JPS6357938A - Viscous damper - Google Patents

Abstract

PURPOSE:To obtain the small-sized viscous damper which has the large vibration suppressing effect by filling the gap between a casing and an inertia body with the high viscous liquid which possesses viscoelasticity and making the ratio of the characteristic frequency of a damper system for the characteristic frequency of a crankshaft system to 0.5 or more. CONSTITUTION:The part between a casing 2 and an inertia body 3 is filled with the high viscous silicone oil, etc. which possesses the viscoelastic property having a viscosity of about 200,000Cst or more and a small viscosity change due to heat. The vibration suppressing characteristic as a viscoelastic damper can be provided, and the damper 1 can be made small-sized. Because of the elastic effect as the viscoelastic damper, the characteristic frequency ratio of the damper can be made 0.5 or more, and the vibration suppressing effect can be increased by making the damping coefficient ratio to 0.2-0.45. Further, since the dynamic shear modulus which represents the elastic property of the viscous liquid can be made about 0.2kg/cm<2> rod or more in case of silicone oil, the clearance of the gap can be made large, and the dispersion of the vibration suppressing characteristic can be reduced.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to a viscous damper used for damping torsional vibrations occurring in the crankshaft of an internal combustion engine.

(Prior Art) In an internal combustion engine with a large number of cylinders and a long crankshaft, this is due to fluctuations in drive torque for each cylinder.

When torsional vibration occurs in the crankshaft and the natural frequency of the crankshaft due to this torsional vibration matches the excitation torque, the amplitude of the torsional vibration increases due to resonance, causing torsional damage to the crankshaft and engine noise.

Therefore, in order to prevent damage to the crankshaft due to this torsional vibration and to prevent carrot noise caused by the torsional vibration, various damper devices have been devised to absorb and suppress torsional vibration.

As shown in FIG. 14, these damper devices 105.

It is attached directly to the shaft end of the crankshaft 103 of the internal combustion engine 101 (the shaft end on the opposite side when the flywheel 102 is installed) or via a crankshaft accessory such as a pulley 104, and is attached to the shaft end of the crankshaft 103 of the internal combustion engine 101 to prevent twisting of the crankshaft 103. It divides or attenuates the resonance peak of the vibration amplitude to lower its resonance level, and acts to absorb the torsional vibration and amplitude of the crankshaft 103, respectively.

Conventionally, as a general example of this damper device, as shown in FIG. Elastic member 109
A rubber type damper device (hereinafter referred to as a rubber damper) having a structure in which an annular metal inertial body 110 is connected via a
106 or as shown in FIG. 16, a ring box-shaped casing 113 is integrally formed along the outer periphery of a damper plate 112 attached to the crankshaft, and a predetermined gap is provided inside the casing 113 with the inner wall of the casing 113. A viscous damper 111 is known in which a ring-shaped metallic inertial body 114 is housed at a distance therebetween, and a gap 115 between the inertial body 114, the inner wall of the casing 113, and the cut is filled with a viscous liquid such as silicone oil. Of these two types of typical damper devices, the rubber damper 106 is a type of viscoelastic damper, and uses viscoelasticity due to shear strain of an elastic body to suppress the torsional vibration generated in the crankshaft to the resonance peak of the amplitude of the torsional vibration. divide, and
It converts the energy of torsional vibration into thermal energy, dissipates it, and attenuates it, suppressing torsional vibration. On the other hand, the conventionally known viscous damper 111 is a viscous damper that converts the energy of torsional vibration generated on the crankshaft due to the viscous resistance of a viscous liquid such as silicone oil into thermal energy and dissipates it. (viscous damping) to suppress torsional vibration.

(Problems to be solved by the invention) In recent years, there has been an increasing demand for lower heat costs, lower pollution, higher output, etc. of internal combustion engines, and there has been a particularly strong trend toward supercharging diesel engines for vehicles. Progress is being made towards higher supercharging through the adoption of intercoolers, etc. This trend is spreading from engines for large vehicles to engines for small and medium-sized vehicles.

