JP5358322B2 - Vibration control device and specification method of vibration control device - Google Patents

Description

  The present invention relates to a vibration control device and a specification setting method for the vibration control device.

  It has been proposed to control a structure using a rotary inertia mass damper using the rotary inertia mass (see, for example, Patent Document 1 and Patent Document 2).

  Also, an additional spring is installed in series with the rotary inertia mass damper, and the rotary inertia mass damper and the additional spring are adjusted so that the natural frequency determined by the rotary inertia mass and the additional spring is synchronized with the natural frequency of the structure. It has been proposed to reduce the response of the structure at this natural frequency by setting the origin (see, for example, Patent Document 3).

JP 2006-125110 A JP 2008-002165 A JP 2008-133947 A

  It is desired to improve the performance of a vibration control device including a rotary inertia mass damper to effectively control the structure.

  In view of the above, an object of the present invention is to effectively control a structure by improving the performance of a vibration control device including a rotary inertia mass damper.

  The invention of claim 1 is a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, and has a mass between the first part and the second part. A rotary inertia mass damper that generates a rotary inertia mass by rotation of the body and a spring are arranged in series, the natural circular frequency determined by the rotary inertia mass and the spring is ωd, and the natural circular frequency of the structure is ω Assuming that the magnification of the rotary inertia mass is αm, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd <1 and αm> 1 according to the damping constant. ing.

  In the first aspect of the invention, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd <1 and αm> 1 according to the damping constant.

  In the load deformation history of the rotary inertia mass damper, the phase with respect to the input of the rotary inertia mass damper is switched at ω / ωd = 1.0. That is, the slope is inverted from negative to positive. Since the rotational inertia mass is a value with respect to acceleration, in order for the rotational inertia mass damper to have resistance to an external force, the inclination needs to be negative. Therefore, when ω / ωd ≧ 1, inertial resistance due to rotational inertial mass does not occur.

  Therefore, ω / ωd <1 is required to obtain a seismic control effect by the rotational inertial mass. Further, in order to amplify the rotational inertial mass, the magnification αm needs to be αm> 1. Furthermore, the range in which αm> 1 varies depending on the attenuation constant.

  Therefore, the rotational inertial mass is amplified by setting the specifications of the rotational inertial mass damper and the spring such that ω / ωd <1 and αm> 1 according to the damping constant. In other words, a large rotational inertial mass can be obtained with a small mass body. Thus, since the performance of the vibration control device is improved, the structure is effectively damped.

  According to a second aspect of the invention, the specifications of the rotary inertia mass damper and the spring are set such that the magnification αm of the rotary inertia mass is maximized.

  In the invention of claim 2, since the magnification of the rotational inertia mass of the rotary inertia mass damper is maximized, the vibration is more effectively suppressed.

  The invention of claim 3 is a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, and has a mass between the first part and the second part. A rotary inertia mass damper that generates a rotary inertia mass by rotation of the body and a spring are arranged in series, the natural circular frequency determined by the rotary inertia mass and the spring is ωd, and the natural circular frequency of the structure is ω Assuming that the spring stiffness magnification of the spring is γk, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd> 1 and γk> 1 according to the damping constant. ing.

  In the invention of claim 3, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd> 1 and γk> 1 according to the damping constant.

  In the load deformation history of the spring, the phase with respect to the input of the spring is switched with ω / ωd = 1.0 as a boundary. That is, the slope is inverted from negative to positive. In order for the spring to have resistance against external force, the inclination needs to be positive. Therefore, when ω / ωd ≦ 1, there is no resistance due to the spring stiffness of the spring.

  Therefore, in order to obtain the damping effect by the spring, it is necessary to satisfy ω / ωd> 1. In order to amplify the spring stiffness of the spring, the magnification γk needs to be γk> 1. Furthermore, the range where γk> 1 varies depending on the attenuation constant.

  Therefore, the spring stiffness of the spring is amplified by setting the rotational inertia mass damper and the specifications of the spring so that ω / ωd> 1 and γk> 1 according to the damping constant. In other words, a large spring rigidity can be obtained with a spring having a small spring rigidity. Thus, since the performance of the vibration control device is improved, the structure is effectively damped.

  According to a fourth aspect of the present invention, the specifications of the rotary inertia mass damper and the spring are set so that the spring rigidity magnification γk of the spring is maximized.

  In the invention of claim 4, since the magnification of the spring stiffness of the spring is maximized, the vibration is more effectively suppressed.

  The invention of claim 5 is a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, and has a mass between the first part and the second part. A rotary inertia mass damper that generates a rotary inertia mass by rotating the body and a damping damper are arranged in parallel, and the rotary inertia mass damper arranged in parallel, the damping damper, and a spring are arranged in series, and the rotary inertia Assuming that the natural circular frequency determined by the mass and the spring is ωd, the natural circular frequency of the structure is ω, and the magnification of the damping coefficient of the damping damper is αc, ω / ωd <1 and a damping constant Accordingly, the specifications of the rotary inertia mass damper and the spring are set so that αc> 1.

  In the fifth aspect of the invention, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd <1 and αc> 1 according to the damping constant.

  In the relationship between the attenuation coefficient magnification αc and ω / ωd, the peak value of the magnification αc becomes smaller than ω / ωd = 1 as the attenuation constant (ratio of attenuation coefficient to critical attenuation coefficient) hd increases. (The peak value is in the range of ω / ωd <1).

