG. The eccentric sleeve 5 is adapted to be fixed to the large end of the connecting rod 6 in the minimum eccentric position, and furthermore, when the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 in the maximum eccentric position,
The connecting rod 6 appears to be in the most extended state to achieve a high compression ratio state, and when the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position, the connecting rod No. 6 appears to be in the most compressed state, making it possible to realize a low compression ratio state.

By the way, in this second embodiment, the structure of the eccentric sleeve locking means 11 is different from that of the above-mentioned second embodiment.

That is, the stopper pin 12 with the piston portion 12a is
By fitting into the through hole formed at the large end of the connecting rod 6 and fixing the cap 16 to the connecting rod 6 with bolts or the like, the medium diameter portion of the through hole is connected to the piston portion 12a to form two chambers. 13.14, which is configured as a hydraulic chamber, each chamber 13.1
The fact that the hydraulic passages 17 and 18 are connected to the hydraulic passages 17 and 18 is the same as in the first embodiment described above, but in this second embodiment,
None of the hydraulic chambers 13, 14 are loaded with return springs.

Instead, the differential pressure α (α′
<α) occurs and the required oil pressure (standard oil pressure) is maintained even on the low pressure side.
Hydraulic pressure is applied to both side pressure chambers 13 and 14 so as to have
') and a state where the oil pressure in the other oil pressure chamber 14 is higher (standard oil pressure + α').

Such means will be explained in further detail.

That is, as shown in FIGS. 9 and 10, the hydraulic passages 17, 18 extend from the crank journal 3 of the crankshaft 1, through the crank arm 4, and from the crank pin 2 to the metal bearing 9. Eccentric sleeve5. The metal bearing 10 and the large end of the connecting rod 6 are connected to hydraulic chambers 13 and 14, respectively, but as shown in FIG. 13, the parts of the hydraulic passages 17 and 18 outside the crankshaft are It is connected to the oil pump 24 or main gear rally 23 side. That is, oil tank 2o
The oil (lubricating oil) is supplied to the main gear rally 23 via an oil filter 22 as oil at the required oil pressure (standard oil pressure) by an oil pump 19 with a relief valve 21.
The oil from this main gear rally 23 is supplied to
The oil pressure from the sub oil pump 24 is supplied to the sub oil pump 24 and discharged as a higher oil pressure (standard oil pressure + α'), but the oil pressure from the sub oil pump 24 is transferred to the oil pressure from the main gear rally 23 via a switching valve 25'. and are selectively supplied to hydraulic passages 17 and 18. That is, when the switching valve 25' is set to the a position as shown in FIG. (li
l! When the switching valve 25 is set to the b position, the standard hydraulic pressure from the main gear rally 23 is supplied to the hydraulic passage 18, and the high hydraulic pressure from the sub oil pump 24 is supplied to the hydraulic passage 17. Hydraulic (S
Semi-hydraulic pressure +α') is supplied. As a result, there is a constant pressure difference α between both wave pressure chambers 13 and 14.
′ has occurred.

Therefore, when the switching valve 25' is set to the a position, the standard hydraulic pressure from the main gear rally 23 is supplied to the hydraulic passage 17, and the high hydraulic pressure (#A semi-hydraulic pressure + α') from the sub oil pump 24 is supplied to the hydraulic passage 18. ) is supplied to the hydraulic chamber 13, standard hydraulic pressure is supplied to the hydraulic chamber 13, and the hydraulic chamber 1
4 is supplied with (standard oil pressure +α'). This results in
Stopper pin 12 with piston part 12a is No. 10.12
As shown in the figure, it moves to the right and assumes the first position, and as shown in FIG. It is fixed to the large end of the connecting rod 6. As a result, the connecting rod 6 appears to be in a state where the amount of soil is expanded, and a high compression ratio state can be achieved.

Furthermore, when the switching valve 25' is set to the b position, the standard oil pressure from the main gear rally 23 is supplied to the oil pressure passage 18, and the high oil pressure (standard oil pressure + α') from the sub oil pump 24 is supplied to the oil pressure passage 17. Since the stopper pin 12 with the piston portion 128 is supplied, the stopper pin 12 with the piston portion 128
Move to the left and assume the second position as shown in Figure 1.
As shown in FIG. 9, the stopper pin 12 and the engaging portion 5a are engaged, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position, so that the connecting rod 6 is The apparent volume of soil also shrinks, making it possible to achieve a low compression ratio state.

