JP2017003089A - Rotational inertia mass device and vibration control structure including the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a rotational inertia mass device capable of effectively avoiding resonance phenomenon due to input of periodic vibration energy, and a vibration control structure including the rotary inertia mass device.SOLUTION: A rotational inertia mass device includes a displacement dependence type adjustment mechanism 2 configured to change rigidity according to displacement, and adjusts rigidity by the displacement dependence type adjustment mechanism 2 when periodic vibration energy is input, so that a natural period changes. Also, the displacement dependence type adjustment mechanism 2 includes a nonlinear rotational inertia element 4 and strong nonlinear rigidity element 3, and the nonlinearities of the nonlinear rotational inertia element 4 and the strong linear rigidity element 3 are dependent on displacement.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a rotary inertia mass device and a vibration control structure including the same.

  When resonance occurs due to long-period ground motion, etc., a large response occurs in the structure. For this reason, in the past, buildings (structures) such as condominiums and office buildings have vibration damping dampers installed in the buildings, and this vibration damping damper absorbs and attenuates the seismic energy (vibration energy) that acts during an earthquake. To reduce the response. Such damping dampers include hysteresis dampers that use the yield strength of steel and frictional resistance of sliding materials, viscous dampers such as oil dampers that use the viscous resistance of viscous materials, and shearing of viscoelastic materials. Viscoelastic dampers using resistance are often used.

  Here, FIG. 25 is a simulation result (resonance curve) for a building that receives a periodic external force, the horizontal axis is the ratio of the frequency of the input wave (sine wave) to the natural period, and the vertical axis is the dynamic analysis result. The dynamic response magnification which normalized the displacement with the displacement when the load 1 is applied statically is shown. In addition, among the three curves in the figure, the center curve indicates the dimensions of the structure m (mass) = 1, k (rigidity) = 1, c (damping coefficient) = 0.04 (damping h = c). /(2√(mk)=0.02) The simulation result of the basic model is shown in the left curve of the basic model when the mass is increased with respect to the basic model (m = 4). The curve on the right shows the result when the stiffness is increased (k = 9).

  From FIG. 25, it was confirmed that a large dynamic response magnification is obtained when resonating in the linear range. On the other hand, if the damping damper is increased and the attenuation h is increased, the peak can be suppressed. However, since the peak of the resonance point is determined by 1 / (2h), it is necessary to install and add a very large number of damping dampers. In reality, the resonance phenomenon can be avoided due to the installation space and the like. In many cases, it is difficult to increase the attenuation.

  On the other hand, as another means for reducing the earthquake response of the building, it has been proposed and put into practical use that a damping device called TMD (Tuned Mass Damper) is installed on the top side of the building (such as the rooftop) ( For example, see Patent Document 1 and Patent Document 2).

  Specifically, TMD reduces the response of the building by moving the seismic energy acting on the building to the TMD side when the added mass synchronized with the primary natural period of the building vibrates in the direction opposite to that of the building. Can be made.

  In addition, a method of reducing the response by installing a seismic isolation device and shifting the primary natural period of a structure such as a building to the long period side is often used (for example, see Patent Document 3).

JP 2000-18323 A JP 2011-220511 A JP 2010-070909 A

However, in order to suppress the resonance by installing TMD as a vibration control device on the roof of a building, an additional mass of about 1% of the entire building is required, and it is necessary to secure a large space on the roof. There are also many.
Furthermore, when applied to a building having different natural periods in two directions, there is a demerit that a TMD having different tuning periods in each direction is required.

  In addition, in seismic isolation structures, it is necessary to study resonance when long-period component waves continue for a long time during an earthquake, such as on soft ground or in a basin.

  The present invention has been made in view of the above circumstances, and an object thereof is to provide a rotary inertia mass device that can effectively avoid a resonance phenomenon caused by the input of periodic vibration energy and a damping structure including the same. To do.

  In order to achieve the above object, the present invention provides the following means.

  The rotary inertia mass device of the present invention includes a three-dimensional cam mechanism that converts linear motion into rotary motion, and the relationship between displacement and rotary inertia mass effect is nonlinear by the three-dimensional cam mechanism.

  The vibration damping structure of the present invention includes the above-described rotary inertia mass device, and is configured such that the natural period changes according to the displacement.

  Moreover, it is desirable that the vibration damping structure of the present invention includes the rotary inertia mass device described above and a variable stiffness device whose stiffness changes according to displacement.

  According to the rotational inertial mass device and the vibration damping structure of the present invention, when periodic vibration energy is input, a nonlinear rotational inertial mass effect is exhibited, and the natural period of the vibration damping structure can be changed. it can. Thereby, it is possible to effectively avoid the resonance phenomenon caused by the input of periodic vibration energy.