  Accordingly, the damping coefficient of the damping damper is amplified by setting the specifications of the rotary inertia mass damper and the spring so that ω / ωd <1 and αc> 1 according to the damping constant, in other words, A large damping coefficient can be obtained with a damping damper having a small damping coefficient. Since the performance of the vibration control device is thus improved, the structure is effectively damped.

  In the sixth aspect of the invention, the specifications of the rotary inertia mass damper and the spring are set so that the magnification αc of the damping coefficient is maximized.

  According to the sixth aspect of the present invention, the damping coefficient magnification αc is maximized, so that the vibration is more effectively suppressed.

  The invention according to claim 7 is a vibration damping device provided between the first part and the second part of the structure that moves relative to each other due to a disturbance, and has one end connected to the first part and an opening at the other end. An outer cylinder having a portion formed therein, a moving part provided in the outer cylinder so as to be movable in the axial direction and not to rotate around the axis, and the other end of the outer cylinder than the moving part A coil-shaped spring disposed between a contact portion provided on the part side and the moving portion, contacting the moving portion and the contact portion, and extending and contracting as the moving portion moves in the axial direction; A shaft body provided in the moving portion and extending in the axial direction in the coiled spring to the other end side, and the outer portion from the opening formed in the other end portion of the outer cylinder. An inner cylinder in which one end is inserted into the coiled spring in the cylinder and the other end is connected to the second part; A rotating body that is rotatably held around an axis in an inner cylinder and into which the shaft body is inserted, and an axial direction of the shaft body provided on an outer peripheral surface of the shaft body and an inner peripheral surface of the rotating body A rotary inertia mass damper having screwing means for converting the linear displacement into a rotational displacement around the axis of the rotating body, and a mass body integrally rotating with the rotating body and rotating around the axis. .

  In the invention of claim 7, the inner cylinder, the rotating body, and the shaft body constituting the rotary inertia mass damper move in the axial direction in the coil-shaped spring. Therefore, compared with a case where the rotary inertia mass damper and the spring are simply connected in the axial direction, the total length in the axial direction of the portion constituted by the rotary inertia mass and the spring in the vibration control device is shortened.

  The invention of claim 8 is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, wherein the first part and the second part Rotational inertia mass damper that generates rotational inertia mass by rotation of the mass body and a spring are arranged in series, and the natural circular frequency determined by the rotational inertia mass and the spring is ωd, When the circular frequency is ω and the magnification of the rotary inertia mass is αm, ω / ωd <1 and αm> 1 according to the damping constant, so that the rotary inertia mass damper and the spring are Set the specifications.

  In the invention of claim 8, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd <1 and αm> 1 according to the damping constant.

  In the load deformation history of the rotary inertia mass damper, the phase with respect to the input of the rotary inertia mass damper is switched at ω / ωd = 1.0. That is, the slope is inverted from negative to positive. Since the rotational inertia mass is a value with respect to acceleration, in order for the rotational inertia mass damper to have resistance to an external force, the inclination needs to be negative. Therefore, when ω / ωd ≧ 1, inertial resistance due to rotational inertial mass does not occur.

  Therefore, ω / ωd <1 is required to obtain a seismic control effect by the rotational inertial mass. Further, in order to amplify the rotational inertial mass, the magnification αm needs to be αm> 1. Furthermore, the range in which αm> 1 varies depending on the attenuation constant.

  Therefore, the rotational inertial mass is amplified by setting the specifications of the rotational inertial mass damper and the spring such that ω / ωd <1 and αm> 1 according to the damping constant. In other words, a large rotational inertial mass can be obtained with a small mass body. Thus, since the performance of the vibration control device is improved, the structure is effectively damped.

  The invention of claim 9 sets the specifications of the rotary inertia mass damper and the spring so that the magnification αm of the rotary inertia mass is maximized.

  In the invention of claim 9, since the magnification of the rotary inertia mass of the rotary inertia mass damper is maximized, the vibration is more effectively suppressed.

  The invention of claim 10 is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, wherein the first part and the second part Rotational inertia mass damper that generates rotational inertia mass by rotation of the mass body and a spring are arranged in series, and the natural circular frequency determined by the rotational inertia mass and the spring is ωd, When the circular frequency is ω and the spring stiffness magnification of the spring is γk, ω / ωd> 1, and γk> 1 according to the damping constant, so that the rotary inertia mass damper and the spring Set the specifications.

  In the invention of claim 10, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd> 1 and γk> 1 according to the damping constant.

  In the load deformation history of the spring, the phase with respect to the input of the spring is switched with ω / ωd = 1.0 as a boundary. That is, the slope is inverted from negative to positive. In order for the spring to have resistance against external force, the inclination needs to be positive. Therefore, when ω / ωd ≦ 1, there is no resistance due to the spring stiffness of the spring.

  Therefore, in order to obtain the damping effect by the spring, it is necessary to satisfy ω / ωd> 1. In order to amplify the spring stiffness of the spring, the magnification γk needs to be γk> 1. Furthermore, the range where γk> 1 varies depending on the attenuation constant.

  Therefore, the spring stiffness of the spring is amplified by setting the rotational inertia mass damper and the specifications of the spring so that ω / ωd> 1 and γk> 1 according to the damping constant. In other words, a large spring rigidity can be obtained with a spring having a small spring rigidity. Thus, since the performance of the vibration control device is improved, the structure is effectively damped.

  In the eleventh aspect of the invention, the specifications of the rotary inertia mass damper and the spring are set so that the spring rigidity magnification γk of the spring is maximized.

  In the invention of claim 11, since the magnification of the spring rigidity of the spring is maximized, the vibration is more effectively suppressed.

