JP2002285857A - Piston drive for internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To make a crank rotation angle required for the move of a piston in each stroke proper to improve sound vibration performance. SOLUTION: The middle point of a stroke in which the piston advances toward a top dead center TDC from a bottom dead center BDC is a first middle point P1, the middle point of a stroke in which the piston advances toward the bottom dead center BDC from the top dead center TDC is a second middle point P2, a crank rotation angle until the piston moves to the top dead center TDC from the first middle point P1 is θ1, a crank rotation angle until the piston moves to the second middle point P2 from the top dead center TDC is θ2, a crank rotation angle until the piston moves to the bottom dead center BDC from the second middle point P2 is θ3, and a crank rotation angle until the piston moves to the first middle point P1 from the bottom dead center BDC is θ4. (θ1+θ2) is not less than (θ3+θ4).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a piston driving apparatus for an internal combustion engine in which a piston reciprocating in a cylinder and a crankshaft are linked.

[0002]

2. Description of the Related Art In a reciprocating internal combustion engine, generally,
A single link type piston drive device in which a piston and a crankshaft are connected by one connecting rod is used.

[0003]

However, in such a single-link type device, the piston speed becomes slower in the latter half than in the first half of the expansion stroke due to the crank-to-rod connecting ratio (see FIGS. 2 to 4). (See solid line (b)). For this reason,
The combustion pressure change rate in the first half of the expansion stroke in which combustion proceeds increases, and the crank rotation angle required for the piston to pass in the first half of the expansion stroke in which the combustion pressure driving the piston is the largest is determined by the fact that the piston passes in the second half of the expansion stroke. Smaller than the required crank rotation angle. For this reason, there has been a problem that torque fluctuation of the crankshaft increases and sound vibration performance deteriorates.

[0004] A multi-link type piston driving device is known in which a piston and a crankshaft are linked by a plurality of links to change an engine compression ratio (Japanese Patent Application Laid-Open No. 2000-73804, Japanese Unexamined Patent Application Publication No. 2000-73804). 2000-5137
No. 79). In such a multi-link device, the degree of freedom in setting the crank rotation angle required for the piston to pass through each stroke is greater than in a single-link device.

One object of the present invention is to make the crank rotation angle appropriate in each stroke of the piston by using, for example, the above-mentioned double-link structure, to optimize the speed of change in combustion pressure and to reduce torque fluctuation in the crankshaft. The purpose is to improve the sound and vibration performance by suppression.

[0006]

SUMMARY OF THE INVENTION The present invention relates to a piston driving apparatus for an internal combustion engine in which a piston reciprocating in a cylinder and a crankshaft are linked. Here, the middle of the stroke of the piston from the bottom dead center to the top dead center is a first middle point, the middle of the stroke of the piston from the top dead center to the bottom dead center is the second middle point, and the piston is the first middle point. The crank rotation angle until the piston moves from the point to the top dead center is θ1, the crank rotation angle until the piston moves from the top dead center to the second intermediate point is θ2, and the piston moves from the second intermediate point to the bottom dead center. Assume that the crank rotation angle until the piston moves is θ3, and the crank rotation angle until the piston moves from the bottom dead center to the first intermediate point is θ4.

In the first invention, (θ1 + θ2) is (θ3 +
θ4) or more.

According to a second aspect, in addition to the first aspect, θ2 is equal to or larger than θ1.

In a third aspect, θ2 is equal to or larger than θ3.

[0010] In a fourth aspect based on the third aspect, θ1 is not less than θ4.

According to a fifth aspect, in addition to the third aspect, θ2 is equal to or greater than θ1.

According to a sixth aspect, in addition to the fourth aspect, θ2 is equal to or larger than θ1.

According to a seventh aspect, in addition to the sixth aspect, θ1 is not less than θ3.

An eighth invention provides a lower link rotatably attached to a crankpin of the crankshaft,
One end is rotatably connected to the lower link, and the other end is rotatably connected to the piston. The other end is rotatably connected to the lower link, and the other end is a support fulcrum on the engine body side. And a control link that is connected rotatably around the center and that controls the degree of freedom of the lower link, and a support position change mechanism that controls the position of the support fulcrum.