It is a figure which shows the model of the damping structure which concerns on one Embodiment of this invention. It is a figure which shows an example of the resonance curve of the damping structure which concerns on one Embodiment of this invention, and is a figure which shows the resonance curve of the type 1 of FIG. It is a figure which shows the example of a displacement dependence setting for the mass fluctuation | variation of a strong nonlinear rigidity element and a nonlinear rotational inertia element. It is a figure which shows the resonance curve of the type 2 of FIG. It is a figure which shows the resonance curve of the type 3 of FIG. It is a figure which shows the resonance curve of a Duffing equation. It is a figure which shows the nonlinear rotational inertia element (nonlinear rotational inertial mass apparatus) of the 1st example which concerns on one Embodiment of this invention, and is a figure which shows the state which the variable inertial mass body rotated and opened. FIG. 8 is a view taken along line X1-X1 in FIG. 7. It is a figure which shows the nonlinear rotational inertia element (nonlinear rotational inertial mass apparatus) of the 1st example which concerns on one Embodiment of this invention, and is a figure which shows the state which the variable inertial mass body closed. FIG. 10 is a view taken along line X1-X1 in FIG. 9. It is a figure which shows the shape of the cam groove of the solid cam mechanism (linear motion conversion mechanism) of the nonlinear rotating inertial element of the 1st example which concerns on one Embodiment of this invention. It is a figure which shows the relationship between the fluctuation | variation part of rotational inertial mass, and an axial displacement. It is a figure which shows the relationship between the variation | change_quantity of a variable rotation inertial mass, and the horizontal direction displacement of a linear slider. It is a figure which shows the shape of the cam groove of the nonlinear rotation inertia element corresponding to the type 1 of FIG. It is a figure which shows the shape of the cam groove of the nonlinear rotation inertia element corresponding to the types 2 and 3 of FIG. It is a figure which shows the nonlinear rotational inertia element (nonlinear rotational inertia mass apparatus) of the 2nd example which concerns on one Embodiment of this invention. It is a figure which shows the strong nonlinear rigidity element (variable rigidity apparatus) which concerns on one Embodiment of this invention. It is a figure which shows the restoring force characteristic of a disk spring single-piece | unit. It is a figure which shows the shape of the cam groove of the solid cam mechanism (linear motion conversion mechanism) of the strong nonlinear rigidity element which concerns on one Embodiment of this invention. It is a figure which shows the relationship between the load and axial displacement which generate | occur | produce. It is the figure used in description which determines the pitch of a thread groove. It is a figure which shows the shape of the cam groove of the strong nonlinear rigidity element corresponding to the type 1 of FIG. It is a figure which shows the shape of the cam groove of the strong nonlinear rigidity element corresponding to the type 2 of FIG. It is a figure which shows the shape of the cam groove of the strong nonlinear rigidity element corresponding to the type 3 of FIG. It is a figure which shows the resonance curve of a linear model.

  Hereinafter, a rotary inertia mass device according to an embodiment of the present invention and a damping structure including the same will be described with reference to FIGS. 1 to 25.

  First, when the vibration damping structure (damping structure) of the present embodiment is displaced by vibration energy such as seismic energy, the rigidity is adjusted (adjusted) according to the displacement and the natural period (natural vibration) The period is changed, and the resonance phenomenon of the structure with respect to the input of the periodic vibration energy is avoided.

  And the damping structure 1 of this embodiment is provided with the displacement dependence type | mold adjustment mechanism 2 which changes rigidity and rotational inertial mass according to a displacement, as shown in FIG. 2 includes a strongly nonlinear stiffness element (Fk) 3 whose nonlinearity depends on displacement and a nonlinear rotational inertia element (Fm) 4.

  That is, in the vibration damping structure 1 of the present embodiment, by adding the displacement-dependent adjusting mechanism 2 including the strong nonlinear stiffness element 3 and the nonlinear rotational inertia element 4 to the above-described basic model shown in FIG. Resonance characteristics are obtained by changing the shape of the resonance curve of the linear model shown in FIG. 25 so as to be bent as shown in FIG. 2, and when the periodic vibration energy is input, the resonance is caused only to the smallest dynamic response magnification. It does not occur.

  Here, FIG. 3 shows a displacement-dependent setting example of the mass fluctuation of the strong nonlinear stiffness element 3 and the nonlinear rotational inertia element 4 of the displacement dependent adjustment mechanism 2.

  In FIG. 3, the horizontal axis indicates the displacement (standardized by the static displacement of the basic model). Further, the vertical axis of the diagram related to the strong nonlinear stiffness element 3 represents the generated load (Fk), and the vertical axis of the diagram related to the nonlinear rotational inertia element 4 represents the mass fluctuation (Ma) of the nonlinear rotational inertia element 4. Yes.

  And in the damping structure 1 of this embodiment, the displacement dependence type | mold adjustment mechanism 2 is provided, and it is made to have the characteristic of the displacement dependence curve of following (a), (b) shown in FIG.

  (A) Although the strong nonlinear stiffness element 3 does not function at the initial stage, the function starts to be exerted from the middle indicated by the displacement Ug in FIG. 3 and reaches a large value fp at the displacement Up. Thereafter, there are a case where the load drops to zero, a case where the function is maintained at fp, a case where the load is increased due to secant rigidity at the displacement Up, and the like.

  (B) The nonlinear rotational inertia element 4 functions sufficiently at the initial stage (maximum value ma), and its function is reduced from the middle indicated by the displacement Ug in FIG. 3, and the minimum value (in FIG. 0). After that, there are cases where the original function is restored and cases where the minimum value is maintained.

  In other words, in the vibration damping structure 1 of the present embodiment, when the displacement is small, the action of the strong nonlinear stiffness element 3 is relatively small, the action of the nonlinear rotational inertia element 4 is large, and the displacement has a certain value (Up When it becomes large (near Up), the action of the strong nonlinear stiffness element 3 is large, and the action of the nonlinear rotational inertia element 4 is small.

  By providing each of the features (a) and (b) as described above, and by providing these features together, if the displacement is Ug or less, the structure is compared with the basic model shown in FIG. Longer period. Further, when the displacement increases from Ug to Up, the instantaneous natural period of the entire structure is rapidly shortened. That is, the vibration characteristics change abruptly from the long-period structure to the short-period structure.

  Note that. In this embodiment, the segmented cubic curves shown in type 1, type 2, and type 3 shown in FIG. 3 are used. However, if the natural period (natural frequency) of the structure can be changed by displacement, The displacement-load relationship and the displacement-rotation inertial mass variation relationship may be linear, stepwise, quadratic or the like.

