JP2001099231A - Roll vibration reduction device for internal combustion engine - Google Patents

Abstract

(57) [Problem] To reduce roll vibration during idling and avoid vibration deterioration due to resonance, regardless of the size and speed of the auxiliary flywheel system. SOLUTION: A first pulley 8 for driving an alternator 7 serving as a sub-flywheel system includes an inner peripheral portion 9 directly fixed to a crankshaft 2 and an outer peripheral portion 1 for receiving a belt 6.
And a rubber layer 12 serving as a spring element.
Are joined by A damper 13 is provided between the outer peripheral portion 10 and the inner peripheral portion 9 to attenuate the relative movement of the two in the rotational direction. The damper 13 is provided in the radial direction of the inner peripheral portion 9 so as to cross the middle portion of the groove. An orifice member 25 that can move forward and backward is provided. The orifice member 25 moves via hydraulic pressure according to operating conditions, and the damping constant is switched. The roll vibration reduction effect due to the anti-resonance of the vibration system can be obtained at the idling rotation speed. In the vicinity of the resonance speed, the damping action is strongly given, and the deterioration of vibration is suppressed.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a roll vibration reduction device for reducing a roll vibration of an internal combustion engine such as an automobile engine, which is caused by fluctuations in combustion pressure and the like.

[0002]

2. Description of the Related Art As a conventional roll vibration reducing device for an internal combustion engine, for example, a device disclosed in Japanese Patent Application Laid-Open No. Hei 6-33990 is known. The main flywheel consisting mainly of a crankshaft and a flywheel is provided with a sub flywheel system that is driven to rotate in the opposite direction to the main flywheel, and is generated by the reaction force acting on the bearings of these two flywheel systems. The roll vibration of the internal combustion engine is canceled by utilizing the roll moment.

[0003]

In the above configuration, the condition for completely canceling the engine roll vibration is given by the following equation (1).

[0004]

I ₁ = ρI ₂ (1) where I ₁ is the moment of inertia of the main flywheel system, I ₂
Is the moment of inertia of the sub flywheel system, and ρ is the speed increase ratio of the sub flywheel system.

In the vicinity of this condition, a large roll vibration reduction effect can be obtained. However, since the main flywheel system originally has a large moment of inertia, in order to completely cancel this, as is apparent from the above equation (1), it is necessary to increase the moment of inertia of the sub flywheel system sufficiently. Alternatively, it is necessary to secure a large speed increase ratio ρ. However, in order to increase the moment of inertia of the sub-flywheel system, a large sub-flywheel is required, which leads to an increase in the weight of the entire engine. Rotates at a high speed, so that the durability of the bearing portion is reduced. Due to these various factors, it is not practically possible to completely cancel the engine roll vibration with the above-described device, and the effect is limited.

The present invention has been made in view of such problems. Under the operating conditions that require a reduction in roll vibration, a greater effect can be obtained, and at the same time, under other operating conditions. An object of the present invention is to provide a roll vibration reduction device for an internal combustion engine, which suppresses undesirable vibration deterioration to a minimum.

[0007]

The invention according to claim 1 is
A main flywheel system that rotates integrally with the crankshaft of the internal combustion engine, a sub-flywheel system rotatably provided substantially parallel to the main flywheel system, and a sub-flywheel system that rotates the crankshaft. And a driving force transmission mechanism for transmitting the rotation of the auxiliary flywheel system to the auxiliary flywheel system to reduce the roll vibration of the internal combustion engine due to the rotation of the crankshaft by the rotation of the auxiliary flywheel system. In the roll vibration reduction device, a vibration element that resonates at a predetermined rotation speed is formed by interposing a spring element in a part of the power transmission system in the driving force transmission mechanism, and a damping constant of the driving force transmission mechanism is set to an internal combustion engine. The present invention is characterized in that damping constant variable means for changing the value in accordance with the operating condition is provided.