In a ninth aspect, in addition to the eighth aspect, (θ1
+ Θ2) is larger at high compression ratios than at low compression ratios.

According to a tenth aspect, in addition to the eighth aspect, (θ
2 + θ3) becomes larger at a high compression ratio than at a low compression ratio.

According to an eleventh aspect, in addition to the eighth aspect, θ2
Is larger at high compression ratios than at low compression ratios.

According to a twelfth aspect, in addition to the eleventh aspect, θ
4 is smaller at a high compression ratio than at a low compression ratio.

[0019]

According to the first aspect of the invention, the crank rotation angle (θ1 + θ2) from the latter half of the compression stroke to the first half of the expansion stroke becomes relatively large, and the time for the piston to pass around the top dead center becomes longer. A sufficient time for the air-fuel mixture to convect before starting combustion is obtained, and the combustion state is improved. Also, since the piston speed after the start of combustion becomes relatively small, the piston position after the end of combustion becomes closer to the top dead center side, and the expansion period in which the combustion gas works substantially becomes longer,
Thermal efficiency is improved. In particular, when a high compression ratio type internal combustion engine aims to improve the thermal efficiency by increasing the expansion ratio, the effect becomes large.

According to the third aspect, since the piston speed in the first half of the expansion stroke is equal to or lower than the piston speed in the second half of the expansion stroke, the combustion pressure change rate in the first half of the expansion stroke in which a rapid pressure change occurs at high pressure is reduced. And the sound and vibration performance is improved. Further, the increase in torque fluctuation of the crankshaft due to the rapid increase of the combustion pressure in the first half of the expansion stroke can be reduced by increasing the crank rotation angle θ2 in the first half of the expansion stroke, and the sound and vibration performance is further improved.

Further, when the piston speed in the first half of the intake stroke is lower than the piston speed in the second half of the intake stroke, the negative pressure at the start of intake is reduced, and pump loss can be reduced. Further, since the amount of intake air immediately after the start of intake is reduced, the lift amount of the intake valve can be reduced. Further, since the piston speed in the latter half of the intake stroke increases, the intake inertia effect increases before and after the bottom dead center, so that more air-fuel mixture can be charged into the cylinder, and the output can be improved.

The operation and effect of the other invention will be apparent from the description of the embodiment described later.

[0023]

Preferred embodiments of the present invention will be described below with reference to the drawings. FIG. 1 is a schematic diagram showing a piston drive device for an internal combustion engine according to this embodiment. A lower link 2 is rotatably attached to a crankpin 11 of the crankshaft 1. One end of the upper link 5 is connected to the first connection pin 10
The other end of the upper link 5 is rotatably connected to the piston 3 via a piston pin 4. Although not shown here, the piston 3 is disposed reciprocally in the cylinder. A combustion chamber is defined above the piston 3 and an intake valve and an exhaust valve (hereinafter, referred to as a piston). An intake / exhaust valve or simply a valve is provided as necessary).

The lower link 2 includes a control link 6
One end of the control link 6 is rotatably connected via a second connection pin 9, and the other end of the control link 6 is rotatably supported by a cylinder block (not shown) or the like on the engine body side around a support fulcrum 12. Have been. This support fulcrum 12
The support position changing mechanism 13 for changing and controlling the position of the control shaft 7 includes a control shaft 7 rotatably supported by the engine body,
And an actuator (not shown) such as a motor that rotationally drives the control shaft 7. An eccentric cam 8 is eccentrically provided on the control shaft 7, and the other end of the eccentric cam 8 is rotatably attached to the control link 6. Therefore,
The center of the eccentric cam 8 is the support fulcrum 12 described above.

With this configuration, when the control shaft 7 is rotationally driven in accordance with the engine operating condition, the motion constraint condition of the lower link 2 by the control link 6 changes, the piston stroke changes, and the engine compression ratio changes. Is done.