  Then, the displacement-dependent adjustment mechanism 2 including the strong nonlinear stiffness element 3 and the nonlinear rotational inertia element 4 as described above is provided, and the simulation result of the resonance curve when the type 1 characteristic of FIG. FIG. 4 shows the simulation results when the characteristics of type 2 are provided, and FIG. 5 shows the simulation results when the characteristics of type 3 are provided.

  2, 4, and 5, the solid line indicates the stable periodic solution, and the broken line indicates the unstable periodic solution. Note that the response value usually converges to the stable periodic solution.

  In the case of a strongly nonlinear structure, it is known that a branching phenomenon occurs in the resonance curve. In FIGS. 2, 4 and 5, ○ is a saddle node branch, ● is a pitch fork branch, and ◎ is a double period branch. It has become a point.

  When there are a plurality of stable periodic solutions for one frequency ratio, the periodic solution to be converged is determined by the initial conditions.

That is, when the state where the structure is stationary is set to the initial value, it converges to the periodic solution with the smallest dynamic response magnification. Specifically, in FIG. 2, the stable periodic solution can be roughly divided into regions indicated by (1), (2), (3), and (4). Converge in the area of (2).
Thereby, it was confirmed that the dynamic response magnification can be significantly suppressed in the vibration damping structure 1 of the present embodiment as compared with the linear model shown in FIG.

  Further, in (2), there is a region where the dynamic response magnification is slightly increased due to branching, but the dynamic response magnification is 2 or less, and it was confirmed that an excellent response reduction effect was obtained.

  Many branching phenomena occur in the area (4), but if the initial state is a state in which the structure is stationary, the response magnification does not increase to the area (4) and can be ignored.

  On the other hand, in the case of sweep excitation to a high frequency, the dynamic response magnification gradually increases from the region (2) to the region (3), but the frequency ratio β is around 2.5. By jumping to the area (1), the response magnification is drastically reduced. As a result, there is a very low possibility that the situation will converge from the stationary state to the region (4). In addition, when a large initial value is given, there is a possibility of being converged in the areas (3) and (4), but the areas (1) and (2) are more than the areas (3) and (4). Since the stability is higher, the response is reduced to the periodic solutions (1) and (2) with a slight disturbance even in this case.

  Next, in type 2, as shown in FIG. 4, since the gradient of the region corresponding to (4) is large, the frequency range that resonates in (4) is narrow. In type 3, as shown in FIG. 5, the resonance range is limited to the vicinity of the frequency ratio β = 3.

  In Type 3, compared with the case of k = 9 in the linear model shown in FIG. 25, the resonance curve in the vicinity of the frequency ratio β = 3 is similar, but in the linear model, it is added at the frequency ratio β = 3. Resonating to the peak of the resonance curve when shaken, the type 3 is decisively different in that it converges in the region (1) (almost zero in the figure).

Here, the Duffing equation of the following equation (1) is a dynamic system in which a term of αω ² × ³ is added to the linear motion equation, and is a well-known equation. Ω is ω = √ (k / m), h is the damping of the structure (h = c / 2√ (mk)), α is the nonlinear spring constant of the structure, p is the external force magnification, β is the structure The frequency ratio of forced external force to ω, m is the mass of the structure, c is the damping coefficient of the structure, and k is the rigidity of the structure.

  FIG. 6A shows an example of a resonance curve according to the equation (1). FIG. 6B is shown in Reference Document 1 (Koichi Watanabe, Shoichi Nakai: Study on vibration characteristics of one-degree-of-freedom system having strong nonlinearity, Architectural Institute of Japan (Kinki), September 14, 2014). It is the resonance curve by the Duffing equation which considered the yielding which was done.

  It is a well-known fact described in many documents that the resonance curve is inclined as shown in FIG. 6 when a term including strong nonlinearity is included in the dynamical system.

  Also, when the resonance curve is tilted, there are multiple periodic solutions for the same frequency ratio β, which periodic solution is converged depends on the initial state of the structure, and the initial value is stationary In this case, it is also known that the dynamic response magnification converges to the periodic solution.

  Further, when the period of the excitation force is changed (sweep excitation), if the sweep is made to the high frequency side, the dynamic response magnification is increased from D → A in FIG. 6 to the low frequency side. It is also known that when the sweep is performed, a path different from C → E → D is followed.

  In addition, as described in Reference 1 above, considering the yield in the strong nonlinear stiffness element 3 having a hardening-type restoring force characteristic is naturally considered first when the design around the device (2) is considered. Is the item to be.

  On the other hand, as shown in Reference Document 1, when the initial value is not zero at the frequency near E in FIG. 6, there is a case where it converges to D and the dynamic response magnification increases.

  On the other hand, in the vibration damping structure 1 and the displacement dependent adjustment mechanism 2 of the present embodiment, by incorporating the nonlinear rotational inertia element (rotational inertial mass device / variable inertial mass device of the present embodiment) 4, point E Is shifted to the long period side (the direction in which the frequency ratio β decreases).

  That is, when the nonlinear rotational inertia element 4 is incorporated, the area portion of (1) in FIG. 2 having a small dynamic response magnification can be expanded to bite into the long period side. Further, in type 3 of FIG. 3, the frequency that is highly likely to resonate can be limited to a narrow range ((4) of FIG. 5), and the possibility of occurrence of the resonance phenomenon can be significantly reduced. it can.

  Therefore, in the damping structure 1 of the present embodiment, when periodic vibration energy is input, a nonlinear rotational inertia mass effect is exhibited, and the natural period of the damping structure 1 can be changed. . Thereby, it is possible to effectively avoid the resonance phenomenon caused by the input of periodic vibration energy. This is particularly effective when a cyclic load close to the natural period of the structure 1 acts for a long time.