However, in engines such as in-line multi-cylinder engines, where torsional vibration occurs due to factors such as a long crankshaft, and which requires a damper to suppress the torsional vibration, high supercharging is required to increase the crankshaft. However, with the rubber damper, there is a limit to the ability to damp torsional vibrations because the elasticity of the elastic member is limited. Furthermore, there are problems such as an increase in strain in the elastic member and deformation, deterioration, and damage of the elastic member due to heat generation, and it cannot be applied to torsional vibrations with large vibrations. On the other hand, viscous dampers can use heat-resistant materials such as silicone oil for the viscous liquid, so they are resistant to high temperatures and have fewer problems such as deterioration.In addition, vibration can be suppressed by increasing the moment of inertia of the inertial body. Improved ability can be measured.

However, with this conventional viscous damper. Since torsional vibration is suppressed by the damping effect caused by the viscous resistance of the viscous liquid, the moment of inertia of the inertial body must be increased to improve the damping ability, which reduces the weight burden on the crankshaft due to the increased weight of the damper. This increases the size and tends to cause deflection and breakage of the crankshaft, and the additional weight increases the size of the damper, causing problems such as space for installation.

(Means for Solving the Problems) The present invention aims to solve the problems of the damper device according to the prior art, and provides the following means.

It is formed by a ring box-shaped casing that is attached to the crankshaft and rotates together with the crankshaft, and an annular inertial body that is housed within the casing with a predetermined gap between the inner wall of the casing and the casing and the above-mentioned casing. In a viscous damper made by filling the gap between the inertial body and the viscous liquid,
A high viscosity liquid having viscoelasticity is used as the viscous liquid, and the ratio of the natural frequency fd (fd/fe) of the viscous damper system to the natural frequency fe of the crankshaft system is 0.5.
The above shall apply.

(Function) By using the above-mentioned high viscosity viscous liquid, the viscous damper according to the present invention acts like a viscoelastic damper against torsional vibration, and can set the natural frequency ratio fd/fe to 0.5 or more. .

(Example) Hereinafter, the present invention will be explained based on the drawings.

FIG. 1 shows an example of a viscous damper shape to which the present invention is applied.

In the figure, a viscous damper 1 is a damper plate portion that is attached directly or via a crankshaft bending part such as a pulley to the end of the crankshaft where the torsional amplitude is large due to torsional vibration of the crankshaft (not shown). 6, an annular cylindrical casing 2 with a square cross section integrally formed with the damper plate part 6 along the outer periphery of the damper plate part 6, and a predetermined gap hs between the inner wall of the casing 2 and the inner wall of the casing 2.
hi.

The casing 2 and the annular inertial body 3 with a square cross section are housed at a distance of 10 mm, and the gap between the casing 2 and the inertial body 3 is filled with a high-viscosity viscous liquid 4 having viscoelasticity.

The damper plate portion 6 and the casing 2 are formed by casting or cutting a ferrous metal that has relatively good heat dissipation properties and strong mechanical strength.

Note that in order to improve heat dissipation and to reduce the influence of the moment of inertia of the casing 2, it is possible to use a light aluminum alloy, but this poses a problem in terms of strength. It is preferable to use a heavy material for the inertial body 3 in order to increase the moment of inertia, but a button cannot withstand use due to its low strength and rigidity. For this reason, it is cast or cut using iron or iron-based alloys, which are relatively heavy and strong. It is formed by J processing or the like.

The viscous liquid, which is an important element in the characteristics of the viscous damper according to the present invention, includes high viscosity silicone oil, etc., which has a viscosity of 200,000 C5t or more, has viscoelastic properties, is heat resistant, and has little change in viscosity due to heat. use.