  The invention of claim 12 is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relative to each other due to a disturbance, wherein the first part and the second part The rotary inertia mass damper and the damping damper that generate the rotary inertia mass by the rotation of the mass body are arranged in parallel, and the rotary inertia mass damper arranged in parallel and the damping damper and the spring are arranged in series. When the natural circular frequency determined by the rotational inertia mass and the spring is ωd, the natural circular frequency of the structure is ω, and the magnification of the damping coefficient of the damping damper is αc, ω / ωd <1 And according to the damping constant, the specifications of the rotary inertia mass damper and the spring are set so that αc> 1.

  According to the invention of claim 12, the specifications of the rotary inertia mass damper and the spring are set so that ω / ωd <1 and αc> 1 according to the damping constant.

  In the relationship between the attenuation coefficient magnification αc and ω / ωd, the peak value of the magnification αc becomes smaller than ω / ωd = 1 as the attenuation constant (ratio of attenuation coefficient to critical attenuation coefficient) hd increases. (The peak value is in the range of ω / ωd <1).

  Accordingly, the damping coefficient of the damping damper is amplified by setting the specifications of the rotary inertia mass damper and the spring so that ω / ωd <1 and αc> 1 according to the damping constant, in other words, A large damping coefficient can be obtained with a damping damper having a small damping coefficient. Since the performance of the vibration control device is thus improved, the structure is effectively damped.

  The invention of claim 13 sets the specifications of the rotary inertia mass damper and the spring so that the magnification αc of the damping coefficient is maximized.

  In the invention of claim 13, since the magnification αc of the attenuation coefficient is maximized, the vibration is more effectively suppressed.

  According to the first aspect of the present invention, the rotational inertial mass of the rotary inertia mass damper is amplified and the performance of the vibration control device is improved compared to the case where this configuration is not applied, so that the structure is effectively controlled. Can shake.

  According to the invention described in claim 2, since the magnification at which the rotational inertial mass is amplified is maximum, the structure can be more effectively damped.

  According to the third aspect of the present invention, the spring rigidity of the spring is amplified and the performance of the vibration control device is improved as compared with the case where the present configuration is not applied. it can.

  According to the fourth aspect of the present invention, since the magnification at which the spring stiffness of the spring is amplified is maximum, the structure can be more effectively damped.

  According to the fifth aspect of the present invention, the damping coefficient of the damping damper is amplified and the performance of the damping device is improved compared to the case where this configuration is not applied, so that the structure is effectively damped. Can do.

  According to the sixth aspect of the present invention, since the magnification at which the damping coefficient of the damping damper is amplified is maximum, the structure can be more effectively damped.

  According to invention of Claim 7, compared with the case where a rotation inertia mass damper and a spring are simply connected to an axial direction, the axial total length of the site | part comprised with the rotation inertia mass and the spring in a damping device Can be shortened.

  According to the eighth aspect of the present invention, the rotational inertial mass of the rotary inertial mass damper is amplified and the performance of the vibration control device is improved compared to the case where the specification setting method is not applied. Can be controlled.

  According to the invention described in claim 9, since the magnification at which the rotational inertial mass is amplified is the maximum, the structure can be more effectively damped.

  According to the tenth aspect of the present invention, since the spring rigidity of the spring is amplified and the performance of the vibration control device is improved as compared with the case where the specification setting method is not applied, the structure is effectively damped. can do.

  According to the eleventh aspect of the present invention, since the magnification at which the spring stiffness of the spring is amplified is maximum, the structure can be more effectively damped.

  According to the twelfth aspect of the present invention, the damping coefficient of the damping damper is amplified and the performance of the vibration control device is improved as compared with the case where the specification setting method is not applied. Can shake.

  According to the thirteenth aspect of the present invention, since the magnification at which the damping coefficient of the damping damper is amplified is maximum, the structure can be more effectively damped.

It is a model figure of the damping device which concerns on embodiment of this invention. It is sectional drawing along the axial direction which shows the damping device of the Example of this invention. It is a section perspective view of the damping device of the example of the present invention. It is a front view which shows the state in which the damping device of the Example of this invention was provided in the frame which comprises a structure. (A) is a figure of the state which the structure moved horizontally to the right side from the state of FIG. 4, (B) is a figure of the state which the structure moved horizontally from the state of FIG. It is a graph which shows the relationship between the rotational displacement of the rotational inertia mass damper of a damping device, the response displacement of the spring rigidity of a spring, and time. It is a graph which shows the relationship between the load curve of a rotation mass damper, and omega / omegad. It is a graph which shows the example of the equivalent rigidity in the relationship between the load curve of a rotation mass damper, and omega / omegad. It is a graph which shows the relationship between magnification (alpha) m of rotation inertia mass, and (omega) / (omega) d. It is a graph which shows the relationship between magnification γk of spring stiffness of a spring and ω / ωd. It is a graph which shows the relationship between magnification αc of an attenuation coefficient, and ω / ωd. It is the graph which rewrote the graph of FIG. It is the graph which expanded FIG.

A vibration control device according to an embodiment of the present invention will be described with reference to FIGS.
As shown in the model diagram of FIG. 1, the vibration damping device 100 of the present invention converts rotational movement of a mass body into rotational motion of a mass body when an interlayer displacement occurs in an arbitrary layer of a structure. The md is generated, and the rotary inertia mass damper 200 and the spring 400 having the spring stiffness Kd are arranged in series.