  The displacement-dependent adjusting mechanism 2 (variable inertial mass device 4 and variable stiffness device 3) of this embodiment having such characteristics is effective when applied to a seismic isolation structure. By adding the displacement-dependent adjustment mechanism 2 of the embodiment, it is possible to suppress resonance even when a large number of long-period components are repeatedly excited during an earthquake. It can also be installed instead of a stopper device in the event of a large earthquake.

  Further, when the displacement-dependent adjusting mechanism 2 of the present embodiment is added in addition to the seismic isolation structure, resonance can be suppressed without installing many damping dampers as in the prior art.

  Furthermore, when a plurality of vibrations including a large number of specific frequency components, such as mechanical vibrations and vibrations during railway running, are assumed, the displacement-dependent adjustment mechanism 2 of this embodiment is added to effectively resonate the resonance phenomenon. Can be suppressed.

  Moreover, the resonance of the surrounding ground can be suppressed by adding the displacement-dependent adjustment mechanism 2 of the present embodiment to the basic portion of the excitation source such as a large shaking table. Furthermore, the present invention can be applied to the suppression of resonance of structures and bridges that receive strong winds.

  Here, specific examples of the strong nonlinear stiffness element 3 and the nonlinear rotational inertia element 4 will be described.

  First, the nonlinear rotational inertial element (nonlinear rotational inertial mass device / variable inertial mass device) 4 of the first example of the present embodiment is a rotational inertial mass device capable of changing the rotational inertial mass effect depending on the displacement, In other words, it is a rotary inertia mass device having nonlinear rotational inertia characteristics, and includes a rotary inertia mass mechanism 5 and a nonlinear characteristic adjustment mechanism (rotational inertia mass variable mechanism) 6 as shown in FIGS. Has been.

  The rotary inertia mass mechanism 5 includes a screw shaft 7 provided with a central axis O1 coaxially arranged with the axis O1 of the apparatus, and a first rotary weight 8 screwed to the screw shaft 7. ing.

  A connecting member 9 such as a ball joint or a clevis is attached to one end (base end) of the screw shaft 7. The screw shaft 7 has a first screw groove (male screw) 10 to which the first rotary weight 8 is screwed on the outer peripheral surface thereof, and the other end (tip) side of the first screw groove 10. A second thread groove (male thread) 11 into which a later-described second rotary weight 14 is screwed is formed on the outer peripheral surface of the first thread. Further, the first screw groove 10 has a small screw pitch and is formed densely. The second screw groove 11 has a larger screw pitch than the first screw groove 10 and is formed sparsely.

  In the rotary inertia mass mechanism 5, the screw shaft 7 is supported by the casing 12 so as to be able to advance and retreat in the direction of the axis O1, and one end side is arranged outward from the casing 12 to form the first screw groove 10 and the second screw groove 11. This portion is provided in the casing 12. Further, a first rotating weight 8 screwed into the first screw groove 10 is disposed in the casing 12, and a substantially disc-shaped first rotating weight 8 is integrally attached to the casing 12 on the outer peripheral side thereof. 13 is rotatably supported and disposed.

  As a result, in the rotary inertia mass mechanism 5, when the screw shaft 7 moves back and forth in the direction of the axis O1 in accordance with the displacement of the structure 1 caused by the action of vibration energy, The first rotating weight 8 screwed into the first thread groove 10 rotates, and an inertial mass effect several thousand times as large as the mass of the first rotating weight 8 can be generated.

  The nonlinear characteristic adjusting mechanism (rotational inertial mass variable mechanism) 6 includes a second rotating weight 14 that is screwed into the second screw groove 11 of the screw shaft 7.

  Thereby, when the screw shaft 7 moves forward and backward in the direction of the axis O1 according to the displacement of the structure 1 caused by the action of vibration energy, the second screw groove together with the first rotary weight 8 according to the advancement and retraction of the screw shaft 7. The second rotary weight 14 screwed to 11 rotates.

  At this time, the first rotary weight 8 is screwed into the dense first screw groove 10, and the second rotary weight 14 is screwed into the sparse second screw groove 11, so that the first The rotary weight 8 rotates at a high speed, and the second rotary weight 14 rotates at a low speed.

  The sparse second screw groove 11 corresponding to the second rotary weight 14 has a plurality of grooves (four in this embodiment) arranged in parallel to ensure the transmission of force, so that the number of turns is insufficient. It is configured to compensate.

  Further, the nonlinear characteristic adjusting mechanism 6 changes the rotational mass effect by the action of the linear motion conversion mechanism 15 that converts the rotational motion of the second rotary weight 14 into the linear motion along the direction of the axis O1 and the linear motion conversion mechanism 15. And a variable rotation inertial mass mechanism 16.

  The linear motion conversion mechanism 15 of the present embodiment is a three-dimensional cam mechanism, and includes a cylindrical cam 20 and a driven body 21.

  The cylindrical cam 20 has a series of cam grooves formed on the outer peripheral surface. Further, the cylindrical cam 20 of the present embodiment is formed as a part of the second rotary weight 14 while the screw shaft 7 is inserted into the inner hole and the axis O1 is made coincident.

  The follower 21 is formed in a substantially cylindrical shape, and is supported by a linear slider 22 provided integrally with the casing 12 so as to be movable back and forth along the direction of the axis O1. The follower 21 is provided coaxially so as to contain the cylindrical cam 20, and is provided with a claw engaged with the cam groove of the cylindrical cam 20. As a result, the driven body 21 advances and retreats in the direction of the axis O1 as the cylindrical cam 20 rotates around the axis O1 and is guided by the cam groove.

  That is, in the linear motion conversion mechanism 15 of the present embodiment, the rotational motion of the second rotary weight 14 (cylindrical cam 20) is converted into the linear motion of the follower 21.

  The variable rotation inertial mass mechanism 16 includes a rack and pinion 23 and a variable inertial mass body 27.