The piston stroke characteristics according to this embodiment are shown in the solid line (a) of FIGS. In addition,
The solid line (b) in FIGS. 2 to 4 shows the piston stroke characteristics of the single link type piston drive device according to the comparative example. Solid lines (c) and (d) in FIGS. 5 to 7 show the piston stroke characteristics of the present embodiment in the setting state of the high compression ratio and the setting state of the low compression ratio, respectively.

Next, the definition of each code will be described with reference to FIG. FIG. 8 is not a diagram showing the piston stroke characteristics of the present embodiment.

The piston moves from the bottom dead center BDC to the top dead center TDC
The middle of the travel (length) toward the first intermediate point P1, and the middle of the travel of the piston from the top dead center TDC to the bottom dead center BDC is the second intermediate point P2. Further, a stroke from the first intermediate point P1 to the top dead center TDC is assumed, a stroke from the top dead center TDC to the second intermediate point P2 is assumed, and a stroke from the second intermediate point P2 to the bottom dead center BDC is assumed. The process from the bottom dead center BDC to the first intermediate point P1 is assumed.

Further, the crank rotation angles required for the piston to move through each stroke,.
θ2, θ3, θ4. That is, the piston
The crank rotation angle from the intermediate point P1 to the movement to the top dead center TDC is θ1, and the crankshaft rotation angle from the top dead center TDC to the second intermediate point P2
The crank rotation angle from the second middle point P2 to the bottom dead center BDC is θ3, and the crank rotation angle from the bottom dead center BDC to the first middle point P1 is θ2. The rotation angle is assumed to be θ4.

Next, with reference to FIGS. 9 and 10, the features of the present embodiment and the effects thereof will be listed. 9 and 10 exaggerate the respective features and do not accurately represent the piston stroke characteristics of the present embodiment shown in FIGS. However, this embodiment satisfies all the requirements (1) to (12) described below. (1). θ1 + θ2 is equal to or larger than θ3 + θ4.

That is, the crank rotation angle (θ1 + θ2) from the latter half of the compression or exhaust stroke (stroke) to the first half of the expansion or intake stroke (stroke) varies from the latter half of the expansion or intake stroke (stroke) to the first half of the compression or exhaust stroke (stroke). Is larger than the crank rotation angle (θ3 + θ4).

As a result, the crank rotation angle (θ1 + θ2) from the latter half of the compression stroke to the first half of the expansion stroke becomes relatively large, and the time required for the piston to pass before and after the compression top dead center becomes longer. The time for convection is sufficiently obtained before burning, and the combustion state is improved. In addition, the piston speed after the start of combustion is relatively low, so the piston position after the end of combustion is closer to the top dead center, and the expansion period during which combustion gas works substantially increases, thus improving thermal efficiency. I do. In particular, when a high compression ratio type internal combustion engine aims to improve the thermal efficiency by increasing the expansion ratio, the effect becomes large. (2). (1) and θ2 is equal to or larger than θ1.

That is, the crank rotation angle θ1 + θ2 from the latter half of the compression or exhaust stroke (stroke) to the first half of the expansion or intake stroke (stroke) is the crank angle from the latter half of the expansion or intake stroke (stroke) to the first half of the compression or exhaust stroke (stroke). The rotation angle θ3 + θ4 or more, and the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke is equal to the crank rotation angle θ in the second half (stroke) of the compression or exhaust stroke.
1 or more.

As described above, by relatively increasing the crank rotation angle in the first half (stroke) of the expansion stroke, the rate of change of the combustion pressure in the first half (stroke) of the expansion stroke can be relatively reduced, and the noise can be reduced. The vibration performance is improved. Also, the piston speed in the first half (stroke) of the expansion stroke becomes relatively slow,
This decrease in speed cancels out the increase in torque fluctuation of the crankshaft due to the rapid increase in combustion pressure in the first half (stroke) of the expansion stroke, so that the sound and vibration performance is further improved.