It is assumed that the pressure-receiving areas on both sides of the piston portion 12a are set equal, and the sliding friction of the piston portion 12a is negligibly small.

Further, reference numeral 26' in FIG. 13 is a relief valve, and this relief valve 26' is for adjusting so that the differential pressure α' between (standard oil pressure + α') and the standard oil pressure is constant.

Furthermore, 27' is an oil control valve for switching control of the switching valve 25', and when this oil control valve 27' is set to the a position, the pilot oil pressure of the switching valve 25' decreases, and the switching valve 25' is set to the a position. When the oil control valve 27' is placed in the b position, the pilot oil pressure of the switching valve 25' increases, allowing the switching valve 25' to be placed in the b position.

A switching control signal from a controller 40 is input to this oil control valve 27', but the controller 40 receives detection signals from an engine load sensor 41 and an engine rotation speed sensor 42, and When a high engine load range or a high engine speed range that is larger than the engine medium load range is detected, a control signal is issued to set the oil control valve 27' to position b, and when an area below the engine medium load is detected, the oil control valve 27' is A control signal is issued to place 27' at position a.

With the above-described configuration, when a region of engine load or below is detected, a control signal is issued to set the oil control valve 27' to position a, so that the switching valve 25'
is also in position a, and the main gear rally 2 is connected to the hydraulic passage 17.
3 is supplied with standard hydraulic pressure, and the hydraulic pressure passage 18
Since the high oil pressure (IN semi-hydraulic pressure + α') is supplied from the sub-oil pump 24 to A differential pressure α' with respect to the standard oil pressure is applied to the piston portion 12a, and this differential pressure α' is applied to the piston portion 12a.
10. Move the stopper pin 12 to the right as shown in FIG. 10 and 12 to assume the first position. As a result, the 10th
As shown in the figure, the stopper pin 12 and the engaging portion 5b engage with each other, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position. As a result, the connecting rod 6 appears to be in its most extended state,
It becomes a high compression ratio state. By creating a high pressure and compression ratio state in this way, thermal efficiency improves, and improvements in fuel efficiency can be expected.

Furthermore, when a high engine load range or a high engine speed range that is larger than the engine medium load range is detected, a control signal is issued to set the oil control valve 27' to position b, so that the switching valve 25' is also set to position b. ,
The standard oil pressure from the main gear rally 23 is supplied to the hydraulic passage 18, and the high oil pressure (standard oil pressure + α') from the sub-oil pump 24 is supplied to the hydraulic passage 17. Since the oil pressure is higher, the stopper pin 12 with the piston portion 12a is
As shown in FIG. 11, move to the left and take the second position, and as shown in FIG.
are engaged with each other, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. As a result, the connecting rod 6 appears to be in the most compressed state, resulting in a low compression ratio state. By creating a low compression ratio state in this way, knocking can be reliably avoided.

In this way, in this second embodiment, a differential pressure α' is generated between the two side pressure chambers 13 and 14, and the hydraulic pressure is applied to both side pressure chambers 13 and 14 so that the required oil pressure (standard oil pressure) is maintained even on the low pressure side. A means that can switch between a state in which the hydraulic pressure in one hydraulic chamber 13 is higher (1'! semi-hydraulic pressure + α') and a state in which the hydraulic pressure in the other hydraulic chamber 14 is higher (standard hydraulic pressure + α'). Therefore, the pressure difference between the two wave pressure chambers does not change due to changes in the engine speed or crank angle, and as a result, the stopper pin 12 is In addition to not causing uncertain operation,
The differential pressure α' between both wave pressure chambers 13 and 14 is smaller than the differential pressure α in the first embodiment described above (this differential pressure α needs to be large enough to overcome the biasing force of the return spring 15). (Theoretically, it should have any positive value), so 1.For example, sub oil pump 2.
The stopper pin 12 can be reliably driven even when the discharge capacity of the stopper pin 4 is low.