  The damping damper 300 having the damping coefficient Cd arranged in parallel with the rotary inertia mass damper 200 is theoretically necessary as described later, but a mechanism for rotating the mass body of the rotary inertia mass damper 200, etc. Therefore, it is not always necessary to add the damping damper 300. In addition, when adding the damping damper 300, you may integrate in the rotary inertia mass damper 200 integrally.

  The configuration (structure) of the rotary inertia mass damper 200 is not particularly limited, and a desired type and characteristics may be arbitrarily adopted.

  Next, the structure of the vibration control device 100 of the model diagram shown in FIG. 1 will be described more specifically with reference to examples. In addition, the damping device 100 of an Example is an example, Comprising: It is not limited to the structure of this Example. For example, rotational inertia mass dampers described in JP-A-2006-125110 and JP-A-2008-002165 are also applicable to the present invention. The principle of the rotary inertia mass damper is the same as the principle described in JP-A-2006-125110 and JP-A-2008-002165.

  As shown in FIG. 4, the seismic control device 100 </ b> L and the seismic control device 100 </ b> R are installed in the frame 24 composed of the left column 20 </ b> L, the right column 20 </ b> R, the upper beam 22 </ b> A, and the lower beam 22 </ b> B that constitute the structure 10. Are arranged side by side. Further, the left side vibration control device 100L and the right side vibration control device 100R are arranged symmetrically in the frame 24.

  Specifically, the vibration control device 100L is rotatably connected at one end to a rotary bearing 30 attached to the substantially central portion of the upper beam 22A, and has the other end at a corner between the lower beam 22B and the column 20L. It is rotatably connected to the attached rotary bearing 40L. The vibration control device 100R has one end rotatably connected to a rotation support 36 attached to the substantially central portion of the upper beam 22A, and the other end attached to a corner of the lower beam 22B and the column 20R. It is rotatably connected to 40R.

  In the following description, when it is necessary to distinguish between left and right, either L or R is added after the symbol, and when it is not necessary to distinguish, L and R are omitted.

  In addition, the installation (arrangement) of the vibration control device as shown in FIG. 4 is an example, and if the horizontal deformation of the frame 24 (see FIGS. 5A and 5B) can be suppressed, how will it be? It may be installed (arranged). Furthermore, you may install in the frame of any layer (floor) of the structure 10. FIG. Moreover, you may install in site | parts other than a frame. That is, what is necessary is just to install so that the damping effect may be acquired by providing the damping device 100 between the 1st site | part which the structure 10 moves relatively by disturbances, such as an earthquake, and a 2nd site | part. Further, a toggle mechanism may be applied as described in JP 2008-002165 A.

  Next, the vibration control device 100 will be described in detail. Since the left and right vibration control devices 100R and 100L have the same structure, LR will be omitted.

  As shown in FIGS. 2 and 3, the vibration damping device 100 of the embodiment has an outer cylinder 110 and an inner cylinder 140, and the inner cylinder 140 is inserted into the outer cylinder 110. The inner cylinder 140 is inserted from the opening 112 of the outer cylinder 110 and can move in the outer cylinder 110 along the axial direction.

  An arm 120 extending in the axial direction is provided at the end of the outer cylinder 110 opposite to the opening 112, and the end of the arm 120 is rotatably connected to the rotary support 30 (see FIG. 4). .

  An arm 190 extending in the axial direction is provided at the end of the inner cylinder 140 opposite to the insertion side of the outer cylinder 110, and the end of the arm 192 is rotatably connected to the rotary support 40 (see FIG. 4).

  A disk 180 is provided in the outer cylinder 110 so as to be movable along the axial direction. Ribs 182 are formed on the side wall of the disk 180 along the axial direction. On the inner wall surface 110 </ b> A of the outer cylinder 110, a concave groove 115 that engages with the rib 182 of the disk 180 is formed along the axial direction. Therefore, the disk 180 is configured to be movable in the axial direction in the outer cylinder 110 without rotating around the axis. Note that a small diameter portion 186 is formed in the disk 180.

  A flange 114 is formed in the opening 112 of the outer cylinder 110. A coiled spring (coil spring) 400 is disposed between the flange 114 and the disk 180. Further, both end portions of the spring 400 are in contact with the disk 180 and the flange 114. One end of the spring 400 is fitted between the small diameter portion 186 of the disk 180 and the inner wall surface 110 </ b> A of the outer cylinder 110.

  A shaft 160 is provided at the center of the disk 180. The shaft 160 extends along the axial direction in the spring 400 toward the opening 112.

  The inner cylinder 140 includes a rotating body holding part 142 and a connecting part 144. The rotating body holding part 142 is inserted into the outer cylinder 110, and the connecting part 144 is connected to the arm 190 described above.

  In the rotating body holding part 142 of the inner cylinder 140, the rotating body 150 is held so as to be rotatable about an axis. A ball 172 is provided between the inner peripheral surface 142A of the rotating body holding portion 142 and the outer peripheral surface 150A of the rotating body 150. The viscous body (viscous liquid) 302 provided (or may be provided) in the gap between the rotation holding unit 142 and the rotating body 150 will be described later.

  The distal end portion 152 on one end side of the rotating body 150 is exposed from the opening portion 146 on the side inserted into the outer cylinder 110 of the rotating body holding portion 142. Further, the other end portion side of the rotating body 150 is exposed from the opening portion 148 on the side that is not inserted into the outer cylinder 110 of the rotating body holding portion 142, and a flange portion 154 is formed at the other end portion. And the flange part 154 of this other end part is engage | inserted by the recessed part 145 formed in the edge part of the connection part 144 mentioned later. Therefore, the rotating body 150 has a structure that rotates around the axis without moving relative to the inner cylinder 140 (the rotating body holding portion 142) in the axial direction. Moreover, the inside of the rotary body 150 is hollow (the rotary body 150 has a hollow structure).