The latter half of the compression or exhaust stroke (stroke)
Is smaller than or equal to the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke, but θ1 + θ2 is larger than θ3 + θ4, so that the crank rotation angle θ1 is set to be sufficiently larger than, for example, a conventional single-link structure. Can be done without becoming too small. Rather, when θ1 becomes relatively small, the crank rotation angle (θ4 + θ1) in the compression stroke is shortened, and the compression heat loss is reduced, so that the thermal efficiency is improved. At the same time, the crank rotation angle (θ4 + θ1) in the exhaust stroke is also shortened, but the exhaust stroke does not cause much problem because the load required for exhausting the combustion gas during the exhaust stroke is reduced due to the blowdown in the latter half of the expansion stroke. (3). θ2 is equal to or larger than θ3.

That is, the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke is equal to or larger than the crank rotation angle θ3 in the second half (stroke) of the expansion or intake stroke.

In this case, since the piston speed in the first half (stroke) of the expansion stroke is equal to or lower than the piston speed in the second half (stroke) of the expansion stroke, the combustion pressure change rate in the first half of the expansion stroke in which a rapid pressure change occurs at high pressure is reduced. Can improve sound and vibration performance. Further, an increase in torque fluctuation of the crankshaft due to a rapid increase in combustion pressure in the first half of the expansion stroke can be reduced by a decrease in piston speed due to an increase in the crank rotation angle θ2 in the first half (stroke) of the expansion stroke, and the sound vibration performance is improved. Further improve.

Further, since the piston speed in the first half (stroke) of the intake stroke is lower than the piston speed in the second half (stroke) of the intake stroke, the negative pressure at the start of intake is reduced, and pump loss can be reduced. Further, since the amount of intake air immediately after the start of intake is reduced, the lift amount of the intake valve can be reduced. Further, since the piston speed in the latter half of the intake stroke increases, the intake inertia effect increases before and after the bottom dead center, so that more air-fuel mixture can be charged into the cylinder, and the output can be improved. When θ2 is equal to or larger than θ3, the piston stroke cannot be realized with the conventional single-link structure, no matter how much the crank-to-rod connecting rod ratio is increased, and a double-link structure is used. This is made possible by: (4). (3) And θ1 is not less than θ4.

That is, the crank rotation angle θ1 in the second half (stroke) of the compression or exhaust stroke is equal to or larger than the crank rotation angle θ4 in the first half (stroke) of the compression or exhaust stroke, and the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke.
Is greater than or equal to the crank rotation angle θ3 in the latter half (stroke) of the expansion or intake stroke.

As a result, in addition to the effect of (3), the increase in the pressure change rate after the start of combustion before and after the top dead center can be further reduced, and the sound and vibration performance is improved. Also, since the exhaust speed is relatively large in the first half of the exhaust stroke where the exhaust pressure is high and the opening of the exhaust valve is large, the exhaust speed is relatively small in the second half of the exhaust stroke where the exhaust pressure is low and the opening of the exhaust valve is small. In addition, the pump loss in the exhaust stroke can be reduced. Furthermore, since the intake speed is relatively small in the first half of the intake stroke in which the opening of the intake valve is small, and the intake speed is relatively large in the second half of the intake stroke in which the opening of the intake valve is large, the pump loss in the intake stroke is reduced. can do. (5). (3) And θ2 is not less than θ1.

That is, the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke is equal to or larger than the crank rotation angle θ3 in the second half (stroke) of the expansion or intake stroke, and the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke.
Is greater than or equal to the crank rotation angle θ1 in the latter half (stroke) of the compression or exhaust stroke.

In this case, in addition to the effect of (3), θ2 is changed to θ
By setting the ratio to 1 or more, the pressure change rate in the first half (stroke) of the expansion stroke in which the pressure change rate is maximum can be further reduced, and the sound vibration performance is further improved. (6). (4) And θ2 is not less than θ1.

That is, the crank rotation angle θ1 in the second half (stroke) of the compression or exhaust stroke is equal to or larger than the crank rotation angle θ4 in the first half (stroke) of the compression or exhaust stroke, and the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke.
Is greater than or equal to the crank rotation angle θ3 in the latter half (stroke) of the expansion or intake stroke, and the crank rotation angle θ2 in the latter half (stroke) of the expansion or intake stroke is greater than or equal to the crank rotation angle θ1 in the latter half (stroke) of the compression or exhaust stroke. .