When the sub flywheel system is driven through the driving force transmission mechanism with the spring element interposed therebetween, a so-called spring-mass vibration system is formed, and resonance occurs at a specific rotation speed. With regard to the roll vibration, since anti-resonance occurs due to resonance, the roll vibration is suppressed to be low at this anti-resonance frequency. The frequency f _{0 at} which the above resonance occurs and the frequency f at which the anti-resonance occurs are expressed by the following equations (2) and (3).

[0009]

(Equation 2)

Where I ₁ is the moment of inertia of the main flywheel system, I ₂ is the moment of inertia of the sub flywheel system, and ρ is the speed increase ratio of the sub flywheel system (ρ in the case of a forward direction with respect to the engine rotation direction). > 0, ρ <for reverse direction
0) and k are the spring constants of the spring elements.

Therefore, if the anti-resonance frequency f is set so as to be in a rotational speed range in which roll vibration is to be reduced, for example, a frequency that becomes a problem in idle rotation, roll vibration during idling is greatly reduced. I do. In other words, even if the inertia moment and the speed increase ratio of the sub flywheel system are the same,
Due to the anti-resonance action of the vibration system, the roll vibration can be further reduced.

For example, assuming that the idle speed at which roll vibration should be reduced is 750 rpm, in the case of a four-cylinder engine, the secondary frequency of the engine speed becomes a problem.
What is necessary is just to set so that anti-resonance may come at 25 Hz.

At this time, the smaller the damping effect on the vibration system, the greater the effect of reducing the roll vibration due to anti-resonance.
However, if the damping action is small, the vibration of the sub flywheel system becomes excessively large in the resonance region. In the present invention, by changing the damping constant of the driving force transmission mechanism according to the operating conditions, while avoiding the deterioration of vibration at the time of resonance,
The roll vibration reduction effect in the region where the roll vibration is to be reduced can be enhanced.

According to a second aspect of the present invention, the attenuation constant changing means is configured to switch the attenuation constant in two stages.

According to the third aspect of the present invention, the damping constant under the first operating condition in which the frequency obtained by multiplying the rotation frequency of the internal combustion engine by (the number of cylinders / 2) substantially matches the resonance frequency of the vibration system is The rotation frequency of the internal combustion engine (number of cylinders / 2)
The damping constant variable means is configured such that the multiplied frequency is higher than the damping constant under the second operating condition that substantially matches the anti-resonance frequency of the roll vibration of the internal combustion engine accompanying the resonance of the vibration system. .

For example, assuming that the idle speed at which the roll vibration should be reduced is 750 rpm, in the case of a four-cylinder engine, the secondary frequency of the engine speed becomes a problem.
The anti-resonance is set at 25 Hz, and the damping constant at this time is preferably small. In contrast, resonance occurs when the engine speed is 810 rpm, for example.
It is desirable that the damping constant at this time is large.

According to a fourth aspect of the present invention, the spring element and the damping constant varying means are provided on a crank pulley coupled to a crankshaft, and the damping constant varying means has an orifice for generating a damping force. Components
The damping constant is switched by moving the damping constant by the hydraulic pressure supplied via the crankshaft. That is, the movement of the orifice member changes the passage area of the orifice through which the fluid flows, and changes the damping constant.

Further, the invention according to claim 5 is characterized in that the damping constant variable means uses an electromagnetic viscous fluid generating a damping force and changes the damping constant by controlling the applied voltage. In this case, the viscosity of the electromagnetic viscous fluid is controlled by the applied voltage, and the damping constant changes.

In the invention according to claim 6, the damping constant changing means switches the damping constant by moving the orifice member for generating the damping force by centrifugal force accompanying rotation of the engine. That is, since the centrifugal force becomes large in a high speed region of the engine and becomes small in a low speed region, the damping constant is mechanically switched using this.

As the spring element, an elastic body such as rubber can be used, or a metal spring can be used. Since a metal spring has a smaller damping of the spring element itself, a greater roll vibration reduction effect can be obtained by utilizing anti-resonance.

According to the present invention, the sub flywheel system rotates in a direction opposite to the main flywheel system. As is apparent from the above equation (3), when the rotation is performed in the opposite direction, the resonance frequency and the anti-resonance frequency are separated from each other, and the interval therebetween is widened. Compared to driving in the direction
The roll vibration reduction effect itself also increases.