  A female thread groove 162 is formed on the outer peripheral surface of the shaft 160 provided in the disk 180 described above. The female screw groove is inserted into the through hole 151 of the tip portion 152 of the rotating body 150. A male screw groove (not shown) into which the female screw groove 162 is screwed is formed on the inner peripheral surface of the through hole 151 of the distal end portion 152 of the rotating body 150. Therefore, when the shaft 160 moves the tip portion 152 of the rotating body 150 in the axial direction. The rotating body 150 is configured to rotate around the axis.

  A mass body 178 having a diameter larger than that of the rotating body 150 is provided on the outer peripheral surface of the rotating body 150 exposed from the opening 148 of the rotating body holding portion 142 of the inner cylinder 140. Therefore, the mass body 178 rotates integrally with the rotating body 150. And the connection part 144 is formed in the substantially square frame shape by the side view so that the outer side of this mass body 178 may be enclosed (it does not interfere).

  The shaft 160, the inner cylinder 140, the rotating body 150, and the mass body 178 constitute the rotary inertia mass damper 200.

  The axis of the rotating body 150, the axis of the disk 180, the axis of the mass body 178, and the axis of the shaft 160 are on the same axis. These are the axis of the rotary inertia mass damper 200 and also coincide with the coil center (axis) of the coiled spring 400.

  Here, as described above, a viscous body (viscous liquid) 302 such as oil is inserted between the rotating body holding portion 142 of the inner cylinder 140 and the rotating body 150 (while the ball 172 is provided), so that the damping damper ( An oil damper) 300 may be used (a damper in which a rotary inertia mass damper and a damping damper (oil damper) are integrally formed). The viscous body 302 is sealed with an oil seal (not shown) or the like, and has a structure that does not leak.

Next, the operation of the vibration control device 100 of the present embodiment will be described. As shown in FIG. 5A, when the structure 10 moves horizontally to the right side due to vibration (disturbance) such as seismic motion, the frame 24 moves to the right. Deform horizontally (deform into parallelogram). That is, the upper beam 22A moves horizontally to the right (the upper beam 22A and the lower beam 22A move relative).

  Therefore, in the frame 24, the distance between the rotary bearing 30 and the rotary bearing 40L becomes longer, and the distance between the rotary bearing 30 and the rotary bearing 40L becomes shorter. Thus, due to the change in the distance between the rotary bearing 30 and the rotary bearings 40L and 40R, the overall length of the vibration control device 100L is expanded, and the total length of the vibration control device 100R is contracted.

  Further, as shown in FIG. 5B, when the structure 10 moves horizontally to the left, the frame 24 is horizontally deformed (deformed into a parallelogram) in the left direction. That is, the upper beam 22A moves horizontally to the left (the upper beam 22A and the lower beam 22A move relative to each other).

  Therefore, in the frame 24, the distance between the rotating bearing 30 and the rotating bearing 40L is shortened, and the distance between the rotating bearing 30 and the rotating bearing 40R is increased. Thus, due to the change in the distance between the rotary bearing 30 and the rotary bearings 40L and 40R, the overall length of the vibration control device 100R contracts, and the total length of the vibration control device 100L extends.

  Therefore, when the frame 24 vibrates left and right due to disturbance (vibration) such as an earthquake, the distance between the rotary bearing 30 and the rotary bearings 40R and 40L changes, and the overall length of the vibration control device 100 expands and contracts.

  Due to this expansion and contraction (change in the total length), the disk 180 moves in the axial direction. As the disk 180 moves in the axial direction, the coiled spring 400 expands and contracts. Further, the rotating body 150 rotates around the axis and the mass body 178 integrated with the rotating body 150 rotates around the axis (the rotating body 150 and the mass body 178 rotate together).

  Next, a description will be given of a vibration control device in which the rotational inertial mass is amplified and a first specification setting method for increasing the magnification αm of the rotational inertial mass md.

  The natural circular frequency determined by the rotational inertia mass md generated by rotating the mass body 178 of the damping device 100 and the spring 400 is ωd,

  If the natural circular frequency of the structure 10 (the natural circular frequency when the frame 24 vibrates (deforms) left and right as shown in FIGS. 5A and 5B) is ω,

  ω / ωd <1 and magnification αm> 1

  The specifications of the rotary inertia mass damper 200 and the spring 400 (specifically, the spring constant of the spring 400, the total length of the spring 400, the weight of the mass body 178, the diameter of the mass body 178, the mass body 178 Rotation amount (ratio between axial movement and rotation amount, etc.) is set.

  Here, FIG. 6 is a graph showing the relationship between the rotational inertia mass of the rotary inertia mass damper 200 of the damping device 100 and the response displacement of the spring stiffness of the spring 400 and time. The DMC-M line indicates the rotational inertia mass of the rotary inertia mass damper 200, and the K line indicates the spring rigidity of the spring 400. 7 and 8 are graphs showing the relationship between the load curve of the rotating mass damper 200 and ω / ωd.

  As shown in FIGS. 6 to 8, the phase with respect to the input of the rotary inertia mass damper 200 is switched at the boundary of ω / ωd = 1.0. That is, the slope is inverted from negative to positive.