By the way, in the conventional viscous damper, a relatively low viscosity viscous liquid is used as the viscous liquid. For this reason, the elastic effect of the viscous liquid is hardly obtained, and the torsional vibration is suppressed by the viscous damping effect due to the viscous resistance of the viscous liquid. However, in the present invention, as described above, the viscous liquid has viscoelasticity in a high viscosity state. Since high viscosity silicone oil etc. are used, a damping effect and a vibration damping effect (dynamic vibration absorption) due to elasticity work, and the characteristics are as follows.
Similar to a rubber damper that utilizes the viscoelastic effect of an elastic body such as rubber, vibration damping characteristics similar to a viscoelastic damper can be obtained.

The damping characteristics of the viscous damper according to the present invention will be explained below using a simple vibration model.

FIG. 2 shows a multi-degree-of-freedom equivalent vibration model for a generally known multi-cylinder engine (six cylinders in the figure) crankshaft system. In the figure, Id indicates the moment of inertia of the damper system 7, Id indicates the moment of inertia of the pulley etc. 10, ■2
~I7 represents the moment of inertia of each cylinder 8, and Xa represents the moment of inertia of the flywheel 9. Also, Kd in the same figure
is the spring constant of the damper system 7, Cd is the damping coefficient of the damper system 7, 1 is the spring constant between each crankshaft system, C
+-C7 indicates the damping coefficient between each crankshaft system. If the vibration characteristics of the damper system 7 are known, simulation results using this multi-degree-of-freedom vibration model are known to be in good agreement with actual results, and predictive calculations can be utilized with sufficient accuracy.

As mentioned above, the viscous damper according to the present invention utilizes viscoelasticity similarly to the rubber damper.

This viscoelastic property changes depending on usage conditions such as temperature, amplitude, frequency, etc., and the material, so it is important to understand this property when designing a viscous damper.

Generally, for the excitation torque frequency depending on the main or rank shaft rotation order in the normal engine rotation range.

Since the natural frequency of the crankshaft torsional vibration resonance mode curve is higher than the secondary mode, only the primary mode needs to be considered. That is, the crankshaft system including the damper can be represented by an equivalent vibration model of a two-degree-of-freedom system as shown in FIG. In the figure, Id is the moment of inertia of the damper system 11, K
d represents a spring constant, and Cd represents a damping coefficient. Furthermore, the moment of inertia Ie of the crankshaft system 12 indicates the resultant moment of inertia of the entire crankshaft system excluding the damper system 11. Also,
Ke and Ce represent the rigidity of the crankshaft system, the spring constant and damping coefficient of the crankshaft system due to friction of the bearings, etc.

Generally, the equation of motion of the two-degree-of-freedom equivalent vibration model shown in Figure 3 is: IdMd+Cd'(Xd-@Xe) = Mt. O(1)I
eXe+Cd (Xe-Xd)+CeXe+KeXe
It can be expressed as =F=Fo Cos (1) and (2).

Here, when expressed using complex representation, the excitation force F is F = Foa''t・-(3) The response amplitudes Xe and Xd of the crankshaft system and damper system are X
e=χeejwt...(4a) Xd=, Cd e"" --
・(4b) Also, the complex damping coefficient C♂ is Cd' = Cd −j・(Kd/ω) ・・
・It becomes (5). Here, Cd is the damping coefficient of the damper, K
d represents the spring constant of the damper, and ω represents the excitation angular velocity.

Next, if we make equations (1) and (2) dimensionless, −λ2
θd+(λd”+j−2ζdλ)(θd−θe)=o・
...(6)−λ8θc+μ(λd”+j・2ζdλ)(
Oe−Cd)+(1+j・2ζCλ)θe=o
...(7). Here, μ= Id/Ie・: Moment of inertia ratio (8
) ωd=, /'iii'a700: Damper system natural angular frequency... (9a) ωe=J remainder 7: Crankshaft system natural angular frequency... (9b) λd=ωd/ωe: Eigen Angular frequency ratio...(1o)λ=ω/ωe :Excitation frequency ratio...(11)ζd=Cd/(2Idωe
): Damping coefficient ratio of damper system - (12a) ζB = (:e/
(2Iec, +e): Crankshaft system damping coefficient ratio (1
2b) χo=Fo/Ke: Static angle change
---(13) θd=χd/xo: Damper system response amplitude ratio...(14a) Oe =Xe/χ0
: Crankshaft system response amplitude ratio (14a).