  The rack and pinion 23 is supported by the first rotary weight 8 so as to be able to advance and retreat in the direction of the axis O1. The rack and pinion 23 is rotated around the axis O1 together with the first rotary weight 8, and the rotation axis is directed in a direction perpendicular to the axis O1. A rotary shaft 25 rotatably supported around the rotation axis on the one-rotation weight 8 and a linear motion that is integrally attached to the rotary shaft 25 and meshed with the rack 24 so that the rack 24 advances and retreats. And a pinion 26 for converting into the rotational motion of

Further, an urging means such as a rotation spring is attached to the rotating shaft 25 and is urged so as to always rotate in one direction. Thereby, the pinion 26 is urged in one direction, and the rack 24 is urged so as to move to the B side of the screw shaft 7 in FIG.

  The variable inertia mass body 27 has one end fixed to the rotating shaft 25 of the rack and pinion 23 and extends in a direction perpendicular to the rotating axis.

  In the rotary inertia mass device of the first example of the nonlinear rotary inertia element 4 having the above-described configuration, when the screw shaft 7 advances and retreats in the direction of the axis O1, the second rotary weight 14 and the first rotary weight 8 rotate. The follower 21 of the linear motion conversion mechanism 15 of the three-dimensional cam mechanism moves forward and backward in the direction of the axis O1 in conjunction with the second rotary weight 14.

  7 and 8, when the driven body 21 moves to the B side of the screw shaft 7 in FIG. 7, the force that the driven body 21 pushes the rack 24 of the variable rotation inertial mass mechanism 16 becomes weaker. At the same time, the rotation shaft 25 of the rack and pinion 23 is rotated by the urging means, and the variable inertia mass body 27 is rotated around the rotation axis to move so as to open the umbrella.

  Further, as shown in FIGS. 9 and 10, when the driven body 21 moves to the screw shaft 7 side of FIG. 7A, the rotating shaft 25 of the rack and pinion 23 is pressed by the driven body 21, and the variable inertia mass body 27 is moved. Move to close the umbrella.

  Thus, the variable inertial mass body 27 can be opened and closed to change the rotational inertial mass effect.

  In addition, the number of claws of the follower 21 of the three-dimensional cam mechanism can be one or plural, but in this embodiment, the claws are provided at four locations, top, bottom, left, and right. In this case, as shown in FIG. 11, the lead length of the cam groove 28 of the cylindrical cam 20 is set so as to be within ± 45 ° rotation with respect to the movement of the screw shaft 7 in the axis O1 direction. That is, the shape of the cam groove 28 of the three-dimensional cam mechanism is determined so that a predetermined displacement in the direction of the axis O1 can be realized according to this ± 45 ° rotation.

In order to realize the type 1, type 2, and type 3 of the nonlinear rotational inertia element (Fm) 4 shown in FIG. 3, the shape of the cam groove 28 is as shown in FIG.
Further, even when the umbrella shown in FIGS. 9 and 10 is closed, there is some rotational inertial mass. If the rotational inertial mass at that time is regarded as zero rotational mass, the type of FIG. 3 can be realized. It will be.

  Further, the relationship between the movement of the follower 21 in the direction of the axis O1 of the linear motion conversion mechanism (three-dimensional cam mechanism) 15 and the rotational inertial mass by the variable inertial mass 27 can be obtained by numerical calculation and represented as a chart or the like. is there.

  Specifically, the geometric position of the variable inertia mass body 27 can be determined from the displacement of the driven body 21, and thereby the variable rotational inertia mass by the variable inertia mass body 27 can be calculated. For example, in order to realize the cubic curve of FIG. 3, the displacement of the variable inertia mass body 27 is obtained in reverse from the necessary rotational inertia mass, and the cam groove 28 and the like are formed so as to realize the displacement. That's fine.

  More specifically, an example of a method for determining the shape of the cam groove 28 for realizing the characteristic of nonlinear rotational inertia mass fluctuation will be described.

  When Type 1, Type 2 and Type 3 in FIG. 3 are shown in an overlapping manner, the relationship between the rotational inertial mass variation and the axial displacement is as shown in FIG.

  As a procedure for determining the shape of the cam groove 28, first, the relationship between the variable inertia mass variation (variable inertia mass body 27) of the first rotary weight 8 and the horizontal displacement of the linear slider 22 is obtained. The horizontal displacement of the linear slider 22 means the displacement of the driven body 21 and the rack 24.

  If the horizontal displacement of the linear slider 22 is determined, the angle at which the variable inertia mass 27 is opened in an umbrella shape can be geometrically calculated from the gear ratio of the rack 24.

Thereby, the relationship between the horizontal displacement of the linear slider 22 and the rotational inertial mass of the variable portion by the variable inertial mass body 27 as shown in FIG. 13 can be obtained.
A, B, C, D, E, and F shown in FIG. 13 are target points for realizing the rotational inertia mass of type 1 in FIG. That is, when the displacement in the direction of the axis O1 in FIG. 12 is Ug, U1, Up, U2, U3, U4, the cam groove 28 may be set so that the horizontal displacement of the linear slider 22 is as shown in FIG. .

  FIG. 14 shows a setting example of the cam groove 28 in the case of type 1. The position C in the figure corresponds to the horizontal displacement of L2 in FIG. Further, the positions of B and D in the figure correspond to the horizontal position of L1 in FIG. FIG. 14B shows the shape of the cam groove 28 from 0 degree to 45 degrees of the second rotary weight 14 for controlling the three-dimensional cam, but 0 degree to −45 degrees has an axis of 0 degree. It becomes the line symmetrical shape.