According to a ninth aspect of the present invention, the damping constant variable means increases a damping constant in a rotation range where the vibration system resonates, and respectively sets a damping constant in a low rotation side and a high rotation side of the rotation range. Is reduced.

As a result, the vibration at the time of resonance is reliably attenuated, and the roll vibration can be reduced in the rotation speed region before and after the resonance.

[0024]

According to the apparatus for reducing roll vibration of an internal combustion engine according to the present invention, the roll vibration can be reduced by the anti-resonant action of the vibration system without increasing the size of the auxiliary flywheel system and increasing the speed increase ratio. In particular, by changing the damping constant of the driving force transmission mechanism according to the operating conditions, it is possible to avoid the deterioration of the vibration at the time of the resonance and to reduce the roll vibration in the region where the roll vibration should be reduced. The roll vibration reduction effect can be enhanced.

[0025]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Preferred embodiments of the present invention will be described below in detail with reference to the drawings.

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

FIG. 1 shows an entire in-line four-cylinder internal combustion engine for a vehicle equipped with the roll vibration reduction device of the present invention. The internal combustion engine body 1 has a flywheel (not shown) at the rear end. A main flywheel system mainly including the crankshaft 2 is provided, and auxiliary machines such as an alternator 7 driven by the crankshaft 2 are provided. As shown in FIG. 2, the alternator 7 has a disk-shaped additional inertial mass body 4, and the inertia moment of the additional inertia mass 4 and the inertia moment of the alternator 7 reduce the secondary flywheel for reducing roll vibration. The system is configured. The alternator 7 has a driving force transmission mechanism mainly composed of a crank pulley 5 coupled to the crankshaft 2, a pulley 34 on the alternator 7 side, and an accessory drive belt 6 wound between the two. Driven.

FIGS. 3 and 4 show details of the crank pulley 5 which is a part of the driving force transmission mechanism. The crank pulley 5 has a two-stage configuration including a large-diameter first pulley 8 located near the internal combustion engine main body 1 and a small-diameter second pulley 14 located outside the first pulley 8. The driven accessory drive belt 6 is hung on the first pulley 8. The first pulley 8 is divided into an inner peripheral portion 9 directly fixed to the crankshaft 2 and an outer peripheral portion 10 for receiving the accessory drive belt 6. 9 is rotatably supported via a bearing 11 and a rubber layer 1 serving as a spring element.
It is connected to the inner peripheral part 9 by 2. Further, a damper 13 is provided between the outer peripheral portion 10 and the inner peripheral portion 9 to provide a damping action for the relative movement of the two in the rotation direction.

Also, a second pulley 1 for driving another auxiliary machine
4 is also divided into an inner peripheral portion 15 and an outer peripheral portion 16 fixed to the crankshaft 2, and a rubber layer 17 is provided between the two. The spring constant of the rubber layer 17 and the moment of inertia of the outer peripheral portion 16 are tuned so as to function as a dynamic vibration absorber for suppressing torsional resonance.

The inner peripheral parts 9, 15 of the two pulleys 8, 14
Are integrally joined by bolts 18, and a sealing material 19 for preventing oil leakage is applied to a joint surface thereof, particularly, a portion around the damper 13.

FIG. 5 shows the details of the damper 13. As shown in the figure, the damper 13 fills an arc 21 with an oil 21 in an arc-shaped groove 20 provided in the inner peripheral portion 9, and at one end of the groove 20, an expandable and deformable bellows 22 is provided, and at the other end, Piston 23 coupled to outer periphery 10
, And the piston 23 slides in the groove 20 in the circumferential direction. Further, an orifice member 25 is provided so as to be able to advance and retreat along the radial direction of the inner peripheral portion 9 so as to cross the middle portion of the groove portion 20, and an orifice 24 is formed at the tip end of the orifice member 25. I have. The orifice member 25 includes a return spring 3
1, while being urged in an inner circumferential side, that is, a retreating direction, while being urged in an outer circumferential side, that is, a protruding direction by a hydraulic pressure supplied through an oil passage 27. As shown in FIG. 4, the above oil pressure is supplied from the oil pump 26 via the inside of the crankshaft 2 and determines whether the oil pressure is supplied to the oil passage 27 side or discharged to the drain 30 side. The switching valve 28 is controlled by a signal from a controller 29.