  Since the rotational inertial mass (dynamic mass (DM)) md is a value with respect to acceleration, the inclination needs to be negative in order for the rotational inertial mass md to have resistance to external force. That is, when ω / ωd ≧ 1, resistance to an external force (dynamic mass effect) due to the rotational inertial mass md cannot be obtained. In other words, it can be seen that ω / ωd <1 is a necessary condition in order to obtain resistance (dynamic mass effect) due to the rotational inertial mass md. Note that ω / ωd = 1 is a state where the structure 10 and the vibration control device 100 are resonated.

Here, the magnification αm will be described.
The equivalent stiffness obtained from the history with damping is taken as the stiffness at the maximum displacement point of the total length of the damping device 100, and the ratio of the stiffness of the rotary inertia mass body (dynamic mass alone) to the dynamic mass with damping is calculated. It is defined as magnification (magnification of rotational inertia mass md generated by mass body 178). An equation for calculating the magnification αm is [Equation 1]. Further, a graph of this [Equation 1] is the graph of FIG.

  The meaning of each symbol is as follows.

：回転慣性質量の倍率

：直列したばねにより増幅された回転慣性質量

：回転慣性質量体（ダイナミックマス単体）の質量

：バネ剛性

：制震装置の固有円振動数

：構造物の固有円振動数

：位相角

：減衰定数（減衰係数と臨界減衰係数との比）

: Rotational inertia mass magnification

: Rotational inertial mass amplified by a series of springs

: Mass of rotating inertial mass (dynamic mass alone)

: Spring stiffness

: Natural frequency of vibration control device

: Natural circular frequency of structure

: Phase angle

: Damping constant (ratio between damping coefficient and critical damping coefficient)

  As shown in FIG. 9, when ωd = 1.0, the magnification αm is −1.0. Therefore, the rotational inertial mass md is amplified by setting the specifications of the rotational inertial mass damper 200 and the spring 400 such that ω / ωd <1.0 and the magnification αm is greater than 1. . In other words, the resistance (dynamic mass effect) to the external force due to the rotational inertial mass is amplified. Therefore, since a large rotational inertial mass can be obtained with the small mass body 178, the structure 10 is effectively damped.

  As shown in FIG. 9, when the damping constant hd of the damping coefficient Cd of the damping damper 300 is large, the magnification αm is suppressed. Therefore, it can be seen that a large effect can be obtained by reducing the attenuation constant hd.

  Note that ω / ωd at which αm> 1.0 and αm is maximum varies depending on the attenuation constant hd, and may be determined by calculation based on the value of hd.

  For example, when hd = 0.2, the range of ω / ωd where αm> 1 is 0.54 <ω / ωd <0.74. In addition, when hd = 0.02, ω / ωd = 0.97 and the magnification αm is maximized, which is optimal. Note that αm may be larger than 1 even though the magnification αm is not necessarily maximum.

  When hd = 0, ω / ωd = 1 does not become continuous and diverges, so hd> 0 needs to be required (attenuation is necessary). However, even when the damping damper 300 (viscous body 302) is not incorporated, the damping always occurs due to various frictional resistances such as a mechanism for rotating the mass body 178 of the rotary inertia mass damper 200 and viscous resistance due to grease. To do. Therefore, the damping damper 300 (viscous body 302) is not necessarily added (incorporated).

  Here, as the vibration control effect of the structure using the rotary inertia mass damper (rotational inertia mass), there are a vibration control effect by the input reduction effect and a vibration control effect by the mode control. Mode control is a technology that effectively controls the structure by controlling the vibration mode of the structure and eliminating the higher-order mode.

  However, in order to perform mode control, it is necessary to generate a rotational inertial mass larger than the mass of the structure. That is, a large and heavy mass body 178 is necessary. However, if the configuration includes a large and heavy mass body 178, an increase in installation space and sufficient strength of the entire vibration control device 100 that supports the heavy mass body 178 are required (for example, an increase in the arms 120 and 190). In addition, a failure may occur.

  However, the damping device 100 set by the setting method of this specification amplifies the rotational inertial mass (dynamic mass effect) without increasing the rotational inertial mass (mass body 178). It is very effective.

  Next, a description will be given of a vibration damping device with amplified spring stiffness and a second specification setting method for increasing the magnification γk of the spring 400 spring stiffness Kd. Further, since the vibration control device itself has the same configuration, the description thereof is omitted.

  As described with reference to FIGS. 6 to 7, the phase with respect to the input to the spring 400 is switched at ω / ωd = 1.0. That is, the slope is inverted from negative to positive. In order for the spring 400 to have resistance against an external force, the inclination needs to be positive. That is, when ω / ωd ≦ 1, resistance is to external force. In other words, it is understood that ω / ωd> 1 is a necessary condition for the spring 400 to have resistance against external force.

Here, the magnification γk will be described.
That is, the ratio of the spring 400 to the spring stiffness is defined as the spring stiffness magnification γk. The equation for calculating the magnification γk is [Equation 2]. Further, the graph of FIG. 10 is a graph of this [Equation 2].

：減衰のある場合の制震装置１００の最大変形時と荷重の関係から求まる剛性が等価バネ剛性

: Rigidity obtained from the relationship between the maximum deformation of the damping device 100 and the load when there is damping is equivalent spring stiffness

  In addition, each code other than the above is the same as that described in [Equation 1].