  FIG. 15 shows a setting example of the cam groove 28 in the case of type 2 and type 3. The position C in the figure corresponds to the horizontal displacement of L2 in FIG. Further, the position B in the figure corresponds to the horizontal position L1 in FIG. Since it is sufficient to maintain the state of C after the point C, the position of the cam groove 28 is as shown in FIG. FIG. 15 (b) shows the shape of the cam groove 28 from 0 degree to 45 degrees of the second rotary weight 14 for controlling the three-dimensional cam. A line-symmetric shape with 0 degree as an axis.

  Next, the nonlinear rotational inertia element (nonlinear rotational inertial mass device) 4 of the second example of this embodiment will be described. The non-linear rotational inertial element 4 of the second example is also a non-linear rotational inertial mass device capable of changing the rotational inertial mass effect in accordance with the displacement as in the first example.

  On the other hand, the nonlinear rotational inertia element 4 of the first example is configured so that the rotational inertia mass-displacement relationship can be expressed by an arbitrary curve, but this relationship may be stepped. For this reason, the non-linear rotational inertial element 4 as the second example is configured so that the rotational inertial mass-displacement relationship can be expressed in a step shape.

  Specifically, the nonlinear rotational inertia element 4 of the second example of the present embodiment includes a rotational inertial mass mechanism 5 and a nonlinear characteristic adjustment mechanism (rotational inertial mass variable mechanism) 6 as shown in FIG.

  The rotary inertia mass mechanism 5 is configured in the same manner as in the first example. That is, the first rotary weight 8 is screwed onto the screw shaft 7 provided with the central axis O1 coaxially arranged with the damper axis O1.

  A first screw groove 10 into which the first rotary weight 8 is screwed and a second screw groove 11 into which the second rotary weight 14 is screwed are formed on the outer peripheral surface of the screw shaft 7. The first screw groove 10 has a small screw pitch and is formed densely. The second screw groove 11 has a larger screw pitch than the first screw groove 10 and is formed sparsely.

  The nonlinear characteristic adjusting mechanism (rotational inertial mass variable mechanism) 6 includes a second rotating weight 14 screwed into the second thread groove 11.

  Further, the nonlinear characteristic adjusting mechanism 6 changes the rotational mass effect by the linear motion converting mechanism 15 that converts the rotational motion of the second rotary weight 14 into the linear motion along the direction of the axis O1 and the linear motion converting mechanism 15 acting. And a variable rotation inertial mass mechanism 30 to be operated.

  The linear motion conversion mechanism 15 is a three-dimensional cam mechanism, and includes a cylindrical cam 20 and a driven body 21 as in the first example, and the cylindrical cam 20 rotates around the axis O1 and is guided by the cam groove 28 to follow the driven body. 21 moves forward and backward in the direction of the axis O1. Further, a pressing mechanism 31 composed of a spring 31a and a ball 31b is provided on the distal end side of the driven body 21.

  In the variable rotation inertial mass mechanism 30 of the second example of the present embodiment, when the driven body 21 moves in the direction of the axis A, the driven body 21 is pressed against the first rotating weight 8 and the first rotating weight 8 is pressed. An additional rotary inertial mass 32 that rotates with the rotation is provided.

  Further, a friction member 33 is provided between the additional rotary inertia mass body 32 and the first rotary weight 8, and the follower 21 moves in the direction of the axis A so that the additional rotary inertia mass body 32 is attached to the first rotary weight 8. It is configured to be rotated integrally with the first rotary weight 8 by being pressed and friction with the friction member 33.

  Then, when the driven body 21 moves in the A side direction in accordance with the advancement and retreat of the screw shaft 7 in the axis O1 direction, the rotational inertia effect of the additional rotating inertia mass body 32 is added, and when the driven body 21 moves in the B side direction, friction occurs. And the rotational inertia effect of the additional rotational inertia mass 32 is lost.

  In this example, the rotational inertia effect is in two stages, ON and OFF, but the rotational inertia mass can be increased or decreased according to the displacement. Further, the nonlinear rotational inertia element (rotational inertial mass device having nonlinear characteristics) 4 using this friction is more compact than the nonlinear rotational inertial element 4 of the first example, and can realize a large additional rotational inertial mass.

  In addition, by arranging a plurality of devices 4 in which the rigidity of the pressing mechanism 31 and the friction coefficient of the friction member 33 are adjusted to various levels, for example, on the floor of the building, the entire layer of the building (structure 1) has a multi-stepped shape. It is also possible to realize the rotational inertial mass-displacement relationship.

  Next, the strongly nonlinear stiffness element (variable stiffness device having nonlinear stiffness characteristics) 3 according to the present embodiment is configured to include a stiffness adjusting mechanism 35 and a curing type stiffness mechanism 36 as shown in FIG. .

  The rigidity adjusting mechanism 35 includes a screw shaft 37 provided with the central axis O2 coaxially arranged with the axis O2 of the device, a first rotating body 38 provided by being screwed to one end side of the screw shaft 37, And a second rotating body 39 that is screwed to the other end of the screw shaft 37.

  A connecting member 40 such as a ball joint or a clevis is attached to one end (base end) of the screw shaft 37. The screw shaft 37 has a first screw groove (male screw) 41 into which the first rotating body 38 is screwed on the outer peripheral surface thereof, and the other end (tip) side of the first screw groove 41. A second screw groove (male screw) 42 into which the second rotating body 39 is screwed is formed on the outer peripheral surface of the second screw 39.

  Further, the screw shaft 37 is supported by the casing 43 so as to be able to advance and retreat in the direction of the axis O2, and one end side from the casing 43 is arranged outside, and a portion where the first screw groove 41 and the second screw groove 42 are formed is formed in the casing 43. It is arranged and arranged. Further, the first rotating body 38 screwed into the first screw groove 41 and the second rotating body 39 screwed into the second screw groove 42 are respectively disposed in the casing 43 and are substantially disk-shaped first. The rotating body 38 and the second rotating body 39 are rotatably supported by bearings 44 integrally attached to the casing 43 on the outer peripheral side.