Further, the stopper pin 12 is configured to be movable in the axial direction of the crankshaft 1.

The stopper pin 12 can be operated reliably even under the influence of inertia generated due to reciprocating movement of the connecting rod or the like.

Further, when the stopper pin 12 assumes the first position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 12 moves in the opposite direction to the second position. Once in this position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the minimum eccentric position, so there is no need to consider the drive timing of the stopper pin 12, and this allows control for changing the compression ratio. can be simplified.

Next, a third embodiment will be described using FIGS. 14 and 15.

Now, in this third embodiment, as shown in FIG. 14, a piston pin 7 as a support shaft inserted into a bearing hole of the connecting rod 6 and the bearing hole is attached to the pivot portion at the small end of the connecting rod 6. An eccentric sleeve 28 is rotatably provided so that it can take a maximum eccentric position and a minimum eccentric position.

Further, an eccentric sleeve locking means 29 is provided,
This eccentric sleeve locking means 29
The two engaging portions formed on the eccentric sleeve 28 are actuated by the hydraulic drive mechanism 35 as a piston-type fluid pressure drive mechanism to actuate the stopper pin 30 as a pin member that can move in a direction perpendicular to the axial direction of the By engaging the stopper pin 30 with 28a and 28b, this eccentric sleeve 28
It is possible to fix the rotation at two positions (the above-mentioned maximum eccentricity position and minimum eccentricity position).

Note that the eccentric sleeve 28 has a notch-shaped engagement portion 28a formed at a portion of its outer periphery where the eccentric sleeve 28 takes the minimum eccentric position, and a notch-like engagement portion 28a is formed at the outer periphery where the eccentric sleeve 28 takes the maximum eccentric position. A notch-shaped engaging portion 28b is formed in the portion where the stopper pin 30 is positioned, and when the stopper pin 30 protrudes, the stopper pin 30 and the engaging portion 28a or 28b engage with each other. .

Then, this stopper pin 30 and the engaging portion 28a or 2
8b to be engaged with is determined by the following means. That is, as disclosed in Japanese Patent Publication No. 63-32972, by counting the number of teeth of the ring gear that rotates with the crankshaft 1 using an electromagnetic pickup, the crank position and eventually the piston position are detected, and the piston position is detected. The timing is determined to engage either the engaging portion 28a or 28b.

As a result, when the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 at the maximum eccentric position or the minimum eccentric position, the connecting rod 6 appears to be in an expanded state or a most contracted state, This makes it possible to achieve a high compression ratio state or a low compression state.

Further, the stopper pin 30 has a piston portion 3 at its intermediate portion.
0a (the pressure receiving areas on both sides of this piston part 30a are also set equal), but
The stopper pin 30 with the piston part 30a opens in the inner peripheral surface of the bearing hole at the small end of the connecting rod 6 and is inserted into a recess having approximately the same diameter as the piston part 30a.Then, the return spring 33 is loaded, and then the stopper pin 30 is capped. 34, hydraulic chambers (fluid pressure chambers) 31 and 32 are formed on both sides of the piston portion 30a. The hydraulic passages 1 are connected to these hydraulic chambers 31 and 32, respectively.
7.18 is connected. As shown in FIG. 14, this hydraulic passage 17, 18 is connected to a connecting rod 6, a metal bearing 9. It passes through the crankshaft 1, and the portion outside the crankshaft 1 is connected to a hydraulic circuit system as shown in FIG.

Furthermore, means for applying a required hydraulic pressure (1F! semi-hydraulic pressure) to both wave pressure chambers 31 and 32 through the hydraulic passages 17 and 18, and a stopper pin 30 with a piston portion 30a that resists the biasing force of the return spring 33 are provided. In order to move the hydraulic pressure toward the hydraulic chamber 31, a means for applying a hydraulic pressure higher than the standard hydraulic pressure (standard hydraulic pressure +α) to the hydraulic chamber 32 is provided. Then, when the above-mentioned high oil pressure acts on the hydraulic chamber 32,
The stopper pin 30 protrudes and engages the engaging portion 28a or 28.
b.