Since the viscous damper according to the present invention uses high viscosity silicone oil, it can be treated as a viscoelastic damper. Therefore, in equation (5) representing the complex damping coefficient Cd' of the damper, the spring constant of the damper system is KdK0. Therefore, as shown in FIG. 4, the peak Po of the torsional resonance 13 of the crankshaft system is divided into two resonance peaks, Ifffl 14 and node 15, by the damper system.

In the same figure, the resonance peak apex A (Al
The positions of tA, ) and B (Bi, 8U) move on the graph using the damper's spring constant Kd and damping coefficient Cd as parameters. Here, in order to suppress the torsional amplitude to a low level in a wide frequency range, it is optimal to make the resonance peak amplitudes of points A and B equal and lowest (0ρtimu
m). In this case, the I and II node resonance peak points are points Aopt and Bopt in FIG. 4.

By the way, in the conventional viscous damper, since a relatively low viscosity silicone oil or the like is used as the viscous liquid, in equation (5) representing the complex damping coefficient Cd'' of the damper.

The spring constant Kd of the damper is 0, and as shown in FIG. 5, there is only one resonance peak of the crankshaft system even when the damper is installed. That is, since the conventional viscous damper has almost no dynamic vibration absorption effect due to elasticity, vibrations are absorbed only by viscous damping due to viscous resistance. Therefore, the point where the damper's damping coefficient Cd=0 is Po, which coincides with the resonance peak of the crankshaft system without the damper. Also,
In the conventional viscous damper, as the damper damping coefficient Cd increases, the resonance peak apex Po decreases, and when the minimum amplitude point Popt is exceeded, the amplitude increases again and becomes cd-4.
When it becomes oo, it reaches point Q. That is, at one point Q, the crankshaft system and damper system of the vibration model shown in FIG. 2 are in a state of vibration as one unit.

Here, the optimal conditions and conditions for damping in FIGS. 4 and 5 are summarized in the table below. It should be noted that since it is considered that the damping coefficient C6 of the crankshaft system is relatively sufficiently small compared to the damping coefficient Cd of the damper at the optimum time, Ce=O is set to simplify the development of Equation 1. Further, μ in the table below is the moment of inertia ratio μ=Id/Ie shown in equation (8) above.

FIG. 6 is a diagram showing the relationship between the viscoelastic characteristics of the viscous damper according to the present invention, the amplitude θ at the resonance peak apex, and the frequency ratio λ. The relationship between the resonance frequency λ and the amplitude O of the resonance peak apex (one resonance peak when Kd=O) is shown. Here, the inertia moment ratio μ=Id/τe of the damper system and the crankshaft system is constant. In the same figure, I
, the position of the resonance peak apex of the II node vibration is λ=λ龜(λ
α indicates the resonant frequency at point Q shown in FIG. In the case of a conventional viscous damper, i.e., a viscous damper (Kd=O), the resonance peak has its apex in the region λ≧λα, and in the case of the viscous damper according to the invention, i.e., a viscoelastic damper (Kl=
This can be regarded as a special example of the node resonance peak of O).

In addition, the resonance peak apex Pop indicating the optimum time of the viscous damper
For the amplitude of t, I, II1 at the optimum time of the viscoelastic damper
Nodal resonance peak point Aopt.

As I showed my ex-husband, the amplitude of Bopt is 1/near T
The ratio of E Otsu to 5 is low. .