  As a result, when the screw shaft 37 moves forward and backward in the direction of the axis O2 according to the displacement of the structure 1 caused by the action of vibration energy, the first rotating body 38 and the second rotating body according to the forward and backward movement of the screw shaft 37. 39 rotates about axis O2.

  In addition, the stiffness adjusting mechanism 35 converts the rotational motion of the first rotating body 38 into a linear motion along the direction of the axis O2, and the nonlinear motion adjusting mechanism 45 that converts the rotational motion of the second rotating body 39 along the direction of the axis O1. And a non-linear characteristic adjusting mechanism 46 for converting to a linear motion.

  Each of the nonlinear characteristic adjusting mechanisms 45 and 46 is a three-dimensional cam mechanism, and includes a cylindrical cam 47 and a driven body 48.

  Also in this embodiment, the 1st rotary body 38 and the 2nd rotary body 39 each serve also as the cylindrical cam 47, and a series of cam grooves are formed on the outer peripheral surface of each rotary body 38, 39.

  The follower 48 is formed in a substantially cylindrical shape, and is supported by a linear slider 49 provided integrally with the casing 43 so as to be movable back and forth along the direction of the axis O2. The follower 48 is provided coaxially so as to contain the cylindrical cam 47, and is provided with a claw engaged with the cam groove of the cylindrical cam 47. As a result, the driven body 48 advances and retreats in the direction of the axis O2 as the first rotary body 38 and the second rotary body 39, which are cylindrical cams 47, rotate around the axis O2 and are guided by the cam grooves.

  The curable rigid mechanism 36 is configured using a spring member. In the present embodiment, the curable rigid mechanism 36 is configured so that a plurality of disc springs 50 are stacked side by side.

  The screw shaft 37 is integrally provided with a disk-like pressing jig 51 at the center between one end side and the other end side. A first group of disc spring bodies 52 is provided between the pressing jig 51 and the driven body 48 of the one nonlinear characteristic adjusting mechanism 45, and is between the pressing jig 51 and the driven body 48 of the other nonlinear characteristic adjusting mechanism 46. The second group of disc spring bodies 53 are interposed respectively. In the present embodiment, in the initial state, there are gaps Ug between the pressing jig 51 and the first group of disc spring bodies 52, and between the pressing jig 51 and the second group of disc spring bodies 53, respectively. It is comprised so that it may be formed.

  In this strongly nonlinear rigid element 3, when the screw shaft 37 moves forward and backward in the direction of the axis O2, the pressing jig 51 integrated with the screw shaft 37 also moves in the direction of the axis O2, and the disc spring body 51 in the moved direction, A reaction force is obtained by compressing 52.

  At this time, two nonlinear characteristic adjusting mechanisms 45 and 46 are provided on both ends, and the movement of the screw shaft 37 in the direction of the axis O2 is converted into the movement of the first rotating body 38 and the second rotating body 39 in the rotating direction. . Further, each rotary body 38, 39 is provided with a cam groove for constituting a three-dimensional cam mechanism, and a linear motion in the direction of the axis O2 of the follower 48 in which the rotational displacement of each rotary body 38, 39 is engaged with the cam groove. Convert to (linear displacement).

Thereby, the space | interval of the pressing jig 51 and each follower 48 changes according to the displacement amount of the axial direction O2 of the screw shaft 37, and the rigidity of the strong nonlinear rigidity element 3 becomes variable. Therefore, the displacement-load relationship after the displacement Up in FIG. 3 can be realized.
In addition, in the displacement-load relationship of the strong nonlinear stiffness element 3 in FIG. 3, the displacement Up is common to Type 1, Type 2, and Type 3, and can be realized by combining a plurality of disc springs 50.

  Further, in the present embodiment, the number of claws of the follower 48 of the three-dimensional cam mechanism can be one or more. For example, the claws are provided at four places on the top, bottom, left, and right. In this case, the lead length of the cam groove of the cylindrical cam 47 is set so as to be within ± 45 ° or less with respect to the movement of the screw shaft 37 in the direction of the axis O2. That is, the shape of the cam groove of the three-dimensional cam mechanism is determined so that a predetermined displacement in the direction of the axis O2 can be realized according to this ± 45 ° rotation.

  Here, FIG. 18 shows an example of the restoring force characteristic of the disc spring alone. A linear restoring force characteristic is exhibited up to a range where the deflection amount f / h with respect to the bending height is about 0.75, and then the rigidity increases and the disc spring 50 is completely crushed (f / h = 1). .0), the stiffness is ∞.

  In order to ensure the durability of the disc spring 50, it is desirable to suppress f / h to 0.75 (75%) or less. On the other hand, for example, when the gap stopper is operated in a seismic isolation structure, the rigidity of the gap portion needs to approach ∞ as the interlayer displacement increases.

  Therefore, in this embodiment, as shown in FIG. 17, f / h is set to 0.75 (75%) by interposing a deflection limiting jig (displacement limiting jig) 55 between the adjacent disc springs 50. ) The rigidity can be set to ∞ while making the following. Then, by connecting a plurality of disc springs 50 having different rigidity or adjusting the thickness of the deflection limiting jig 55, it is possible to realize a smoother curable restoration characteristic. Further, by providing the gap Ug between the disc spring 50 and the pressing jig 51, it becomes possible to realize a gap (a region where no load is generated).