Furthermore, a switching control signal from a controller 40 is input to the oil control valve 27 shown in FIG. Upon receiving detection signals from the engine rotation speed sensor 42 and the electromagnetic pickup (piston position sensor), a control signal is output to move the oil control valve 27 to position b when the lock is applied. When changing the joint, a control signal is issued to move the oil control valve 27 to position a. Note that the timing of outputting the control signal is determined based on the signal from the piston position sensor.

With the above-described configuration, when a region of engine load or below is detected, a control signal is issued to set the oil control valve 27 to the a position, and the switching valve 25 is set to the a position. Then, the standard hydraulic pressure from the main gear rally 23 is supplied to both the hydraulic passages 17 and 18, and the hydraulic pressure chamber 13
．． 14 is supplied with standard oil pressure. As a result, the stopper pin 30 with the piston portion 30a is retracted by the biasing force of the return spring 15, and the engagement between the stopper pin 30 and the engaging portion is disengaged. After the locked state is released in this manner, a control signal is issued at a required timing to move the oil control valve 27 to position b. As a result, the switching valve 25 is also in the b position, high hydraulic pressure from the oil pump 24 is supplied to the hydraulic passage 18, and this high hydraulic pressure (standard hydraulic pressure + α) is supplied to the hydraulic chamber 32. At this time, the standard oil pressure from the main gear rally 23 is supplied into the oil pressure chamber 31 via the oil pressure passage 17, so that as a result, the differential pressure α between the high oil pressure (standard oil pressure + α) and the standard oil pressure is applied to the piston. 30a
As a result, this differential pressure α resists the biasing force of the return spring 33 and causes the stopper pin 30 with the piston portion 30a to protrude. As a result, the protruding stopper pin 30 and the engaging portion 28b engage with each other in relation to the above-mentioned control signal output timing, and the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 at the maximum eccentric position. . As a result, the connecting rod 6 appears to be in the most extended state, resulting in a high compression ratio state. Setting the compression ratio to such a high state improves thermal efficiency and can be expected to improve fuel efficiency.

Also, if a high engine load range or a high engine speed range that is larger than the engine medium load range is detected, a control signal is first issued to set the oil control valve 27 to position a, and the stopper pin 30 and the engaging portion are connected. After disengaging and releasing the locked state, at another required timing, a control signal is issued to set the oil control valve 27 to position b, so that the switching valve 25 is also set to position b, and the oil pressure is increased. A high hydraulic pressure is supplied from the oil pump 24 to the passage 18, and this high hydraulic pressure (standard hydraulic pressure +α) is supplied to the hydraulic chamber 32. This causes the stopper pin 30 with the piston portion 30a to protrude against the biasing force of the return spring 33 due to the differential pressure α, as in the case described above. Therefore, in relation to the above control signal output timing, the protruding stopper pin 30 and the engaging portion 28a engage, and the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 at the minimum eccentric position. the result.

The connecting rod 6 appears to be in the most compressed state and is in a low compression ratio state. By creating a low compression ratio state in this way, knocking can be reliably avoided.

In this way, in the case of the third embodiment as well, the eccentric sleeve locking means 29 is arranged at the small end of the connecting rod 6, and the protruding operation and timing of the stopper pin 30 allows a high compression ratio state and a low compression state to be achieved. Although this is different from the first embodiment in that a ratio state is achieved, standard oil pressure is applied to both hydraulic chambers 31 and 32 as a base, so changes in the engine speed and crank angle The pressure difference between the pressure chambers on both sides does not change, so that the protruding and retracting operation of the stopper pin 30 does not become uncertain due to centrifugal force or acceleration of the reciprocating movement of the connecting rod 6.

In the case shown in FIGS. 14 and 15, the return spring 33 is removed and the hydraulic circuit system of the second embodiment described above (see FIG. 13) is used to connect both hydraulic chambers 31 and 32.
A differential pressure α'(α' α) is generated between both hydraulic chambers 31.3 so that the required hydraulic pressure (II semi-hydraulic) is maintained even on the low pressure side.
2 and enables switching between a state where the oil pressure in one hydraulic chamber 31 is higher (standard oil pressure + α') and a state where the oil pressure in the other oil pressure chamber 32 is higher (standard oil pressure + α'). You can.