In the state where the hydraulic pressure is supplied, FIG.
As shown in (a), the tip of the orifice member 25 protrudes into the groove 20, and the groove 20 is partitioned by the orifice 24, so that the inner peripheral portion 9 and the outer peripheral portion 10 rotate relatively, and the piston 23 moves into the groove 20. When sliding inside, the oil 21
Flows through the orifice 24 to generate a large damping force.
On the other hand, when the supply of the hydraulic pressure is stopped by the valve 28, the orifice member 25 is retracted by the spring force of the return spring 31 as shown in FIG. The damping constant becomes smaller.

As described above, when the auxiliary flywheel system is driven via the spring element, a so-called spring-mass vibration system is formed, and resonance occurs at a predetermined frequency. Then, anti-resonance occurs due to this resonance, and the roll vibration is suppressed to a low level at the frequency of this anti-resonance.

The frequency f _{0 at which} this resonance occurs and the frequency f at which anti-resonance occurs are expressed by the above-mentioned equations (2) and (3).

By setting the anti-resonance frequency f to be a frequency that causes a problem in idle rotation,
Idle vibration can be significantly reduced. When the idling rotational speed is 750 rpm, in the case of a four-cylinder internal combustion engine, the secondary rotation frequency becomes a problem.
What is necessary is just to set so that anti-resonance may come to z.

FIG. 6 shows the characteristics of roll vibration in this embodiment. As shown in FIG.
When the damping due to the vibration is small, the vibration reduction effect at the anti-resonance near 750 rpm increases, but at around 810 rpm where the secondary frequency of the engine rotation becomes the resonance frequency, the deterioration of the vibration due to the resonance becomes remarkable. Therefore, in the present embodiment, the orifice member 25 is switched to the position shown in FIG. 5A in the region where the resonance frequency is 810 rpm and the vicinity thereof where the vibration is deteriorated, and the damping constant is increased. As a result, FIG.
As shown by the solid line in FIG. More specifically, a point where the characteristic when the damping constant is small and the characteristic when the damping constant is large intersects, for example, 780r
The attenuation constant is switched near pm. In a high-speed region beyond the resonance point, the magnitude of the damping does not significantly affect the vibration. Therefore, in order to reduce the work due to the oil pressure supply, the oil pressure supply is stopped and the damping is reduced. That is, as shown in FIG.
The attenuation constant becomes large in the region near m, and becomes smaller on both the low speed side and the high speed side.

In combination with such a variable control of the damping constant, finally, a roll vibration reduction characteristic as shown by a solid line in FIG. 6 is obtained. The figure also shows, as a reference example, the characteristics in the case where the driving force transmission mechanism is not provided with a spring element. As is apparent from the comparison between the two, the vicinity of the idle speed (750 rpm) where roll vibration is a problem is shown. Thus, the roll vibration can be significantly reduced.

In addition, since the rotation speed of about 1000 rpm or more is used when the vehicle is running, the deterioration of the vibration at the rotation speed near the resonance point rarely causes a problem. Such a rotational speed range may be used. In the present invention, even in such a case, deterioration of vibration due to resonance can be reliably avoided.

Next, a second embodiment of the present invention will be described with reference to FIGS.
It will be described based on.

In the second embodiment, as shown in FIGS. 8 and 9, the spring element of the first pulley 9 of the crank pulley 5 is changed to a metal coil spring 32 instead of rubber. According to such a metal coil spring 32, since the damping of the spring element itself is smaller than that of rubber, the damping of the entire vibration system combined with the damper 13 can be further reduced.