In this case, in addition to the effect of (4), the crank rotation angle θ2 in the first half (stroke) of the expansion stroke in which the rate of pressure change is maximum becomes the largest among θ1 to θ4, and the rate of pressure change is most effective. The sound and vibration performance is further improved. (7). (6) And θ1 is not less than θ3.

That is, the crank rotation angle θ2 in the first half (stroke) of the expansion or intake stroke is the largest, and the crank rotation angle θ1 in the second half (stroke) of the compression or exhaust stroke is θ2.
In the following, the crank rotation angle θ4 in the first half (stroke) of the compression or exhaust stroke is at most θ1 and the crank rotation angle θ3 in the second half (stroke) of the expansion or intake stroke is at most θ4 or less.

Therefore, θ2 is the largest, θ1 is the second largest after θ2, and the first half of the expansion stroke (stroke) at which the rate of pressure change is the largest, and the first half of the compression stroke (stroke) at which the rate of pressure change is the second largest. And the pressure change rate can be effectively reduced, and the sound vibration performance is improved. In addition, the top dead center side (,
Since the piston speed of ()) is relatively lower than the piston speed of the bottom dead center side (,), the displacement of the piston compression top dead center position to the valve side due to the inertia force of the piston can be reduced. The clearance between the valve and the valve can be reduced, and it is possible to further increase the compression ratio. Further, since the piston speed is the slowest in the first half (stroke) of the intake stroke, the suction negative pressure is reduced, and by reducing the lift of the intake valve, the distance between the intake valve and the piston can be reduced to increase the compression ratio. Therefore, the thermal efficiency can be more effectively improved by increasing the compression ratio and the expansion ratio. (8). As described above, the lower link 2, the upper link 5, the control link 6, and the support position changing mechanism 13
It has. That is, it is a multi-link type configuration in which the engine compression ratio can be changed.

As described above, since the engine is of a multi-link configuration in which the engine compression ratio can be changed, the piston strokes (1) to (7), which are impossible with a single-link configuration, can be realized. It is possible to reduce pump loss, improve thermal efficiency by increasing compression ratio and expansion ratio, improve output by increasing charging efficiency, reduce combustion pressure change rate, and improve sound vibration performance by reducing crankshaft torque fluctuation, etc. Become. In addition, since the compression ratio can be changed, the following (9) to (12) can be realized. (9). The crank rotation angle (θ1 + θ2) from the latter half (stroke) of the compression or exhaust stroke to the first half (stroke) of the expansion or intake stroke is larger in the setting state of the high compression ratio than in the setting state of the low compression ratio ( Or equal).

If the closing timing of the intake valve is the same, when the engine compression ratio is increased, the maximum compression pressure is increased as compared with the case of the low compression ratio. Therefore, the latter half of the compression stroke (stroke) to the first half of the expansion stroke (stroke) ), And the sound-vibration performance tends to deteriorate. In order to achieve the same effective compression ratio as at the time of the low compression ratio, it is necessary to delay the closing timing of the intake valve by a well-known variable valve operating device or the like to delay the compression start timing. As a result, the pressure after the start of compression sharply rises, so that the pressure in the expansion stroke drops sharply. Therefore, the crank rotation angle θ1 + θ from the latter half of the compression stroke (stroke) to the first half of the expansion stroke (stroke)
By making 2 larger at high compression ratios than at low compression ratios, the piston speed at top dead center becomes relatively small when the high compression ratio is set, and the pressure change rate can be reduced. Therefore, sound vibration performance at a high compression ratio can be improved.