However, in this case as well, the controller 40 outputs a switching control signal to the oil control valve 27' shown in FIG.
Unlike the second embodiment described above, the oil control valve 27' is moved to position a when locked in response to detection signals from the engine load sensor 41, engine speed sensor 42, and electromagnetic pickup (piston position sensor). When changing the engaging portion that engages with the stopper pin 28, a control signal that moves the oil control valve 27' to position b is issued. Note that the timing of outputting the control signal is determined based on the signal from the piston position sensor.

With the above-mentioned configuration, when a region below the engine load is detected, a control signal is first issued to set the oil control valve 27' to position b. Then, the switching valve 25' is also in position b, and the standard oil pressure from the main gear rally 23 is supplied to the oil pressure passage 18, and the high oil pressure (standard oil pressure + α') from the sub oil pump 24 is supplied to the oil pressure passage 17. is supplied, and the hydraulic pressure in the hydraulic chamber 31 becomes higher, so the stopper pin 30 with the piston portion 30a is retracted, and the engagement between the stopper pin 30 and the engaging portion is disengaged. After the locked state is released in this manner, a control signal is issued at a required timing to move the oil control valve 27' to position a.

As a result, the switching valve 25' is also in the a position, the standard oil pressure from the main gear rally 23 is supplied to the oil pressure passage 17, and the high oil pressure (standard oil pressure + α') from the sub oil pump 24 is supplied to the oil pressure passage 18. ) is supplied, and the hydraulic pressure in the hydraulic chamber 32 becomes higher.
As a result, a differential pressure α' between the high oil pressure (standard oil pressure +α') and the standard oil pressure is applied to the piston portion 30a, and this differential pressure α' causes the stopper pin 30 with the piston portion 30a to protrude. As a result, the stopper pin 30 and the engaging portion 28b engage with each other in relation to the above-mentioned control signal output timing, and the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 at the maximum eccentric position.

As a result, the connecting rod 6 appears to be in the most extended state, resulting in a high compression ratio state.

By creating a high compression ratio state like this, thermal efficiency improves,
Improvements in fuel efficiency, etc. can be expected.

Also, when detecting a high engine load range or a high engine speed range that is larger than the engine medium load range, a control signal is first issued to set the oil control valve 27' to position b, and the stopper pin 30 and the engaging part are After disengaging the oil control valve 27' from the locked state and releasing the locked state, a control signal is issued at another required timing to set the oil control valve 27' to position a, so the switching valve 25'
is also in position a, and the main gear rally 2 is connected to the hydraulic passage 17.
3 is supplied with standard hydraulic pressure, and the hydraulic pressure passage 18
The high oil pressure from the sub oil pump 24 (standard oil pressure +
α') is supplied, and the oil pressure in the hydraulic chamber 32 becomes higher, so that as a result, the differential pressure α' between the above-mentioned high oil pressure (If semi-hydraulic pressure + α') and the standard oil pressure is applied to the piston portion 30a. As a result, this differential pressure α' is applied to the piston portion 30a.
The attached stopper pin 30 is made to protrude. As a result, in relation to the above control signal output timing, the stopper pin 30
and the engaging portion 28a engage with each other, and the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 at the minimum eccentric position. As a result, the connecting rod 6 appears to be in the most compressed state, resulting in a low compression ratio state. By creating a low compression ratio state in this way, knocking can be reliably avoided.

In this way, similarly to the second embodiment described above, a differential pressure α' is generated between the two wave pressure chambers 31 and 32, and both wave pressure chambers 31 and 32 are arranged so that the required oil pressure (S quasi-hydraulic) is maintained even on the low pressure side. Pressure chamber 31.
32, and it is possible to switch between a state in which the oil pressure in one hydraulic chamber 31 is higher (is semi-hydraulic pressure + α') and a state in which the oil pressure in the other hydraulic chamber 32 is higher (standard oil pressure + α'). Since the pressure difference between the two wave pressure chambers does not change due to changes in the engine speed or crank angle, the stopper pin is In addition, the differential pressure α' between the two wave pressure chambers 31 and 32 is the same as the differential pressure α in the embodiment shown in FIG. 14.15 (this differential pressure α is (Theoretically, it can have any positive value), so it can be used, for example, when the sub-oil pump 24 is used at low engine speeds. Even when the discharge capacity is low, the stopper pin 30 can be reliably projected.