Therefore, as shown in the characteristic diagram of FIG.
By switching the damper 13 to a low damping constant near the idle speed (750 rpm) where the secondary rotational frequency becomes the anti-resonant frequency, a greater vibration reduction effect can be obtained. Also, the secondary rotational frequency becomes the resonance frequency.
In the vicinity of 10 rpm, since the attenuation of the spring element is reduced, the deterioration of vibration due to resonance becomes more remarkable.
The orifice 24 of No. 3 is set to have a smaller diameter. As a result, the damping action of the damper 13 increases, and deterioration of vibration can be sufficiently suppressed.

In addition, by using the metal coil spring 32, compared to rubber, it is possible to withstand a break to a larger amplitude, so that the rigidity of the spring can be further reduced.
Therefore, if the same anti-resonance frequency is secured, the moment of inertia of the sub flywheel system can be further reduced. Therefore, as shown in FIG. 7, it is not necessary to provide an additional inertial mass body in the alternator 7, and the anti-resonance frequency can be reduced to the secondary rotational frequency of the idle speed only by the inertia moment of the rotor 33 of the alternator 7. Becomes possible.

Next, a third embodiment will be described with reference to FIGS.

In the third embodiment, the spring element and the damping element are provided not on the crank pulley 5 but on the alternator pulley 34 side.

As shown in FIGS. 11 and 12, the alternator pulley 34 is divided into an inner peripheral portion 36 directly fixed to the rotating shaft 35 and an outer peripheral portion 37 on which the accessory drive belt 6 is hung. The outer peripheral portion 37 is rotatably supported on the inner peripheral portion 36 via a bearing 38 and is coupled to the inner peripheral portion 36 via a metal coil spring 32 and a damper 39. As described above, even in the configuration in which the spring element and the damping element are provided on the alternator pulley 34, the rotor portion 33 (see FIG. 7) of the alternator 7 and the inner peripheral portion 36 become the inertial mass of the sub flywheel system, and The same operation and effect as those of the first and second embodiments can be obtained.

In this embodiment, an electromagnetic viscous fluid 40 is used as a means for varying the damping constant of the damper 39. That is, as shown in FIG. 13, the damper 39 has an inner peripheral portion 36 similar to the above-described damper 13.
At one end of an arc-shaped groove 20 provided at the other end, a bellows 22 which can be expanded and contracted is provided.
And a piston 23 coupled to the
Is configured to slide in the groove 20 in the circumferential direction.
An electromagnetic viscous fluid 40 whose viscosity changes with voltage is sealed in the groove 20 instead of ordinary oil. The plus and minus electrodes 42 and 43 are arranged in the groove 20 via an insulator 41 such as a ceramic. Voltage application device 4 controlled by controller 29
When a voltage is applied between the electrodes 42 and 43 according to 4, the viscosity increases as the voltage increases, and the damping constant of the damper 39 increases.

Therefore, in the vicinity of 810 rpm at which the secondary rotation frequency becomes the resonance frequency, the attenuation is increased by applying a high voltage between the electrodes 42 and 43, and the deterioration of vibration due to resonance can be suppressed. In the vicinity of idle rotation (750 rpm) where the secondary frequency becomes the anti-resonance frequency, the voltage application is stopped to reduce the attenuation, and a large vibration reduction effect can be obtained. In a high rotation speed region beyond the resonance point, the voltage application is stopped to suppress power consumption. Thereby, characteristics similar to those in FIG. 10 can be obtained.

In this embodiment, the attenuation characteristics can be changed continuously by continuously changing the voltage applied to the electrodes 42 and 43. Therefore, by gradually changing the damping constant at the time of switching the damping constant, it is possible to avoid a sudden change in the tension of the accessory driving belt 6, prevent the occurrence of transient vibration of the belt 6, and improve the durability of the belt 6. Can be improved.

Next, a fourth embodiment will be described with reference to FIGS.

In the fourth embodiment, the means for varying the damping constant is mechanically switched by using centrifugal force instead of external control.