When the compression ratio is increased, the maximum compression ratio is limited because the distance between the piston and the valve is reduced, but the crank rotation from the latter half of the compression stroke (stroke) to the first half of the expansion stroke (stroke) is restricted. By making the angle θ1 + θ2 larger at the high compression ratio than at the low compression ratio, the piston speed at the top dead center becomes relatively small at this high compression ratio, and the high compression ratio at which the piston comes closest to the valve is obtained. Sometimes, the displacement of the piston position due to the inertial force at the time of high-speed rotation can be reduced, so that the clearance between the piston and the valve can be reduced, and a higher compression ratio can be achieved. Further, since the intake speed in the first half (stroke) of the intake stroke in the setting state of the high compression ratio becomes relatively slow, the intake valve can be made smaller without increasing the pump loss, and the compression ratio is further increased. It becomes possible. FIG. 5 shows the piston stroke characteristics of the present embodiment in which the stroke + the crank rotation angle θ1 + θ2 required for the stroke is relatively longer at a high compression ratio than at a low compression ratio. (10). The crank rotation angle (θ2 + θ3) in the expansion stroke or the intake stroke is larger (or equal) in the setting state of the high compression ratio than in the low compression ratio.

On the high compression ratio side, it is conceivable to improve the fuel efficiency by increasing the expansion ratio, and on the low compression ratio side, increase the charging efficiency to increase the output.
Since the compression stroke on the high compression ratio side is relatively shortened, the heat loss time due to compression can be reduced, and the effect of improving fuel efficiency by increasing the expansion ratio can be further increased. On the other hand, on the low compression ratio side, the piston speed during the intake stroke is relatively high, and the piston speed during the compression stroke is relatively low. The output can be improved by increasing the efficiency and the effective compression ratio. FIG. 5 shows the piston stroke characteristics of the present embodiment in which the stroke + the crank rotation angle θ2 + θ3 required for the stroke is relatively longer at a high compression ratio than at a low compression ratio. (11). The crank rotation angle θ2 in the first half of the expansion stroke or the intake stroke is larger (or equal) at the high compression ratio than at the low compression ratio.

As a result, the piston speed after the start of combustion in the stroke at a high compression ratio becomes relatively small, and the piston position after the end of combustion becomes closer to the top dead center. The expansion period of the heat is increased, and the thermal efficiency is improved. Therefore, the thermal efficiency is further improved by increasing the expansion ratio when the compression ratio is high. FIG. 5 shows the piston stroke characteristics of the embodiment in which the crank rotation angle θ2 required for the stroke at the time of the high compression ratio is extended on the high compression ratio side. (12). In addition to (11), the crank rotation angle θ4 in the compression stroke or the first half of the exhaust stroke is made smaller (or equal) at a high compression ratio than at a low compression ratio.

As shown in (11), in order to increase the crank rotation angle θ2 in the first half (stroke) of the expansion stroke in order to improve the thermal efficiency by increasing the expansion ratio at the time of a high compression ratio, the other strokes must be shortened. There is. In terms of sound and vibration performance, the first half of the compression stroke (θ4) in which the pressure in the cylinder is smaller than the latter half of the expansion stroke (θ3) where the combustion pressure is applied and the latter half (θ1) where the compression pressure is applied and the pressure rises due to the start of combustion. It is desirable to reduce the heat loss because it has the least practical harm, but rather increases the compression speed and improves the thermal efficiency. Therefore,
In (12), the crank rotation angle θ4 in the compression stroke or the first half of the exhaust stroke is smaller at a high compression ratio than at a low compression ratio. FIG. 5 shows the piston stroke characteristics of the present embodiment in which the crank rotation angle θ2 required for the stroke is lengthened and the crank rotation angle θ4 required for the stroke is shortened in such a setting state of the high compression ratio. I have.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a piston drive device for an internal combustion engine according to an embodiment of the present invention.

FIG. 2 is a graph showing a piston position with respect to a crank rotation angle in the embodiment (a) and a comparative example (b).

FIG. 3 is a graph showing a piston speed with respect to a crank rotation angle in the embodiment (a) and a comparative example (b).

FIG. 4 is a graph showing piston acceleration with respect to a crank rotation angle in the embodiment (a) and the comparative example (b).

FIG. 5 is a graph showing a piston position with respect to a crank rotation angle in a setting state of a high compression ratio (c) and a low compression ratio (d) of the embodiment.

FIG. 6 is a graph showing a piston speed with respect to a crank rotation angle in a setting state of a high compression ratio (c) and a low compression ratio (d) of the embodiment.

FIG. 7 is a graph showing a piston acceleration with respect to a crank rotation angle in a setting state of a high compression ratio (c) and a low compression ratio (d) of the embodiment.