That is, damping of the viscous damper is achieved by consuming vibrational energy as heat that is proportional to the square of the relative shear velocity amplitude between the moment of inertia (or mass of inertia) of the crankshaft system and the damper system. Therefore, when the damper damping coefficient Cd increases and the relative amplitude between the damper system and the crankshaft system is suppressed, no more vibration energy can be consumed, and the resonance peak reaches the maximum consumption point, that is, Pop. The amplitude has a minimum value, and even if the Cd value is further increased, the amplitude will increase. On the other hand, in a viscoelastic damper such as the viscous damper of the present invention, a part of the vibration energy of the entire system is shared by the spring system due to the spring effect of the damper system, so that the vibration energy of the crankshaft system is reduced. By optimizing the characteristics of the damper, dividing the grip vibration, lowering its resonance peak, and appropriately combining the effects of consuming vibration energy as heat using the damping coefficient Cd term, a large vibration damping effect can be obtained. It will be done.

As described above, the viscous damper as a viscoelastic damper according to the present invention has more efficient vibration damping ability than the viscous damper as a conventional viscous damper. This is clear from the comparison of the amplitude at the resonance peak points Aopt and Bopt at the optimum time of the viscoelastic damper shown in Fig. 6 and the amplitude at the resonance peak point Popt at the optimum time of the viscous damper. The damping ability of the viscoelastic damper in terms of moment ratio is doubled.

As explained above, the viscous damper according to the present invention can be handled as a viscoelastic damper by using high viscosity silicone oil having viscoelasticity as the viscous liquid, and compared with the viscous damper as a conventional viscous damper. It is possible to measure the improvement of the vibration damping ability against torsional vibration.

Next, the complex damping coefficient Cd' that provides the damping characteristics of the viscous damper of the present invention will be described.

This complex damping coefficient Cd can be calculated from the damper constant A representing the shape of the damper and the dynamic characteristics of the silicone oil. The damper constant is determined using the viscous damper having the shape shown in FIG. Here, ri represents the inner radius of the inertial body 3, and V″0 represents the outer radius. Using this damper constant A, the complex damping coefficient Cd′ is expressed as follows.
Cd=A・μ”...(16
). Here, μm indicates the complex viscosity coefficient.

G′ r=μ′−j− (17) ω. μ′ and G° are coefficients representing the dynamic characteristics of viscous liquid,
μ′ is the dynamic viscosity coefficient, and Q l is the dynamic shear elastic coefficient, which is an important factor in determining the damping effect 1 elastic effect of the damper. Figures 7 and 8 show the silicone oil viscosity and G' when the measurement temperature and frequency are constant.

The relationship between μ′ is shown.

In addition, the dynamic shear modulus G' at a temperature of 30°C, 1 frequency, IQOHz, and a silicone oil viscosity of 10xlO'C5t.
is 80-90X10-3 Kg/am', and the dynamic viscosity coefficient μ' is 4X10-' KgS/Cm2.

Further, the spring constant Kd, the damping coefficient Cd, and the above QI, which are important elements for the characteristics of a viscoelastic damper.

To find the relationship with μ′, using equation (5), Kd=A−
G'...(18), Cd=A・μ'・
...(19), respectively, and the dynamic shear modulus G
′, is proportional to the dynamic viscosity coefficient μ′.

Figures 9 and 10 show the data obtained by actually measuring the relationship between the spring constant Kd, damping coefficient Cd, and silicone oil viscosity with a viscous damper installed in an in-line 6-cylinder engine under constant degrees and constant frequency, respectively. This is a diagram originally created.

The above two characteristic curves are the dynamic shear modulus G' and the dynamic viscosity coefficient μ' shown in FIGS. 7 and 8, respectively.
It shows the same characteristics as the characteristic curve showing the relationship between silicone oil viscosity and silicone oil viscosity. As shown in FIG. 9, it can be seen that the spring constant Kd of the viscous damper decreases rapidly below 500,000 CS, and the damping ability of the viscoelastic damper decreases below 500,000 CS. As described above, it is desirable to use high viscosity silicone oil having a viscosity of about 500,000 CS for a viscous damper as a viscoelastic damper.