In this embodiment, as in the third embodiment, a spring element and a damping element are provided on the alternator pulley 34 side, and a metal coil spring 32 is provided between an inner peripheral portion 36 and an outer peripheral portion 37. And a damper 45 are provided.
The damper 45 is similar to the damper 13 described above, and
6 is filled with oil 21 in an arc-shaped groove 20, and one end of the groove 20 is provided with an expandable and contractible bellows 22.
And a piston 23 coupled to the outer peripheral portion 37 is disposed at the other end, and the piston 23 slides in the groove 20 in the circumferential direction. Further, an orifice member 46 is provided so as to be able to advance and retreat in the radial direction of the inner peripheral portion 36 so as to cross the intermediate portion of the groove portion 20, and the orifice 24 is formed at the tip end of the orifice member 46. I have. The orifice member 46 is constantly urged in an inner circumferential side, that is, a retreating direction by a return spring 49 disposed between a lower end 47 and a projection 48 of the inner circumferential portion 36.

When the alternator pulley 34 rotates, a centrifugal force is generated in the orifice member 46 toward the outer periphery, and the centrifugal force increases as the rotation speed increases. Then, the orifice member 46 slides to a position where the spring force of the return spring 49 and the centrifugal force balance. Therefore, at the time of low rotation, the orifice member 46 retreats inward as shown in FIG. 16A, and the damping constant decreases. At the time of high rotation, the orifice member 46
Is projected outward due to centrifugal force, and the oil moves through the orifice 24, so that the damping constant increases.

The damping constant is desirably large near 810 rpm at which the secondary rotation frequency becomes the resonance frequency, and becomes small near idle rotation (750 rpm) at which the secondary rotation frequency becomes the anti-resonance frequency. As the position of the orifice member 46 switches at the rotation speed,
Spring constant of return spring 49 and orifice member 4
6 is set. As a result, as in FIG. 10, it is possible to sufficiently suppress the vibration deterioration near the resonance point while obtaining a large vibration reduction effect near the idle speed. In particular, in this embodiment, a control means for variably controlling the damping constant is not required, so that the system can be simplified.

Next, a fifth embodiment will be described with reference to FIGS.

In the fifth embodiment, an alternator 7 serving as a sub-flywheel system is driven in the reverse direction with respect to the crankshaft 2 by driving the alternator 7 on the back surface of the accessory drive belt 6.

By rotating the motor in the opposite direction, the interval between the resonance frequency f ₀ and the anti-resonance frequency f is widened, as is apparent from the above equation (3). Therefore, when the damping constant is switched, even if there is a slight error in the switching point with respect to the rotation speed, a small damping constant is obtained at the rotation speed at which the anti-resonance frequency is obtained, and a large damping constant is obtained at the rotation speed at which the resonance frequency is obtained. Can be obtained. In other words, a larger error can be tolerated in the rotational speed at which the switching is actually performed, and the setting thereof becomes easier. Further, as shown in FIG. 18, the effect of reducing the roll vibration due to the anti-resonance is further increased as compared with the case where the crankshaft 2 is rotated in the same direction. It should be noted that any of the configurations of the first to fourth embodiments described above can be applied to the reverse drive as in the fifth embodiment.

[Brief description of the drawings]

FIG. 1 is a front view of an internal combustion engine for a vehicle according to a first embodiment of the present invention.

FIG. 2 is a side view of the alternator according to the first embodiment.

FIG. 3 is a sectional view of the crank pulley according to the first embodiment, taken along line AA of FIG. 4;

FIG. 4 is a sectional view of the crank pulley according to the first embodiment, taken along line BB in FIG. 3;

FIGS. 5A and 5B are cross-sectional views showing details of the damper of the first embodiment and showing a switching state thereof.

FIG. 6 is a characteristic diagram showing characteristics of roll vibration in the first embodiment.

FIG. 7 is a partially cutaway sectional view of an alternator according to a second embodiment.

FIG. 8 is a sectional view of the crank pulley of the second embodiment, taken along line AA of FIG. 9;

FIG. 9 is a sectional view of the crank pulley of the second embodiment, taken along line BB of FIG. 8;

FIG. 10 is a characteristic diagram showing characteristics of roll vibration in the second embodiment.

