JPH0735842B2 - Pendulum type vibration control device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial field of use] The present invention is a pendulum type damping device for expanding the damping force of a weight at the tip of a pendulum and a damper connected to the pendulum by lever action to enhance the damping effect. A pendulum type damping system that always obtains a good damping effect by changing the lever ratio of the pendulum, even if the natural frequency of the vibration system to which the pendulum is attached, particularly the vibration system, changes. Regarding the shaking device.

[Technical background of the related art and its problems] Many of the vibration problems that occur in mechanical systems are due to insufficient damping of mechanical structures. Therefore, various vibration dampers have been proposed as a method of externally applying damping to a mechanical structure. This vibration suppressor can be roughly classified into a passive type and an active type.

The active type generally provides a large damping effect, and in particular, the damping effect does not significantly decrease with respect to parameter fluctuations of the main system that is the object of damping (for example, changes in natural frequency due to changes in rigidity and mass). have. However, it has drawbacks such as a complicated structure, it is necessary to supply energy from the outside, and maintenance is difficult. On the other hand, the passive type generally has a simple structure and does not require the supply of energy from the outside, so that the maintenance is easy and the maintenance cost is low.

However, the passive vibration suppressor such as the dynamic vibration reducer and the hood damper is subjected to the optimum adjustment method under the condition that the natural frequency of the main vibration system to be damped is kept constant. Therefore, if the natural frequency of the main vibration system fluctuates from the design value, the vibration damping effect of the vibration damper, which was well adjusted in advance, would be greatly impaired.

Therefore, the emergence of a quasi-active type vibration suppressor that has both passive and active characteristics is awaited.

[Object of the Invention] The present invention was devised to solve the above-mentioned problems of the prior art.

A first object of the present invention is a pendulum that is basically a passive type vibration damping device and can cope with the fluctuation of the natural frequency of the vibration system to be dampened and can always maintain a good vibration damping effect. Type vibration control device.

A second object of the present invention is to provide a self-tuning pendulum type vibration damping device capable of automatically adjusting to optimum vibration damping characteristics in response to natural vibration fluctuation of a vibration system which is a vibration damping target. Especially.

[Summary of the Invention] In a first invention, a pendulum type vibration damping device is provided with an adjusting structure capable of changing a lever ratio of a pendulum, and a damping characteristic can be changed by operating a lever ratio. . Here, the pendulum type vibration damping device is a device for trying to effectively obtain a vibration damping effect by expanding the dynamic acting force of the weight damping attached to the tip of the pendulum and the damper by the lever action. , A large damping effect can be obtained by giving a small weight and a slight damping force. In the present invention, focusing on the fact that the damping effect provided by this pendulum type damping device changes by changing the lever expansion ratio, the quasi-active that positively utilizes it to optimally control the lever ratio. It is intended to build a mold damping device.

According to a second aspect of the present invention, an operation for operating a detector for detecting the natural frequency of the vibration system to be dampened and the adjusting mechanism for moving and adjusting the fulcrum of the pendulum according to the detection signal. Means are provided so that optimum adjustment having a high vibration damping effect is automatically performed.

Embodiments Embodiments of the present invention will be described in detail below with reference to the accompanying drawings.

FIG. 1 shows the basic structure of a pendulum type vibration damping device of the present invention. As shown in the figure, the main vibration system 1 to be damped is provided on the foundation 2 and is vibrated with the vibration of the foundation 2. A pendulum 4 is suspended from each upper part of the main vibration system 1 via a bearing 3 so that the pendulum 4 can vibrate.
A weight 5 is attached to the lower end of the pendulum 4, and a damper 6 for damping the vibration is connected to the weight 5. The damper 6 is fixed to the foundation 2, and the pendulum 4 is attached to the foundation 2 surface.
Gives damping force to. In addition, a support structure 7 is provided on the foundation 2 for supporting the pendulum 4 in the middle thereof and setting its own vibration characteristics so that a predetermined elastic force acts on the fulcrum of the pendulum 4 when the pendulum 4 swings. There is. An adjusting mechanism 8 for supporting the pendulum 4 and moving the pendulum 4 up and down to adjust the fulcrum of the pendulum 4 is provided above the support structure 7.