In each of the above embodiments, pressure oil was used as the drive means for the stopper pin 12.30 of the eccentric sleeve lock means 11.29, but other fluids (liquid or gas) at the required pressure may be used. You can use whatever you have.

[Effects of the Invention] As detailed above, according to the variable compression ratio device for an internal combustion engine of the present invention as set forth in claim 1, one end is pivotally supported by a piston that reciprocates within a cylinder of an internal combustion engine. A connecting rod is provided with the other end pivotally supported by the crankshaft, and one of the pivot parts at both ends of the connecting rod is provided with an eccentricity that makes the bearing hole of the connecting rod and the support shaft inserted through the bearing hole eccentric with respect to each other. A pin member in which the sleeve is rotatably provided, an eccentric sleeve locking means capable of fixing the rotation of the eccentric sleeve at a predetermined position, and the eccentric sleeve locking means capable of engaging with an engaging portion formed on the eccentric sleeve. and a piston type fluid pressure drive mechanism that can drive the pin member by moving a piston part connected to the pin member, and the piston type fluid pressure drive mechanism has fluid pressure chambers on both sides of the piston part. At the same time, one fluid pressure chamber is provided with a return spring that urges the piston portion to move toward the other fluid pressure chamber, and a required fluid pressure is applied in advance to both of the fluid pressure chambers. and means capable of applying a fluid pressure higher than the required fluid pressure to the other fluid pressure chamber in order to move the piston portion toward one fluid pressure chamber against the biasing force of the return spring. As a result, the pressure difference between the two fluid pressure chambers does not change due to changes in engine speed or crank angle, and this allows the pin member to remain secure even under the influence of centrifugal force or acceleration of the reciprocating movement of the connecting rod. It has the advantage of being able to operate.

Further, in the variable compression ratio device for an internal combustion engine according to the present invention, the piston type fluid pressure drive mechanism is provided with fluid pressure chambers on both sides of the piston portion, and a pressure difference is generated between both the fluid pressure chambers, resulting in a low pressure. Fluid pressure is applied to both fluid pressure chambers so as to have the required fluid pressure on both sides, and the fluid pressure in one fluid pressure chamber is higher and the fluid pressure in the other fluid pressure chamber is higher. The pressure difference between the pressure chambers on both sides does not change due to changes in the engine speed or crank angle, and as a result, due to centrifugal force or acceleration of the reciprocating movement of the connecting rod, etc. In addition to not causing uncertainty in the operation of the pin member, there is an advantage that the pin member can be operated reliably even when the discharge capacity of the pressurizing means for supplying high fluid pressure is low.

[Brief explanation of drawings]