FIG. 8 is an explanatory diagram for explaining the definition of each code.

FIG. 9 is an explanatory view showing the features of (1) to (7) of the embodiment.

FIG. 10 is an explanatory diagram showing the features of (9) to (12) of the embodiment.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 ... Crankshaft 2 ... Lower link 3 ... Piston 5 ... Upper link 6 ... Control link 13 ... Support position change mechanism

 ──────────────────────────────────────────────────続 き Continued on the front page (72) The inventor, Kenshi Ushijima, Nissan Motor Co., Ltd., 2 Takaracho, Kanagawa-ku, Yokohama, Kanagawa Prefecture (72) The inventor Katsuya Mogi, 2 Nihonsan Takaracho, Kanagawa-ku, Yokohama, Kanagawa F term (reference) 3G092 AA01 AB02 DD06 DG01 EA27 FA04 FA05 FA14 HA01Z HB01Z HB03Z HC01Z HE03Z 3J033 AA04

Claims (12)

[Claims] 1. A piston driving device for an internal combustion engine in which a piston reciprocating in a cylinder and a crankshaft cooperate with each other, wherein the piston moves between a bottom dead center and a top dead center in a first stroke.
The intermediate point, the middle of the stroke of the piston from top dead center to the bottom dead center is the second intermediate point, the crank rotation angle until the piston moves from the first middle point to the top dead center is θ1, and the piston is The crank rotation angle from the dead center to the second intermediate point is θ2, the crank rotation angle from the second intermediate point to the bottom dead center is θ3, and the piston is from the bottom dead center to the first intermediate point. (Θ1 + θ2) is equal to or more than (θ3 + θ4), where θ4 is a crank rotation angle required to move to a point. 2. The piston drive device for an internal combustion engine according to claim 1, wherein θ2 is equal to or larger than θ1. 3. A piston driving device for an internal combustion engine in which a piston reciprocating in a cylinder and a crankshaft cooperate with each other, wherein a middle of a stroke of the piston from a bottom dead center to a top dead center is set to a first position.
The intermediate point, the middle of the stroke of the piston from top dead center to the bottom dead center is the second intermediate point, the crank rotation angle until the piston moves from the first middle point to the top dead center is θ1, and the piston is The crank rotation angle from the dead center to the second intermediate point is θ2, the crank rotation angle from the second intermediate point to the bottom dead center is θ3, and the piston is from the bottom dead center to the first intermediate point. A piston driving device for an internal combustion engine, wherein θ2 is equal to or larger than θ3, where θ4 is a crank rotation angle required to move to a point. 4. The piston driving device for an internal combustion engine according to claim 3, wherein θ1 is equal to or larger than θ4. 5. The piston driving device for an internal combustion engine according to claim 3, wherein θ2 is equal to or larger than θ1. 6. The piston driving device for an internal combustion engine according to claim 4, wherein θ2 is equal to or larger than θ1. 7. The piston driving device for an internal combustion engine according to claim 6, wherein θ1 is equal to or larger than θ3. 8. A lower link rotatably attached to a crankpin of the crankshaft, an upper link rotatably connected at one end to the lower link, and rotatably connected at the other end to the piston, and one end. Is rotatably connected to the lower link, and the other end is rotatably connected to the engine main body side about a support fulcrum, and a control link that regulates the degree of freedom of the lower link,
The piston driving device for an internal combustion engine according to any one of claims 1 to 7, further comprising a support position changing mechanism that changes and controls a position of the support fulcrum. 9. The piston driving device for an internal combustion engine according to claim 8, wherein (θ1 + θ2) is larger at a high compression ratio than at a low compression ratio. 10. The piston drive device for an internal combustion engine according to claim 8, wherein (θ2 + θ3) is larger at a high compression ratio than at a low compression ratio. 11. The piston driving device for an internal combustion engine according to claim 8, wherein θ2 is larger at a high compression ratio than at a low compression ratio. 12. The piston driving device for an internal combustion engine according to claim 11, wherein θ4 is smaller at a high compression ratio than at a low compression ratio.