  As can be seen from the graph of FIG. 10, by setting the specifications of the rotary inertia mass damper 200 and the spring 400 so that ω / ωd> 1, γk> 1, that is, the amplification effect of the spring stiffness Kd. Occurs. However, as ω / ωd is increased, γk converges to 0, that is, γk ≦ 1. Therefore, the specifications of the rotary inertia mass damper 200 and the spring 400 are set so as to be in the range of ω / ωd where γk> 1.

  Therefore, by setting the specifications of the rotary inertia mass damper 200 and the spring 400 so that ω / ωd> 1 and γk> 1, the spring stiffness of the spring 400 is amplified, in other words, the spring stiffness is small. Since large spring rigidity is obtained by the spring 400, the structure 10 is effectively damped.

  Further, as shown in FIG. 10, when the damping constant hd of the damping coefficient Cd of the damping damper 300 is large, the magnification γk is suppressed. Therefore, it can be seen that a large effect can be obtained by reducing the attenuation constant hd.

  Note that ω / ωd at which γk> 1.0 and γk is maximized varies depending on the attenuation constant hd, and may be calculated and determined based on the value of hd.

  For example, in the case of hd0.01, 1.0 <ω / ωd <1.415 and γk> 1. Further, when hd = 0.02, ω / ωd = 1.03 and the magnification γk is maximized, so it is optimal. Note that γk only needs to be larger than 1 even though the magnification γk is not necessarily the maximum.

  Next, a damping device in which the damping coefficient Cd of the damping damper (oil damper) 300 is amplified, and a third specification setting method for increasing the magnification αc of the damping coefficient Cd of the damping damper (oil damper) 300 will be described. . Further, since the vibration control device itself has the same configuration, the description thereof is omitted.

The magnification αc will be described.
An equivalent attenuation coefficient obtained from the history area when there is attenuation is obtained, and the ratio with the attenuation coefficient in the case of the damping damper 300 alone (oil damper alone) is defined as a magnification αc of the damping coefficient in the case of damping. The magnification αc is the square of the response magnification of the rotational inertia mass md (dynamic mass). The equation for obtaining the magnification αm is [Equation 3]. Further, FIG. 11 is a graph of this [Equation 3].

：減衰のある場合の履歴面積から求まる等価な減衰係数 Also,

: Equivalent attenuation coefficient obtained from the history area with attenuation

  In addition, each code | symbol other than the above is the same as that of what was demonstrated by [Equation 1] and [Equation 2].

  As shown in FIG. 11, in the relationship between the magnification αc of the attenuation coefficient Cd and ω / ωd, the peak value of the magnification αc becomes smaller than ω / ωd = 1 as the attenuation constant hd increases ( The peak value is in the range of ω / ωd <1).

  Therefore, by setting the specifications so that ω / ωd <1 and αc> 1, the damping coefficient Cd of the damping damper 300 is amplified. In other words, the damping damper 300 having a small damping coefficient Cd has a large damping. Since the coefficient is obtained, the structure 10 is effectively damped.

  Further, ω / ωd at which αc> 1.0 and αc is maximum varies depending on the attenuation constant hd, and may be determined by calculation based on the value of hd.

  As can be seen from the graph of FIG. 11, when the attenuation constant hd of the attenuation coefficient Cd is reduced, the magnification αc is increased.

  Therefore, according to the value of hd, ω / ωd is calculated so that the product of the magnification αc and the damping coefficient Cd of the damping damper 300 alone is maximized, that is, the damping coefficient Cd after amplification is maximized. In addition, the specifications of the rotary inertia mass damper, the spring, and the damping damper may be determined.

  More specifically, when ω / ωd = 1 (resonance point),

It becomes.
Since the magnification αc increases in inverse proportion to the square of the damping constant hd, as described above, in the vicinity of ω / ωd = 1 (near the resonance point), the smaller the damping coefficient Cd, the larger the damping coefficient when considering the magnification αc. Will get.

を満足する

に対する増幅された等価減衰定数

の比率を求めると、図１２と図１３となる。なお、図１３は、図１２のグラフを拡大したグラフである。 So, redraw the graph in Fig.

Satisfy

Amplified equivalent damping constant for

When the ratio is obtained, FIG. 12 and FIG. 13 are obtained. FIG. 13 is an enlarged graph of the graph of FIG.

得られることが分かる。 As can be seen from the graphs of FIGS. 12 and 13, when ω / ωd = 0.95, the attenuation constant hd is half that of the attenuation constant hd = 0.10. The larger equivalent damping constant is

You can see that

  When the attenuation coefficient Cd is amplified and used as in this mechanism, ω / ωd = 0.90 to 0.95 and hd = 0.05 to 0.10 can effectively obtain the attenuation effect. .

  The first item setting method (the item setting method for amplifying the rotational inertia mass magnification αm) and the third item setting method (the item setting method for amplifying the attenuation coefficient magnification αc) are compatible. Is possible. Therefore, the structure 10 can be more effectively damped by the amplified rotational inertial mass and the amplified attenuation coefficient. At this time, the vibration control method, the structure of the structure, the installation method of the vibration control device, etc. should be used so that the amplified rotational inertial mass and the amplified damping coefficient exhibit the most effective vibration control effect. Considering this, the specifications of the rotary inertia mass damper, the spring, and the damping damper may be determined.

  Here, the inventors of the present invention have discovered for the first time the theory of amplification of the rotational inertia mass, the phenomenon of amplification of the spring stiffness, and the phenomenon of amplification of the damping coefficient in the vibration damping device in which the rotary inertia mass damper and the spring are arranged in series. (Formula). Experiments have confirmed that the theory is correct.