FIG. 11 shows the results of comparing the measured values of the rotational 6th order component of the crankshaft torsional vibration when the viscosity of the silicone oil of the viscous damper was changed in an in-line 6-cylinder engine equipped with a viscous damper. In addition, the theoretical value of the viscous damper (Kd=O) in Fig. 5 was calculated by assuming that the horizontal axis representing the engine speed ratio to the maximum engine speed is the axis of the excitation frequency corresponding to the crankshaft rotation @ 6th order. The optimum resonance peak apex Pope is also shown in the figure. In FIG.

The dashed line in A is 200,000 CS silicon, and the solid line in B is 5
00,000 cst,, the chain double dotted line of c is 1 million C3t, D
The broken line indicates the case of a conventional viscous damper (Kd=O).

°As is clear from the same figure, the viscosity of silicone oil is 200,000 C3.
t, there is almost no difference in effectiveness compared to conventional viscous dampers;
, the torsional amplitude of the resonance peak decreases significantly, and the resonance frequency also changes significantly.

In addition, if the silicone oil viscosity is set to 1,000,000S,
Compared to the case of 00,000C3L, there is almost no difference in torsional amplitude, but the resonance frequency has changed.

In this way, a viscous damper with a silicone oil viscosity of 500,000 C or more acts effectively as a viscoelastic damper when considered on the diagram of the amplitude-frequency ratio characteristic of the resonance peak apex shown in Figure 6. Therefore, the vibration can be suppressed to a lower vibration level than the optimum point Popt of a conventional viscous damper.

Next, optimal design conditions for a viscous damper as a viscoelastic damper according to the present invention will be determined. Note that, as described above, when the silicone oil viscosity is about 500,000 CS, the viscoelastic properties are stabilized. Previously, the dynamic characteristics of the damper and the amplitude-frequency ratio relationship of the torsional resonance peak were shown in Fig. 6 using the damper's spring constant Kd and damping coefficient Cd as parameters. Torsional amplitude vs. frequency obtained based on the results of actual measurements using the ratio fd/fe of the natural frequency fd and the damping coefficient ratio ζd of the damper as parameters, the temperature and moment of inertia ratio being constant, and the crankshaft natural frequency fe being approximately 200 Hz. The relationship between the ratios is shown in FIG.

The natural frequency fd of the damper system is expressed as ω in equation (9a).
As d=2πfd.

Similarly, the natural frequency fe of the crankshaft system is also (9
b) From formula. Also, the damping coefficient ratio ζd of the damper is A from (12a).

In addition, the line fd=o in FIG. 12 represents the characteristics of a conventional viscous damper as a viscous damper with Kd=O,
t is its optimal value.

As is clear from the figure, the optimum design conditions for maintaining the damping ability of the viscous damper according to the present invention as a viscoelastic damper and increasing the degree of freedom in design are as follows: ≧0.5.

That is, by designing the damper so that the natural frequency ratio is 0.5 or more, a damping effect as a viscoelastic damper can be obtained.

By the way, in order to ensure the vibration damping effect of the viscous damper, the natural frequency ratio should be set to 0.5≦fd/fe≦1.0.
It is better to

Furthermore, the damping effect can be further increased by limiting the damping coefficient ratio ζd to 0.2≦ζd≦0.45.

By the way, as described above, the spring constant d of the damper is proportional to the dynamic shear elastic coefficient G' of the high viscosity liquid and the damper constant A. Also, the spring constant K of the damper is
d and the moment of inertia Id of the damper.

Therefore, using equation (18) Kd=A-G', becomes.

Further, as described above, the damper constant A is determined by the damper shape, and is the case for the viscous damper having the shape example shown in FIG.

From the above, in the viscous damper according to the present invention, the moment of inertia Id of the damper, the damper constant A, and the dynamic shear elastic coefficient of the high viscosity liquid are set such that the natural frequency ratio fd/fe is 0.5 or more. Determine G'.