The adjusting mechanism 8 is provided so as to surround the rod 9 of the pendulum 4 and has a supporting portion 10a for supporting the rod 9 on the inner circumference thereof. The slider 10 is screwed on the outer circumference, and is meshed with the screw on the outer circumference of the slider 10. It is mainly composed of an annular gear 11 rotatably attached to the support structure 7. When the gear 11 rotates forward and backward, the slider 10 moves up and down along the rod 9, and the fulcrum of the pendulum 4 supported by the support portion 10a changes. Gear 11
The servomotor 12 as an operating means for driving this
Are connected. The shaft of the servomotor 12 has a gear
11 A gear 13 that meshes with the outer peripheral teeth is attached. Further, the main vibration system 1 is provided with a detector (not shown) for detecting its natural frequency, the servo motor 12 is rotationally driven based on the detection signal of the detector, and the slider is provided via gears 13 and 11. It is configured such that the lever ratio of the pendulum 4 changes by moving up and down.

The change in the natural frequency of the main vibration system 1 can be considered to be due to the structure of the main vibration system 1 or due to secular change.
When it is caused by the structure, it is often possible to infer the change in the natural frequency by measuring some physical quantity. on the other hand,
The change due to aging is due to, for example, wear or corrosion of the structure or looseness of the joint, which cannot be known unless the natural frequency is directly measured. . In any case, the change in the natural frequency of the main vibration system 1 can be known in some way. In the following, it is assumed that the main vibration system 1 changes its natural frequency due to a change in mass or rigidity.

Next, with respect to changes in the natural frequency of the main vibration system 1, the pendulum 4
The optimum adjustment condition is clarified by how to change the fulcrum of to maintain the good damping effect.

First, in analyzing the apparatus shown in FIG. 1, consider a dynamic model as shown in FIG. The main vibration system 1 and the support structure 7 are represented by an undamped one-degree-of-freedom system having a mass M, m and a spring constant K, k, respectively. The pendulum 4 is assumed to be a rigid body, and the distance between the upper end of the pendulum 4 and the fulcrum is l ₁ , and the distance from the fulcrum to the weight 5 of the pendulum 4 is l ₂ . The mass of the weight 5 of the pendulum 4 at the lower end of the pendulum 4 is m _p , and the viscous damping coefficient given to the weight 5 is C. In this analysis, the response such as the mass displacement X ₁ of the main vibration system 1, the mass displacement x ₂ of the supporting structure 7, the displacement X ₃ of the weight 5 of the pendulum 4 and the like to the horizontal vibration u of the foundation 2 is obtained.

Based on the dynamic model shown in FIG. 2, a motion equation regarding mass point is established. _{x 1 / u, x 2 /} u, Expressed as a frequency transfer function between the transfer functions of _{x 3 / u x 1, x} 2, complex amplitude ₁ x _{3, 2, 3} and the base amplitude U, following become that way.

Where the oscillation equation △ is Since the object of vibration suppression here is the displacement of the main vibration system 1, it suffices to find a condition for minimizing the amplitude magnification of the equation (1). Therefore, the following terms are defined in order to have generality by making the equation dimensionless.

Using these terms, the amplitude scale factor X ₁ / U in Eq. (1) and its phase φ
Is represented as follows. However, the term of numerator ω ³ is ignored.

here Other amplitude magnifications X ₂ / U and X ₃ / U are also made dimensionless by using the above 6 terms. Therefore, the six parameters of Ωn, ωn, μs, u, ζ, and λ are intervening factors for damping design. In the present invention, in particular, the lever ratio λ is used as an adjustable variable.
The value of λ that can be adapted to the change in natural frequency of the main vibration system will be investigated.

The optimum adjustment method of this device is obtained by applying the fixed point theory used for a dynamic vibration absorber. According to it, the intersection points P and Q of the amplitude multiplication curve of the equation (5) given to the damping rates ζ = 0 and ζ = ∞ are 2
Since there are two fixed points, the optimum tuning condition is obtained from the condition that the heights of the P and Q points are equal. Also, the best damping is
It is obtained as a damping value such that the vicinity of points P and Q is the maximum value of the amplitude multiplication curve. The optimum tuning condition is satisfied by the following equation.

2EAF = D (BF + D-EC) (12) Substituting equations (7) to (11) into this equation and rearranging for μ gives the following equation.

Optimum tuning conditions are satisfied if the relationships among the parameters are determined so as to satisfy this equation.

On the other hand, for the best attenuation, the square of Eq. (5) is differentiated by ω and ∂ (X ₁ / U) ² / ∂ω ² = 0 (14) P, Q frequency ω _P and ω _Q are substituted. them if, ω _P, ω _Q to the corresponding attenuation factor ζ _P, ζ _Q is obtained. The best damping ζ _opt may use the average value of both.

ζ _opt = (ζ _P + ζ _Q ) / 2 (15) When optimal tuning is performed by equation (13) and the best damping is given by equation (15), the maximum amplitude magnification of the main vibration system is To be suppressed.

Optimal tuning and damping diagrams prepared for the example of λ = 5 by using the equations (13), (15), and (16) are shown in FIGS. 3 and 4. In each figure, the horizontal axis shows the natural frequency ratio Ωn / ωn, and the parameter shows the mass ratio μs.

In FIG. 3, by designating Ωn, ωn, μs, the mass ratio μ and the damping ratio ζ required for optimum tuning can be read. In Fig. 4, the maximum amplitude magnification and ω _P ,
The value of ω _Q is read. That is, it is possible to know the outline of the response magnification curve of the main vibration system with the vibration damping device that is optimally adjusted from this figure.

Next, we will examine the conditions for maintaining optimum tuning by changing the lever ratio with respect to the fluctuation of the natural frequency of the main vibration system.
The condition (13) for optimum tuning can be rearranged into a polynomial for the lever ratio λ as follows.

b ₂ c ₄ λ ⁶ + (b ₂ c ₃ + b ₁ c ₄ ) λ ⁵ + (b ₂ c ₂ + b ₁ c ₃ + b ₀ c ₄ -a ₄ ) λ ⁴ + (b ₂ c ₁ + b ₁ c ₂ + b ₀ c ₃ -a ₃ ) λ ³ + (b ₂ c ₀ + b ₁ c ₁ + b ₀ c ₂ -a ₂ ) λ ² + (b ₁ c ₀ + b ₀ c ₁ + a ₁ ) Λ + (b ₀ c ₀ -a ₀ ) = 0 (17) Here, if α = Ωn / ωn is set, each coefficient is as follows: a ₄ = (1-a ² ) (2a ² / μs ^{2 + 1 / μs-1)} a 3 = (1-a 2) (8a 2 / μs 2 + 2 / μs) a 2 = (1-a 2) (12a 2 / μs 2 + 1 / μs) a 1 = (1 -A ² ) (8a ² / μs ² ) a ₀ = (1-a ² ) (2a ² / μs) b = μ (1 + 1 / μ) b = 2μ / μ b = μ / μ c = (2a / μ + 4 / μ + 2) c = (1 + 8a / μ + 9a / μ) …… (18) c = (12a / μ + 6a / μ) c = (8a / μ + a / μ) c = Za / μ Equation (17) is a sixth-order polynomial for λ Is. Therefore,
Although there are 6 λs, if this polynomial is solved, there is only 1 positive real root. The real root is the value of the lever ratio that gives the optimum tuning.

As described above, the natural frequency fluctuation of the main vibration system may be due to mass change or spring constant change.
First, consider the case where the spring constant K changes. In this case, since μ and μs are constant values, it is only necessary to change α in Equation (17) to obtain the value of λ. α, μ, μ
When s is given and the value of λ is determined, these values are substituted into the equation (15) and the best attenuation is also determined.

FIG. 5 illustrates an example of the calculation result of the optimum values of λ and ζ due to the fluctuation of the natural frequency. In this figure, α = Ωn /
The optimum values of λ and ζ can be read from the vertical axis when Ωn fluctuates in the range of 0.5 to 1.5 around ωn = 1/3. Maximum response magnification X ₁ / U given optimal values of λ and ζ
| max can also be found on the vertical axis. The solid line is μs = 0.1, μ
= 0.01, the broken line is μs = 0.044, μ = 0.0044
It was drawn about. From this figure, it is understood that a substantially linear relationship is established between the increase / decrease in Ωn and the optimum lever ratio. On the other hand, the best damping rate ζ seems to change significantly with respect to the variation of Ωn. However, from the defined parameters, the viscous damping coefficient C is as follows, and ζ decreases as C = 2ζMΩn …… (19) Ωn increases. Therefore, the value of C is increased for the fluctuation of Ωn in this range. No need to change.
Practically, it suffices to change the lever ratio, and there is no problem in keeping C at a constant value.

Next, when the mass M changes, μs and μ also change. Therefore, in this case, μs and μ must be expressed as a function of α. That is, μs = α ² · (k / K) (20) μ = μs · (mp / m) (21) Here, since the spring constant ratio k / K and the mass ratio m _p / m are constant values, μs and μ are uniquely determined by the value of α. After the above preparations, the relationship between α and λ, ζ can be similarly known in this case as well by the equations (17) and (15).

Similarly to FIG. 5, FIG. 6 also shows the optimum values of λ and ζ when Ωn fluctuates in the range of 0.5 to 1.5 around α = Ωn / ωn = 1/3. In this figure, the actual battle is k / K = 0.9, m
For p / m = 0.1, dashed lines are drawn for k / K = 0.4, mp / m = 0.1. These values are set so that the same conditions as in FIG. 5 are obtained when Ωn = 1. (Hereinafter, the unit rad / s of Ωn and ωn is omitted).

Compared with FIG. 5, the differences in FIG. 6 are that the change of the lever ratio is large with respect to the variation of α, that there is no linear relationship, that the damping rate decreases as α decreases, and that the maximum amplitude The point where the change in magnification is small is raised. In this case as well, the viscous damping coefficient C is expressed as follows.

C = 2 [zeta] K / [Omega] n (22) Therefore, also in this case, the fluctuation range of [zeta] as seen in FIG.

Next, a calculation example of the amplitude magnification curve will be described.

First, after replacing equations (1) to (3) with the six defined parameters, x ₁ , x ₂ and x ₃ using the prepared chart
An example of numerical calculation of a point amplitude magnification curve is shown.

If Ωn = 1, ωn = 3, ζ = 5, μs = 0.1, the values of μ = 0.01 and ζ = 0.0055 can be read from FIG. 3, and X ₁ /U=2.38,ω from FIG. _P = 0.94, the value of ω _Q = 1.47 is found.
FIG. 7 is an amplitude magnification curve of each part drawn by using these equations by substituting the above six parameters into equations (1) to (3). As can be seen, the amplitude multiplication curve of the main vibration system agrees with the result obtained from FIG. 4, and it can be seen that the best vibration damping effect is obtained. The maximum amplitude magnification of the support structure is about twice that of the main vibration system, and that of the pendulum weight is about eight times.
This value increases in proportion to the increase of the lever ratio. Therefore, if the maximum amplitude of the pendulum is restricted, the value of λ cannot be increased.

As an example of numerical calculation, Ωn = 1, ω = 3, μs = 0.1 (k / K = 0.
9), μ = 0.01 (m / m = 0.1), λ = 5, ζ = 0.0055 (C =
Consider a case where the main vibration system with a vibration damping device optimally adjusted to 2ζMΩn = 0.011M) fluctuates within the range of Ωn = 0.6 to 1.4. First, FIG. 8 shows a case where the spring constant K changes.
In the figure, the natural frequency Ωn varies depending on the fluctuation of the spring constant K.
Shows the transition of the amplitude magnification curve when changes from 0.6 (curve a), 0.8 (curve b), 1 (curve c), 1.2 (curve d), 1.4 (curve e).

In FIG. 10 (a), when λ = 5 and C = 0.011M (optimal value at Ω = 1) are kept constant, FIG. 10 (b) shows that for each Ωn according to the relationship shown in FIG. When optimum values λ _opt and ζ _{opt of} ζ and λ are given, (c) shows the value of ζ as C =
The case where the optimum value λ _opt is given only to λ while keeping 0.011M is shown.

From this figure, it can be seen that the lever ratio needs to be adjusted especially when the natural frequency becomes higher than that during the optimum adjustment. Also,
It will also be understood that the damping value may remain constant if there is a lever ratio adjustable mechanism that can realize the relationship shown in FIG.

FIG. 9 shows the case where the mass M changes. FIG. 9 shows an amplitude magnification curve when the natural frequency Ωn changes in the range of 0.6 to 1.4 due to the change of the mass M. (A), (b) and (c) of this figure are drawn under the same conditions as in FIG. As is clear from a comparison with FIG. 8, when the mass change occurs, damping of damping effect can be seen regardless of whether the natural frequency is higher or lower than that during optimum adjustment, and the lever ratio is optimized. It is characterized in that if it is adjusted, a vibration damping effect superior to the case where the spring constant K changes is obtained. Therefore, it can be said that the effect of the vibration damping device is more exerted when the mass changes. In this case as well, the mechanism for adjusting the attenuation value is considered to be unnecessary in practice.

Next, the experimental results for the apparatus of FIG. 1 will be shown.

The specifications of the experimental device having the structure shown in FIG. 1 are as follows.

M = 8.08Kg K = 9.42 × 10 ⁴ N / m m = 0.44Kg k = 2.92 × 10 ⁴ N / m m _P = 0.045Kg The natural frequency was changed by the addition of mass. A weight of 4 kg was used as the additional mass. As a result, the natural frequency of the main vibration system without added mass is 108 rad / s, that with added mass is 88.3 rad / s, that of the supporting structure is 257.6 rad / s, and the former lever ratio and damping ratio are λ = 4.72, ζ = 0.0018, and the latter is λ = 6.2, ζ = 0.0012, respectively.

In this device, the magnetic damper was adjusted so that ζ = 0.0016 when there was no additional mass.

FIG. 10 shows two types of amplitude magnification curves obtained by vibrating the bench. ● Lever ratio is λ = 4.85 without added weight
When adjusted to, the ○ mark shows the case where added mass is added and adjusted to λ = 6.1. It can be seen that both agree well with the theoretical value (solid line) and are well damped. These damping effects are clearly understood by impulse response experiments. FIG. 11 shows a comparison of impulse responses with and without a damping device in the absence of added mass.
The figure shows a comparison with additional mass.

Next, an embodiment in which the vibration damping device of FIG. 1 is specifically applied will be described. The following embodiment is an example in which the natural frequency of the main vibration system that is the damping target changes due to its structure.

FIG. 13 shows an example applied to a spherical tank for liquid storage. The spherical tank 20 is supported by a large number of pillars 21 provided along the circumferential direction. Further, a gate-shaped support structure 22 is erected on the outside of the spherical tank 20. The upper end of the pendulum 4 is oscillatably suspended between the support structure 7 and the spherical tank 20. The fulcrum 23 of the pendulum 4 is movable up and down, and the weight 6 at the lower end of the pendulum 4 has a damper 6
Are connected. The natural frequency changes depending on the amount of liquid stored in the spherical tank 20. Therefore, the amount of liquid is detected by the detector to obtain the natural frequency, and the fulcrum 23 of the pendulum 4 is moved so that the optimum lever ratio is obtained according to the magnitude of the natural frequency. Thereby, a good vibration damping effect is always obtained, and the seismic resistance of the spherical tank 20 can be improved.

Fig. 14 shows an example applied to boring machines and ram structures of portal machine tools. Inside the ram structure 24, there is provided a pendulum 4 having a base end vibratably attached to the blade end at the lower end and having a weight 5 at the upper end. A fulcrum 26 of the pendulum 4 is supported by a cylindrical support structure 25, and the fulcrum 26 can be moved up and down by an adjusting mechanism 27. The weight 5 is inserted in a bag 29 made of rubber or the like containing the liquid 28,
The liquid 28 functions as a damper. In this machine tool, the natural frequency changes as the ram structure 24 moves up and down, so the lever ratio of the pendulum 4 is adjusted based on the vertical position of the ram structure 24. In this example, crack vibration can be prevented.

FIG. 15 shows an example applied to an articulated robot arm. A pendulum 4 is provided in the arm 30 along its axial direction. A fulcrum of the pendulum 4 is supported by a cylindrical support structure 31, and the support portion has a structure that allows sliding and vibration by, for example, a linear bearing and a ball bearing. The weight 5 at the tip of the pendulum 4 is inserted into the damping liquid 28 in the support structure 31. The liquid 28 is liquid-tightly sealed by a rubber seal 32 provided between the rod of the pendulum 4 and the inner wall surface of the support structure 31. In multi-joint robots, the number of arms has recently increased, and accompanying this, vibrations of the arms themselves have been regarded as a problem that impairs positioning accuracy. Then, the natural frequency of the first arm changes depending on the posture before the second arm. Therefore, for example, when damping the first arm, the postures of the second arm and the arms beyond it are detected, and the pendulum 4 is detected according to the state.
Move and adjust the fulcrum of.

[Effects of the Invention] In summary, according to the present invention, the following excellent effects are exhibited.

(1) A change in the natural frequency of the vibration system to be damped can be dealt with by changing the lever ratio of the pendulum, and a good vibration damping effect can always be maintained.

(2) Although it has an adjusting mechanism to adjust the lever ratio, it is basically a passive vibration suppressor, and can adapt to natural frequency fluctuations only by changing the lever ratio of the pendulum, simplifying the structure and facilitating maintenance management. And maintenance costs can be reduced.

(3) Since the natural frequency of the vibration system is detected by the detector and the lever ratio of the pendulum is changed by the actuating means based on the detection symbol, a good damping effect is always obtained.

(4) Between the natural frequency of the vibration system and the optimum lever ratio,
Since there is a substantially proportional relationship, automation by means of detectors and operating means can be easily implemented.

[Brief description of drawings]

FIG. 1 is a configuration diagram showing an embodiment of a vibration damping device according to the present invention attached to a main vibration system to be damped, FIG. 2 is a diagram showing a dynamic model of the device of FIG. 1, and FIG. Figures 9 to 9 are vibration analyzes for the dynamic model of Figure 2, Figure 3 shows the optimum tuning and damping, Figure 4 shows the damping effect at the time of optimum adjustment, and Figure 5 shows FIG. 6 is a diagram showing an adaptive condition due to a change in spring constant of the main vibration system, FIG. 6 is a diagram showing an adaptive condition due to a mass change of the main vibration system, FIG. 7 is a diagram showing an amplitude magnification curve of each part, and FIG. 8 is a main vibration. FIG. 9 is a diagram showing an amplitude magnification curve due to a change in the spring constant of the system, FIG. 9 is a diagram showing an amplitude magnification curve due to a change in the mass of the main vibration system, and FIG. 10 is a view comparing experimental values and theoretical values depending on the amplitude magnification. Figures 11 and 12 show the results of impulse response experiments with and without a damping system.
13 is a front view showing an example in which the vibration damping device of the present invention is applied to a spherical tank, FIG. 14 is a vertical sectional view showing an example in which the vibration damping device of the present invention is applied to a boring board, and FIG. It is a longitudinal cross-sectional view showing an example in which the vibration damping device of the present invention is applied to a joint robot arm. In the figure, 1 is a main vibration system, 2 is a base, 3 is a bearing, 4 is a pendulum, 5 is a weight, 6 is a damper, 7 is a support structure, 8 is an adjusting mechanism, 9 is a rod, 10 is a slider, and 11 is a slider. Gear, 12 is a servo motor, 13 is a gear, 20 is a spherical tank, 21 is a pedestal, 22
Is a support structure, 23 is a fulcrum, 24 is a ram structure, 25 is a support structure, 26 is a fulcrum, 27 is an adjusting mechanism, 28 is a liquid, 29 is a bag, 30
Is an arm, 31 is a support structure, and 32 is a rubber seal.

Claims (2)

[Claims] 1. A pendulum including a rod rotatably attached to a vibration system whose base end is to be damped and a weight attached to a tip as a free end of the rod, and a pendulum weight for suppressing vibration of the pendulum. A support structure for supporting the rod in the middle and forming a fulcrum of the rod when the pendulum swings, and a damper provided on the support structure and moving itself along the axial direction of the rod. An adjusting mechanism for moving and adjusting the fulcrum of the rod by means of the slider, wherein the adjusting mechanism has a slider provided around the rod of the pendulum and having a support portion for supporting the rod on the inner circumference, and a screw on the outer circumference, It is characterized by comprising a gear that is rotatably attached to a support structure and is formed in an annular shape to surround the slider, and has an inner periphery formed with a screw that meshes with a screw on the outer periphery of the slider. Pendulum vibration damping device to. 2. A pendulum including a rod rotatably attached to a vibration system whose base end is to be damped, and a weight attached to a tip as a free end of the rod, and a pendulum weight for suppressing vibration of the pendulum. A support structure for supporting the rod in the middle and forming a fulcrum of the rod when the pendulum swings, and a damper provided on the support structure and moving itself along the axial direction of the rod. By doing so, the adjusting mechanism for moving and adjusting the fulcrum of the rod, the detector for detecting the natural frequency of the vibration system, and the adjusting mechanism for moving and adjusting the fulcrum of the rod based on the detection signal from the detector are operated. The adjusting mechanism includes an operating means, and the adjusting mechanism is provided to surround the rod of the pendulum, has a supporting portion for supporting the rod on the inner circumference, and has a slider screwed on the outer circumference; A gear that is attached to the outer periphery of the gear and that is formed in an annular shape to surround the slider and that has a screw that meshes with a screw on the outer periphery of the slider and that has teeth formed on the outer periphery of the gear. A pendulum type vibration damping device comprising a gear that meshes with teeth to rotate the gear to move the slider.