Figures 1 to 8 show a variable compression ratio device for an internal combustion engine as a first embodiment of the present invention. Figure 1 is an overall configuration diagram showing the situation when it is in a low compression ratio state, and Figure 2 is a Figure 3 is a front view of the connecting rod showing how it looks when it is in a low compression ratio state, Figure 3 is an overall configuration diagram showing how it looks when it is in a high compression ratio state, and Figure 4 shows how it looks when it is in a high compression ratio state. A front view of the connecting rod shown in Fig. 5 is v in Figs. 1 and 3.
FIG. 6 is a cross-sectional view of the hydraulic drive mechanism showing the state when it is in a low compression ratio state. FIG. 7 is a cross-sectional view of the hydraulic drive mechanism showing the state when it is in a high compression ratio state. Fig. 8 is a hydraulic circuit diagram thereof, and Figs. 9 to 13 show a variable compression ratio device for an internal combustion engine as a second embodiment of the present invention.
Figure 9 is an overall configuration diagram showing the situation when the compression ratio is low, Figure 10 is an overall configuration diagram showing the situation when the compression ratio is high, and Figure 11 is the overall configuration diagram showing the situation when the compression ratio is low. Figure 12 is a cross-sectional view of the hydraulic drive mechanism showing the state in a high compression ratio state, Figure 13 is its hydraulic circuit diagram, and Figures 14 and 15 are This shows a variable compression ratio device for an internal combustion engine as a third embodiment of the invention, and FIG. 14 is an overall configuration diagram showing its main parts broken away, and FIG. 15 is a sectional view of its hydraulic drive mechanism. FIG. 16 is a diagram illustrating the dimensions of each part of the crankshaft, and FIG. 17 is a characteristic diagram of the hydraulic pressure applied to the hydraulic drive mechanism and its hydraulic pressure supply system. 1.--crankshaft, 2.--crank pin, 3
Crank journal, 4--crank arm, 5.2
- Eccentric sleeve, 5a, 5b - Engaging portion, 6 - ml connecting rod, 7 - Piston pin, 8 - Piston, 9.10 - Metal bearing, 11 - Eccentric slide block means, 11A -. Hydraulic drive mechanism as piston-type fluid pressure drive mechanism, 12--stopper pin (pin member), 12
a- Piston part, 13.14- Hydraulic chamber (fluid pressure chamber), 1
5.-Return spring, 16.--Cap, 1
7.18・-Hydraulic passage, 19-Oil pump, 20-・
Oil tank, 21-Relief valve, 22-Oil filter, 23-Main gear rally, 24.-Sub oil pump, 25° 25'--Switching valve, 26.2
6'--Relief valve, 27.27"--Oil control valve, 28-Eccentric sleeve, 28a, 28
b-Engaging portion, 29--Eccentric sleeve locking means, 30
- Stopper pin (pin member), 30a - Piston part, 31.32 - Hydraulic chamber (fluid pressure chamber), 33 - Return spring, 34 - Cap, 35 - Piston type fluid pressure drive mechanism Hydraulic drive mechanism, 40-controller, 41-engine load sensor, 42--engine rotation speed sensor.

Claims (2)

[Claims] (1) A connecting rod is provided, one end of which is pivotally supported by a piston that reciprocates within a cylinder of an internal combustion engine, and the other end of which is pivotally supported by a crankshaft; On one side, an eccentric sleeve is rotatably provided to make the bearing hole of the connecting rod and a support shaft inserted through the bearing hole eccentric with respect to each other, and an eccentric sleeve locking means is provided to fix the rotation of the eccentric sleeve at a predetermined position. and the eccentric sleeve locking means can drive the pin member by moving a pin member that can engage with an engaging portion formed on the eccentric sleeve and a piston portion connected to the pin member. The piston type fluid pressure drive mechanism has fluid pressure chambers on both sides of the piston part, and the piston part is connected to one fluid pressure chamber and the piston part is connected to the other fluid pressure chamber. A means for applying a required fluid pressure to both of the fluid pressure chambers in advance; A variable compression ratio device for an internal combustion engine, comprising means capable of applying a fluid pressure higher than the required fluid pressure to the other fluid pressure chamber in order to move the fluid pressure toward the other fluid pressure chamber. . (2) A connecting rod is provided, one end of which is pivotally supported by a piston that reciprocates within a cylinder of an internal combustion engine, and the other end of which is pivotally supported by a crankshaft; On one side, an eccentric sleeve is rotatably provided to make the bearing hole of the connecting rod and a support shaft inserted through the bearing hole eccentric with respect to each other, and an eccentric sleeve locking means is provided to fix the rotation of the eccentric sleeve at a predetermined position. and the eccentric sleeve locking means can drive the pin member by moving a pin member that can engage with an engaging portion formed on the eccentric sleeve and a piston portion connected to the pin member. The piston-type fluid pressure drive mechanism has fluid pressure chambers on both sides of the piston part, and a pressure difference occurs between the two fluid pressure chambers, even on the low pressure side. Fluid pressure is applied to both fluid pressure chambers so as to have the required fluid pressure, and switching is possible between a state in which the fluid pressure in one fluid pressure chamber is higher and a state in which the fluid pressure in the other fluid pressure chamber is higher. A variable compression ratio device for an internal combustion engine, characterized in that it is equipped with a means capable of doing so.