Figure 13 shows the gap hs between the casing 2 of the viscous damper and the side surface of the inertial body 3, and the high viscosity silicone oil 4, with the damper's natural frequency ratio fd/fe constant.
The results of actually measuring the relationship between the dynamic shear modulus G' and the dynamic shear modulus G' are shown below. As is clear from the figure, the clearance hs can be increased by increasing the dynamic shear elastic modulus G' while keeping the natural frequency ratio constant, that is, the damping effect is kept approximately constant. In addition, when using silicone oil, the dynamic shear modulus G' is G'≧0.2Kg/am2・ra
d allows for a larger clearance hs and a smaller moment of inertia of the damper.

Further, by increasing the clearance hs, it is possible to reduce the variation in the clearance hs, and it is possible to reduce the variation in the dynamic characteristics of the damper, that is, the variation in the damping characteristics.

(Effects of the present invention) As explained above, the viscous damper according to the present invention can have the distribution characteristics as a viscoelastic damper by using a high viscosity liquid such as high viscosity silicone oil as the viscous liquid. Compared to a conventional viscous damper, an excellent vibration damping effect can be obtained, and the damper can be made smaller. Further, in the viscous damper according to the present invention, the natural frequency ratio of the damper can be set to 0.5 or more due to the elastic effect as a viscoelastic damper.

Furthermore, by setting the damping coefficient ratio ζd to 0.2≦ζd≦0.45, the damping effect can be further increased.

In addition, the dynamic shear modulus of elasticity, which represents the elastic properties of viscous liquid, can be increased to 0.2 Kg/am" rad or more in the case of silicone oil, so a large clearance can be achieved between the casing of the viscous damper and the inertial body. Therefore, variations in the dynamic characteristics of the viscous damper, that is, variations in the damping characteristics, can be reduced.

[Brief explanation of the drawing]

Fig. 1 is a sectional view of a main part showing an example of the shape of a viscous damper according to the present invention, Fig. 2 is an explanatory diagram of a multi-degree-of-freedom equivalent vibration model of a crankshaft system of a multi-cylinder engine, and Fig. 3 is a cross-sectional view of a crankshaft system of a multi-cylinder engine. An explanatory diagram of a vibration model replaced with an equivalent vibration model with two degrees of freedom, and FIGS. 4, 5, and 6 are vibration characteristic diagrams of the damper device. FIG. 7 and FIG. 8 are diagrams showing the relationship between dynamic shear modulus of elasticity, dynamic viscosity coefficient, and silicone oil viscosity. Figures 9 and 10 are diagrams showing the relationship between the spring constant and damping coefficient of the damper and the silicone oil viscosity, and Figure 11 is the relationship between the engine speed ratio and torsional amplitude when the silicone oil viscosity of the viscous damper is used as a parameter. Figure 12 is a diagram showing the relationship between torsional vibration and frequency ratio when the natural frequency ratio and damping coefficient ratio are used as parameters, and Figure 413 is a diagram showing the relationship between the dynamic shear elastic modulus and the damper's 14 is a side view of an internal combustion engine with a damper device installed; FIG. 15 is a sectional view of the main parts of a conventional rubber damper;
FIG. 6 is a sectional view of a main part of a conventional viscous damper. 1... Viscous damper, 2... Casing, 3
...Inertial body, 4...Viscous liquid. 74 Art and Crafts〉shi゛-turn/transferred r and (according to the magnification bar 2 constant (Niu C powder pma) crazy g times Q (ζj=oo) h yS prisoner !77 death Nl! Fr bullet/l 5 A series number G' and waste tank 2・m
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Claims (1)

[Claims] It is formed by a ring box-shaped casing that is attached to the crankshaft and rotates together with the crankshaft, and an annular inertial body that is housed within the casing with a predetermined gap between the inner wall of the casing and the above-mentioned casing. In a viscous damper made by filling the gap between the inertial body and the viscous liquid,
The viscous liquid is a high viscosity liquid having viscoelasticity, and the ratio of the natural frequency fd (fd/fe) of the viscous damper system to the natural frequency fe of the crankshaft system is 0.5.
A viscous damper characterized by the above.