FIG. 11 is a sectional view of the alternator pulley according to the third embodiment, taken along line AA of FIG. 12;

FIG. 12 is a sectional view of the alternator pulley of the third embodiment, taken along line BB of FIG. 11;

FIG. 13 is an enlarged sectional view showing details of a damper according to a third embodiment.

FIG. 14 is a sectional view of the alternator pulley of the fourth embodiment, taken along line AA of FIG. 13;

FIG. 15 is a sectional view of the alternator pulley of the fourth embodiment, taken along line BB of FIG. 14;

FIGS. 16A and 16B are cross-sectional views showing details of a damper according to a fourth embodiment and showing a switching state thereof.

FIG. 17 is a front view of an internal combustion engine for a vehicle according to a fifth embodiment.

FIG. 18 is a characteristic diagram showing characteristics of roll vibration in the fifth embodiment.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 5 ... Crank pulley 7 ... Alternator 8 ... First pulley 9 ... Inner peripheral part 10 ... Outer peripheral part 12 ... Rubber layer 13 ... Damper 25 ... Orifice member 32 ... Metal coil spring 34 ... Alternator pulley 39 ... Damper 40 ... Electromagnetic viscosity Fluid 45 ... Damper 46 ... Orifice member

Claims (9)

[Claims] 1. A main flywheel system rotating integrally with a crankshaft of an internal combustion engine, a sub flywheel system rotatably provided substantially parallel to the main flywheel system, and a rotational driving force of the crankshaft. A drive force transmitting mechanism for transmitting the rotation to the sub flywheel system by transmitting the rotation to the sub flywheel system, wherein the rotation of the sub flywheel system reduces roll vibration of the internal combustion engine accompanying rotation of the crankshaft. In the roll vibration reduction device for an internal combustion engine, a spring element is interposed in a part of the power transmission system in the driving force transmission mechanism to form a vibration system that resonates at a predetermined rotation speed. A roll vibration reduction device for an internal combustion engine, comprising: a damping constant variable means for changing a damping constant according to an operating condition of the internal combustion engine. 2. The roll vibration reduction device for an internal combustion engine according to claim 1, wherein said damping constant variable means is configured to switch the damping constant in two stages. 3. The rotational frequency of the internal combustion engine is set to (number of cylinders /
2) The frequency obtained by multiplying the rotation frequency of the internal combustion engine by (number of cylinders / 2) is the damping constant under the first operating condition in which the multiplied frequency substantially matches the resonance frequency of the vibration system. 3. The damping-constant varying means is configured to be larger than a damping constant under a second operating condition that substantially coincides with an anti-resonance frequency in roll vibration of the internal combustion engine accompanying the above. A device for reducing roll vibration of an internal combustion engine according to claim 1. 4. The crank element according to claim 1, wherein the spring element and the damping constant variable means are provided on a crank pulley coupled to a crankshaft, and the damping constant variable means includes an orifice member for generating a damping force and a crankshaft. By the hydraulic pressure supplied via the
The roll vibration reduction device for an internal combustion engine according to any one of claims 1 to 3, wherein a damping constant is switched. 5. The internal combustion engine according to claim 1, wherein said damping constant variable means uses an electromagnetic viscous fluid generating a damping force and changes the damping constant by controlling an applied voltage. Engine roll vibration reduction device. 6. The damping constant changing means according to claim 1, wherein said damping constant changing means switches the damping constant by moving an orifice member for generating a damping force by centrifugal force accompanying rotation of the engine. The roll vibration reduction device for an internal combustion engine according to any one of the above. 7. The roll vibration reduction device for an internal combustion engine according to claim 1, wherein the spring element is a metal spring. 8. The roll vibration reduction device for an internal combustion engine according to claim 1, wherein said auxiliary flywheel system rotates in a direction opposite to said main flywheel system. 9. The variable damping constant means increases a damping constant in a rotation range where the vibration system resonates, and decreases a damping constant in a region on a lower rotation side and a higher rotation side than the rotation region. The roll vibration reduction device for an internal combustion engine according to claim 1 or 2, wherein: