JP2006258189A - Variable rigidity type dynamic vibration absorbing device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a variable rigidity type dynamic vibration absorbing device, capable of solving the problem of a pattern formation phenomenon if any generated due to vibration of a contact roll system, for example, to which there is no fundamental measure of prevention at site, giving fatal damage to the product, or causing a situation of setting an upper limit of line speed. <P>SOLUTION: This variable rigidity type dynamic vibration absorbing device is composed of a pair of same rigidity long rigid members having axial centers on the same perpendicular plane to the vibration direction of a vibration body, and having the longer axis and the shorter axis in a cross sectional surface, a vibration absorbing body installed on the vibration member to be rotatable on both sides of the pair of the long rigid members, a weight installed to be rotatable between the pair of the long rigid members, and a mechanism to vary a synthetic rigidity coefficient while holding the synthetic vibration direction of the pair of the long rigid members set to the linear vibration direction of the vibration body by mutually reversely rotating the pair of the long rigid members in a surface-symmetric state. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a variable stiffness dynamic vibration absorber for translational vibration.

In the industry, regular patterns are gradually formed on the roll and the system in contact with the roll with the operation of the contact rotating system, which also induces severe vibration or is transferred to the product. There are many phenomena that result in defective products. For example, railroad vehicles and rails, drive rolls and bobbin holders via yarn balls of textile machinery winders, a pair of paper machine rubber winding rolls, automobile tires and roads, steelmaking machines, machine tools, etc. have specific patterns. It is formed. These phenomena are called pattern formation phenomena, and the mechanism of their occurrence has been elucidated and countermeasures have been studied.

In particular, in steelmaking machines, chatter marks for tension levelers and chattering during hot / cold rolling have become serious problems in product management as the accuracy of products increases. On the other hand, the roll polygonalization phenomenon of smoother rolls and gate roll size presses in the press part of the paper machine is also the main cause that hinders the improvement of the line speed.
Most of these phenomena are pattern formation phenomena that occur in a contact roll system in which the roll rotates in contact with the roll.

From the viewpoint of global research, this kind of contact rotation system pattern formation phenomenon is mainly researched on chattering of machine tools and corrugation of railway rails in Japan. The research on the pattern formation phenomenon of the contact roll system in which the roll rotates is in the developing stage. In particular, the chattering phenomenon that occurs in steelmaking machines such as hot rolling and cold rolling, the polygonal deformation phenomenon of rolls in paper machines, and the abnormal vibration in the winding process of thin strips of paper, etc., still have a clear mechanism. It has not been elucidated. Therefore, as a countermeasure in the field against the pattern formation phenomenon that occurs in such a contact roll system, only the line speed is slowed down or the rolls are replaced early, and the fundamental problem is not solved.
When this phenomenon occurs, the industry has to deal with fatal damage or set an upper limit of the line speed. Therefore, the industry is urgently required to develop a countermeasure.

However, such preventive measures have not yet been studied.
In previous studies, as measures to prevent the pattern formation phenomenon, we studied methods for reducing the instability by optimizing the roll diameter ratio and measures for delaying pattern formation that fluctuates the roll rotation speed, and proposed design guidelines. I have done it. The diameter ratio optimization method is effective when the paired rolls are deformed equally and objectively, but is not effective when one of the rolls is deformed. Moreover, since there are many cases where it is necessary to operate at a constant rotation speed, such as an iron-making machine or a papermaking machine, the types of machines that can take countermeasures are limited.
In addition, we theoretically analyzed how the pattern formation phenomenon is affected by attaching a passive dynamic vibration absorber. As a result, unlike the vibration damping effect of a dynamic vibration absorber for general self-excited vibration, there is a problem that a drastic vibration damping cannot be expected and a vibration mode to be damped appears newly.
Further, in a winder system of a textile machine, a winder of a thin ribbon such as paper, etc., the natural frequency of the system changes every moment, so vibration suppression is impossible with a passive dynamic vibration absorber.

Therefore, the present invention provides a variable stiffness type dynamic vibration absorber that solves the above-mentioned problems, and features thereof are as follows.
(1). A pair of identical rigid long rigid bodies having respective axial centers located on the same orthogonal plane with respect to the vibration direction of the vibrating body and having a long axis and a short axis in the cross section, and the pair of long rigid bodies attached to the vibrating body A vibration-absorbing body that is rotatably mounted on both sides thereof, a weight that is rotatably mounted at the center of the pair of long rigid bodies, and the pair of long rigid bodies that are rotated in opposite directions in a plane-symmetrical state. And a mechanism for changing the combined stiffness coefficient while maintaining the combined vibration direction of the pair of long rigid bodies in the linear vibration direction of the vibrating body.
(2). Two pairs of the same large rigid long rigid body, the same small rigid long rigid body, the same small rigid long rigid body, and the vibrating body, each axis centering on the same orthogonal plane with respect to the vibration direction of the vibrating body and having a long axis and a short axis in the cross section A vibration-absorbing body that is mounted on both sides of the two pairs of long rigid bodies so as to be rotatable, and a weight that is rotatably mounted on the center of the two pairs of rigid bodies, and has the same large rigidity in parallel relation. The same small rigid long rigid body pair is disposed outside or inside the long rigid body pair, and one side and the other side of the two pairs of long rigid bodies are rotated in opposite directions in pairs with each other in a plane symmetrical state. And a mechanism for changing the combined stiffness coefficient while maintaining the combined vibration direction of the two pairs of long rigid bodies in the linear vibration direction of the vibrating body.

That is, the variable stiffness type dynamic vibration absorber of the present invention is attached to a vibrating body, and both sides of each of the long rigid bodies are rotatably mounted by a vibration absorbing body, and a weight can be rotated at the center of the long rigid body. As a result, the rotation angle of the long rigid body having anisotropy in the stiffness coefficient can be reliably and accurately compared to the so-called single-support type in which a weight is rotatably mounted on one side of the long rigid body. It can be changed in a symmetric state or in an exact parallel relationship. When the single-support type is applied to the vibration in the translational direction, the vibration in the rotational direction is coupled, and it is difficult to obtain effective vibration suppression particularly in the forced vibration system. As a result, the composite stiffness coefficient of the pair of long rigid bodies is continuously changed to accurately match the natural frequency of the vibration absorption body side with the natural frequency of the vibrator, thereby suppressing or reducing the vibration of the long vibrator. It is delayed.
In addition, by combining two pairs of long rigid bodies and making one pair as the same large rigid long rigid body for basic adjustment and the other pair as the same small rigid long rigid body for fine adjustment, synthesis of the long rigid body The rigidity coefficient can be set to a more accurate value.
That is, the pair of large rigid long rigid bodies and the pair of small rigid long rigid bodies are, for example, coarse adjustment rigid bodies having a large difference in width and height (large anisotropy) of the rectangular cross section. A rigid body for fine adjustment having a small anisotropy is referred to as a small rigid long rigid body.
In terms of structure, the rigidity coefficient changing mechanism for the pair of long rigid bodies can be either a manual type or an automatic type and is simple.
In addition, the rotation angle of a pair of long rigid bodies can be adjusted in a plane symmetry or in a parallel relationship, and the natural frequency on the vibration absorption body side can be adjusted to the natural vibration of the vibration body. It is possible to stop, reduce, and delay control.
The present invention, as will be introduced next, has many examples of application to iron making machines, paper making machines, and the like, and has extremely beneficial effects industrially. In the case where the roll rotates like an iron making machine or a paper making machine, the following method can be considered in addition to using the vibration absorbing device mounted on the roll bearing portion.
a). In a cylindrical roll, a bearing such as a slide bearing can be provided via a central axis, and the bearing can be arranged via the bearing.
b). A backup roll is attached to the roll surface, and the vertical vibration attached to the backup roll bearing is absorbed.
In these examples, the vibration damping effect can be obtained even in the vibration mode in which the amplitude of the roll bearing portion contacting the roll surface is small.
c). It is applied as a corrugation for rails and a polygonal prevention device for wheels.
As the train repeats running, the surface of the rail and the wheels are deformed into a polygon (corrugation), and noise, a decrease in the service life of the rail, and driving safety problems occur. Corrugation is prevented by attaching a vibration absorber to the axle portion through a bearing and suppressing the vertical vibration of the wheel.
d). Apply to machine tools.
In order to suppress chatter vibration of the machine tool, a vibration absorber is mounted on the tool.
e). Applied to automobile brake squeal prevention device.
The disc brake squeal of an automobile is self-excited vibration caused by the combination of a rotor and a pad or a rotor and a caliper. The latter, which is a particular problem, can be reliably suppressed by attaching a dynamic vibration absorber to the caliper. However, when a vibration absorber tuned by a bench test is attached to an actual vehicle, the effect is often not obtained due to the difference in chassis structure. Since this vibration absorber can be easily tuned, it is also effective in the brake squeal phenomenon of an actual vehicle. Possible mounting locations include calipers, rotor interiors (in which case a plurality are required), and hydraulic piston interiors. In addition, the former squeal is attached to the pad.
f). Applied to automobile suspension.
By mounting this device on the suspension upper part of the automobile, it is possible to minimize the vibration to the vehicle body due to the forced vibration of the road surface.
g). Other applications include building vibration suppression due to earthquakes, electric wire self-excited vibration (galloping) suppression, and train squeal noise prevention.

In the variable stiffness type dynamic vibration damping device of the present invention, the shape of the long rigid body is the same for each pair, and the combined vibration direction received by the weight from the pair of rigid bodies is always the same direction as the vibration direction of the vibrating body. In addition, a cross-sectional shape having a major axis and a minor axis such as a rectangle or an ellipse whose composite stiffness coefficient continuously changes by changing the plane symmetry state or the parallel state of the rotation angle is formed.
Further, as the rigidity coefficient changing mechanism for the long rigid body, a single or shared long rigid body rotating device is used. For example, an example of a gear-type rotating device as introduced in Example 1 and Example 2 described later is a pair of long rigid bodies so that the pair of long rigid bodies rotate in plane symmetry simultaneously with their rotational directions reversed from each other. A pair of main gears of the same size are fitted to each other at a reference position where the long axis of each long rigid body is parallel to the vibration direction, or between the gears. In this case, an intermediate transmission gear is interposed between the main gear and one of the main gears is rotationally driven by a motor drive mechanism.
Next, modifications of the present invention will be listed.
(I). The variable stiffness type dynamic vibration absorber having a plurality of pairs of long rigid bodies and changing the stiffness coefficient between the pairs.
For example, a pair of two long rigid bodies are provided for rough frequency adjustment, and another pair of two long rigid bodies are provided for fine frequency adjustment.
(B). The variable stiffness type dynamic vibration damping device in which a viscous fluid is sealed in the weight so that the damping effect can be controlled.
(C). The variable stiffness type dynamic vibration absorber, wherein the vibration absorbing body is mounted on a vibrating body so as to be three-dimensionally rotatable, and the vibration absorbing body is rotated in a direction corresponding to vibration of an external force.
(D). The variable rigidity type dynamic vibration absorber capable of moving a tuning range by inserting an elastic rod between the weight and the vibration absorbing body.
(E). A control method of the natural frequency of the vibration absorbing body that performs tuning based on the phase difference between the vibration of the vibrating body and the weight as well as the amplitude of the vibrating body.
(F). Hybrid variable stiffness dynamic vibration absorber.
By connecting multiple pairs of long rigid bodies and the weights in parallel or in series, and expanding the vibration absorber from a one-degree-of-freedom vibration system to a multi-degree-of-freedom vibration system, a plurality of vibration frequencies can be simultaneously controlled by a single vibration absorber. A hybrid variable rigidity dynamic vibration absorber that can control vibration.
(G). A variable stiffness type dynamic vibration absorber in which a rigid body rotating motor of a stiffness coefficient changing mechanism is miniaturized in place of the weight portion.
(Chi). A variable-rigidity dynamic vibration absorber that can control both viscosity and rigidity by using MR fluid or the like in the vibration absorber body or the operating part of the stiffness coefficient changing mechanism.

Next, an embodiment (specific example) of the present invention (1) will be described with reference to FIGS.
FIG. 1 is an explanatory view showing an embodiment of the device of the present invention, and is a plan sectional view taken along line AA in FIG. FIG. 2 is a side cross-sectional view from the arrow BB in FIG. FIG. 3 is a front view from the arrow CC of FIG.
1 to 3, a vibration absorbing body 2 is attached to the vibrating body 1, and both sides of a pair of leaf springs (long rigid bodies) 3, 4 having the same rigidity coefficient are rotatably mounted on the vibration absorbing body 2. A spring stiffness coefficient changing mechanism 6 is attached, and a weight 5 is rotatably mounted in the middle of the leaf springs 3 and 4.
The lower surface of the vibration absorbing body 2 is attached to the vibrating body 1 as a plane H1 orthogonal to the vibration direction V of the vibrating body 1.
The leaf springs 3 and 4 have the same rectangular shape (a kind of rectangle) as shown in FIG. 3, and their axial cores C 1 and C 2 are parallel to a plane H 2 orthogonal to the vibration direction V of the vibrating body 1. The cross-sectional major axes L1 and L2 are aligned in parallel with the vibration direction V at the reference position.
The two sides of the leaf springs 3 and 4 and the vibration absorbing body 2 are mounted by forming the both side portions of the leaf springs 3 and 4 on the rotating shafts 3a and 4a, and providing the bearing body 2a and 2b with the bearing body 2a. Type bearings 2a, 2b with rotating shafts 3a, 4a attached. The center part of the leaf springs 3, 4 and the weight 5 are attached to the center part of the leaf springs 3, 4 on the rotating shafts 3b, 4b. The bearing 5 is formed with bearings 5a and 5b, and the bearings 5a and 5b are equipped with the rotary shafts 3b and 4b.
The leaf spring stiffness coefficient changing mechanism 6 fixes main gears 6a and 6b of the same size to the rotation shafts 3a and 4a of the leaf springs 3 and 4 at the reference position, and meshes them with each other. By rotating the pair of leaf springs 3 and 4 in a plane symmetrical range of 0 ° to 90 ° of the reference position, the rotation angle is continuously changed to make the pair of leaf springs C-shaped or reverse-shaped, and the combined stiffness coefficient. It is possible to change the natural frequency of the entire device by lowering, and the rotation drive of this gear is enabled by engaging the drive gear mechanism 6c of the servo motor 6g for rotation adjustment with one of them, for rotation adjustment The servo motor 6g is provided with a controller 6f that controls the rotational operation to a leaf spring rotation angle that matches the vibration frequency detectors 6d and 6e of the vibration body 1 and the vibration absorbing body 2 based on the detected vibration frequencies. .

Next, one embodiment (specific example) of the present invention (2) will be described with reference to FIG.
In the example of FIG. 4, the leaf springs 3 and 4 shown in FIGS. 1 to 3 are thickened (same large rigid long rigid body), and a thin plate spring for fine adjustment (same small rigid long rigid body) 30 is provided on the outside thereof. 40, and the same parts as those in FIGS. 1 to 3 are denoted by the same reference numerals, and detailed description thereof is omitted.
The leaf springs 3 and 4 and the leaf springs 30 and 40 have their axial centers positioned in a parallel relationship on a plane orthogonal to the vibration direction of the vibrating body 1, and their cross-sectional major axes coincide with the vibration direction in parallel at the reference position. I'm allowed.
The both sides of the leaf springs 30 and 40 and the vibration absorbing body 2 are mounted by forming the both side portions of the leaf springs 30 and 40 on the rotating shafts 30a and 40a, and providing the bearing body 20a and 20b on the vibration absorbing body 2, respectively. Rotary bearings 20a and 20b are mounted with rotary shafts 30a and 40a, and the central part of leaf springs 30 and 40 and the weight 5 are attached to the central part of leaf springs 30 and 40 to the rotary shafts 30b and 40b. The bearing 5 is formed with bearing bearings 50a and 50b, and the bearings 50a and 50b are equipped with rotating shafts 30b and 40b.
In this example of the stiffness coefficient changing mechanism for the leaf spring, a gear mechanism is provided on the weight 5 for balance, and for the leaf springs 3 and 4, the rotating shaft 3b of the leaf springs 3 and 4 at the reference position is provided. The main gears 60a, 60b of the same size are fixed and meshed with 4b, and the rotation angle of the leaf springs 3, 4 is continuously changed symmetrically with the rotation of the rotational drive shaft D1 of the leaf spring 3, thereby a pair of leaf springs 3 4 is set to a C shape or an inverted C shape, and the combined rigidity coefficient is lowered and changed to roughly set the natural frequency of the entire apparatus.
For the leaf springs 30 and 40, the main gears 61a and 61b of the same size are fixed to the central rotating shafts 30b and 40b of the leaf springs 30 and 40 at the reference position, and these are fixed to the rotating shafts 3b of the leaf springs 3 and 4, respectively. 4b is engaged with the intermediate gears 62a and 62b loosely engaged with each other, and the rotation of the rotational drive shaft D2 of the leaf spring 30 causes the leaf springs 30 and 40 to move through the intermediate gears 62a and 62b from 0 ° to 90 ° of the reference position. When changing the natural frequency of the whole device by changing the rotation angle continuously and symmetrically changing the rotation angle in a range of ° to make the pair of leaf springs 30 and 40 into a C shape or reverse C shape and changing the combined stiffness coefficient. Fine adjustment is performed following the leaf springs 3 and 4.

  The variable stiffness type dynamic vibration damping device of the present invention has the above-described structure that can be made compact so that the natural frequency on the vibration damping device side can be easily and accurately matched to the natural frequency of the upper and lower vibrating bodies to suppress the vibration of the vibrating body. Alternatively, it can be reduced or delayed, and can be widely applied to many vibrators introduced in the above-mentioned effects, and this industrial applicability is extremely high. The present invention is also effective for a system having a vibration mode in the rotational direction such as vibration when a disc brake caliper squeals.

It is explanatory drawing which shows one Example of this invention apparatus (1), and is a plane sectional view from arrow AA-AA of FIG. 2, and arrow AA of FIG. The sectional side view from arrow BB-BB of FIG. 1 and arrow BB of FIG. The front view from arrow CC-CC of FIG. 1 and arrow CC of FIG. It is explanatory drawing which shows one Example of this invention apparatus (2), and is equivalent to FIG.

Explanation of symbols

1 Vibrating body 2 Vibration absorbing body 3, 4 Rigid body 5 Weight
3a, 4a rotation axis
2a, 2b Bearing bearing
5a, 5b Bearing bearing 6 Rigidity coefficient changing mechanism
6a, 6b Main gear
6c Drive gear mechanism
6f controller
6g Servo motor for rotation adjustment
30, 40 leaf spring
30a, 40a rotating shaft
20a, 20b bearing bearings
30b, 40b rotation axis
50a, 50b bearing bearings
60a, 60b main gear
D1, D2 Rotary drive shaft
61a, 61b main gear
62a, 62b Intermediate gear

Claims (2)

  A pair of identical rigid long rigid bodies, each having an axial center located on the same orthogonal plane with respect to the vibration direction of the vibrating body and having a long axis and a short axis in the cross section, and both sides of the pair of rigid bodies attached to the vibrating body A vibration-absorbing main body having a rotationally mounted portion, a weight having a central portion of the pair of rigid bodies rotatably mounted thereon, and the pair of rigid bodies rotated in opposite directions in a plane-symmetric state to the pair of rigid bodies A variable-rigidity dynamic vibration damping device comprising: a mechanism for changing the combined stiffness coefficient while maintaining the combined vibration direction of the vibration member in the linear vibration direction of the vibrating body. Two pairs of the same large rigid long rigid body, the same small rigid long rigid body, the same small rigid long rigid body, and the vibrating body, each axis centering on the same orthogonal plane with respect to the vibration direction of the vibrating body and having a long axis and a short axis in the cross section A vibration-absorbing body that is mounted and rotatably mounted on both sides of the two pairs of rigid bodies, and a weight that is rotatably mounted on the center of the two pairs of rigid bodies, and in parallel relation and the same large rigid length The same pair of small rigid and long rigid bodies are disposed outside or inside the rigid body pair, and each of the two pairs of rigid bodies is rotated in the reverse direction in pairs with each other in a plane symmetrical state. A variable-rigidity dynamic vibration damping device comprising: a mechanism for changing a composite stiffness coefficient while maintaining a composite vibration direction of the rigid body in a linear vibration direction of the vibration body.

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