  In addition, [Equation 1] to [Equation 3] are theoretical equations to the last, and various noises and the like are added in the actual design, so that desired specifications may deviate from these equations. Therefore, in actual design, adjustment (tuning) may be made as appropriate based on these mathematical expressions.

  The spring 400 arranged in series with the rotary inertia damper 200 yields when a predetermined deformation amount is exceeded, and functions as a steel-based elastic-plastic damper. Therefore, when the spring 400 is arranged in series as described above, a fail-safe function is provided when the direct earthquake motion input is larger than expected.

  The present invention is not limited to the above embodiment. Needless to say, various embodiments can be implemented without departing from the scope of the present invention.

10 Structure 24 Frame 30 Rotating bearing (first part)
40L Rotating bearing (second part)
40R Rotating bearing (second part)
100 Damping device 110 Outer cylinder 114 Flange (contact part)
140 Inner cylinder 160 Shaft (shaft)
178 Mass 180 Disc (moving part)
200 Rotating inertia mass damper 300 Damping damper 400 Spring 142 Rotating body md Rotating inertia mass Cd Damping coefficient Kd Spring stiffness

Claims (13)

A vibration damping device provided between the first part and the second part of the structure that moves relative to each other due to a disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of the mass body and a spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
When the magnification of the rotary inertia mass is αm,
ω / ωd <1
and,
Depending on the damping constant hd
αm> 1 so that
A vibration control device in which specifications of the rotary inertia mass damper and the spring are set.   The vibration control device according to claim 1, wherein specifications of the rotary inertia mass damper and the spring are set so that a magnification αm of the rotary inertia mass is maximized. A vibration damping device provided between the first part and the second part of the structure that moves relative to each other due to a disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of the mass body and a spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
If the spring stiffness magnification of the spring is γk,
ω / ωd> 1
and,
Depending on the damping constant,
so that γk> 1.
A vibration control device in which specifications of the rotary inertia mass damper and the spring are set.   The vibration control device according to claim 3, wherein specifications of the rotary inertia mass damper and the spring are set so that a spring rigidity magnification γk of the spring is maximized. A vibration damping device provided between the first part and the second part of the structure that moves relative to each other due to a disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of a mass body and a damping damper are arranged in parallel, and the rotary inertia mass damper arranged in parallel and The damping damper and the spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
When the magnification of the damping coefficient of the damping damper is αc,
ω / ωd <1
and,
Depending on the damping constant,
αc> 1 so that
A vibration control device in which specifications of the rotary inertia mass damper and the spring are set.   The vibration control device according to claim 5, wherein specifications of the rotary inertia mass damper and the spring are set so that a magnification αc of the damping coefficient is maximized. A vibration damping device provided between the first part and the second part of the structure that moves relative to each other due to a disturbance,
An outer cylinder having one end connected to the first part and an opening formed at the other end;
A moving part provided in the outer cylinder so as to be movable in the axial direction and not to rotate around the axis;
It is arranged between the moving part and the abutting part provided on the other end side of the moving part in the outer cylinder, abuts on the moving part and the abutting part, A coiled spring that expands and contracts with axial movement;
A shaft provided in the moving portion and extending in the coiled spring along the axial direction toward the other end, and the outer cylinder from the opening formed in the other end of the outer cylinder An inner cylinder in which one end is inserted into the coiled spring inside and the other end is connected to the second portion; and the shaft is rotatably held around the axis in the inner cylinder A screw provided on a rotating body into which the body is inserted, an outer peripheral surface of the shaft body, and an inner peripheral surface of the rotating body, and converts a linear displacement in the axial direction of the shaft body into a rotational displacement around the axis of the rotating body. A rotary inertia mass damper having a coupling means and a mass body that rotates integrally with the rotary body around an axis;
The vibration control device according to any one of claims 1 to 6, further comprising: It is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relatively by disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of the mass body and a spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
When the magnification of the rotary inertia mass is αm,
ω / ωd <1
and,
Depending on the damping constant,
αm> 1
The specifications setting method of the damping device which sets the specifications of the said rotary inertia mass damper and the said spring so that it may become.   The specification setting method of the damping device of Claim 8 which sets the specification of the said rotary inertia mass damper and the said spring so that the magnification (alpha) m of a rotary inertia mass may become the maximum. It is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relatively by disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of the mass body and a spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
If the spring stiffness magnification of the spring is γk,
ω / ωd> 1
and,
γk> 1
The specifications setting method of the damping device which sets the specifications of the said rotary inertia mass damper and the said spring so that it may become.   The specification setting method of the damping device according to claim 10, wherein the specifications of the rotary inertia mass damper and the spring are set so that the spring rigidity magnification γk of the spring is maximized. It is a specification setting method for a vibration damping device provided between a first part and a second part of a structure that moves relatively by disturbance,
Between the first part and the second part, a rotary inertia mass damper that generates a rotary inertia mass by rotation of a mass body and a damping damper are arranged in parallel, and the rotary inertia mass damper arranged in parallel and The damping damper and the spring are arranged in series,
The natural circular frequency determined by the rotational inertial mass and the spring is ωd,
The natural circular frequency of the structure is ω,
When the magnification of the damping coefficient of the damping damper is αc,
ω / ωd <1
and,
αc> 1
The specifications setting method of the damping device which sets the specifications of the said rotary inertia mass damper and the said spring so that it may become.   The specification setting method of the damping device of Claim 12 which sets the specification of the said rotary inertia mass damper and the said spring so that the magnification (alpha) c of the said damping coefficient may become the maximum.

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