Further, the shape of the cam groove 57 for realizing the type 1, type 2 and type 3 of the strong nonlinear stiffness element 3 shown in FIG. 3 is as shown in FIG. 19, for example.
Further, the relationship between the movement of the follower 48 in the direction of the axis O2 of the nonlinear characteristic adjusting mechanisms (three-dimensional cam mechanisms) 45 and 46 and the displacement-load relationship of the disc spring 50 can be obtained by numerical calculation and represented as a chart or the like. Yes, the cam groove position can be easily set by obtaining the cam groove position that realizes FIG. 3 from the chart.

  More specifically, an example of a cam groove shape determination method for realizing the characteristics of the strong nonlinear stiffness element 3 will be described.

  When Type 1, Type 2, and Type 3 in FIG. 3 are shown in an overlapping manner, the relationship between the generated load and the axial displacement is as shown in FIG. Further, the axial displacement U4 in FIG. 20 is a use limit displacement of the strong nonlinear stiffness element 3.

  In this embodiment, since the follower 48 of the three-dimensional cam mechanism has four claws, one claw is controlled by the cam groove 57 in the range of 360 degrees / 4 = 90 degrees. Further, the pitch of the thread grooves 41 and 42 of the first rotating body 38 and the second rotating body 39 involved in the operation of the driven body 48 is moderate, and when the usage limit range U4, these rotating bodies 38 and 39 are ± 45. The pitch of the screw shaft 37 is determined so that rotation of the degree occurs (see FIG. 21).

FIG. 22 shows a setting example of the cam groove in the case of type 1.
FIG. 22 shows only 0 to 45 degrees, which is the upper half of the ± 45 degree portion. Here, 0 to 45 degrees is a case where the screw shaft 37 moves in the direction in which the disc spring 50 compresses. In addition, the first rotating body 38 and the second rotating body 39 need to reverse the shape of the cam groove 57 (replace left and right). In addition, since it is a direction in which the disc spring 50 is pulled from 0 degrees to -45 degrees, it is not necessary to move the follower 48, so the cam groove 57 may be provided at the same position as the initial position.

  Up to the displacement Og in the direction of the axis O2, the load becomes zero due to the gap Ug of the disc spring portion shown in FIG. Up to the axial O2 direction displacement U1 (point A in FIG. 20), it is the same as the curable restoring force characteristic by the disc spring 50, so the driven body 48 does not need to move from the initial state. After that, the effective displacement by the follower 48 may be reduced by the displacement difference between the black line and the broken line in FIG. For example, at point B, the displacement effective by the follower 48 is reduced by db. Similarly, at point C, the effective displacement by the follower 48 is reduced by dc. Similarly, the corresponding cam groove position is determined.

  FIG. 23 shows a setting example of the cam groove 57 in the case of type 2. Up to point A is the same as type 1. Thereafter, in order to maintain the load fp, the displacement should be reduced by the driven body 48 by the horizontal distance from the vertical line in FIG. Therefore, after the point B, a straight cam groove shape passing through the point E is obtained.

  FIG. 24 shows a setting example of the cam groove 57 in the case of type 3. Although up to point B is the same as type 2, after that point, it is only necessary to relax the displacement by the difference between the solid line and the broken line, so that a straight line is formed after point B. The difference between Type 2 and Type 3 is the slope of the straight line after point B.

  Although one embodiment of the vibration damping structure and the rotary inertia mass device according to the present invention has been described above, the present invention is not limited to the above embodiment, and can be appropriately changed without departing from the spirit thereof. It is.

In the present embodiment, the first example of FIGS. 7 to 10 and the second example of FIG. 16 are shown as the rotary inertia mass device 4, but for example, a variable inertia mass flywheel disclosed in JP 2013-213579 A You may apply what changes the coupling | bonding state of two flywheels arbitrarily by changing the viscosity of a magnetic fluid with the change of an electric current, and makes rotational inertial mass variable.
In this case, the current is controlled according to the value of the displacement sensor, and the value of the rotational inertial mass depending on the displacement can be set arbitrarily, so that the restoring force characteristic of the rotational inertial element required by the present invention can be realized. is there.

Further, as the rotary inertia mass device 4, a variable lead rolling element screw device disclosed in Japanese Patent No. 4366215 and an inertia mass damper disclosed in Japanese Patent Application Laid-Open No. 2010-19347 may be applied in combination.
The variable lead rolling element screw device can smoothly change the conversion magnification into the rotational motion according to the displacement by changing the lead length when converting the axial displacement of the screw shaft into the rotational motion. .

1 Damping structure 2 Displacement-dependent adjustment mechanism 3 Strong nonlinear stiffness element (nonlinear variable stiffness device)
4 Nonlinear rotational inertial elements (nonlinear rotational inertial mass device / variable inertial mass device)
5 Rotational inertial mass mechanism 6 Nonlinear characteristic adjustment mechanism (rotational inertial mass variable mechanism)
7 Screw shaft 8 First rotary weight 9 Connecting member 10 First screw groove 11 Second screw groove 12 Casing 13 Bearing 14 Second rotary weight 15 Linear motion conversion mechanism (three-dimensional cam mechanism)
16 Variable Rotation Inertia Mass Mechanism 20 Cylindrical Cam 21 Follower 22 Linear Slider 23 Rack Pinion 24 Rack 25 Rotating Shaft 26 Pinion 27 Variable Inertia Mass 28 Cam Groove 30 Variable Rotation Inertia Mass Mechanism 31 Pressing Mechanism 31a Spring 31b Ball 32 Additional Rotation Inertia Mass body 33 Friction member 35 Stiffness adjusting mechanism 36 Curing type rigid mechanism 37 Screw shaft 38 First rotating body 39 Second rotating body 40 Connecting member 41 First screw groove 42 Second screw groove 43 Casing 44 Bearing 45 One nonlinear characteristic adjustment Mechanism (three-dimensional cam mechanism)
46 The other nonlinear characteristic adjustment mechanism (three-dimensional cam mechanism)
47 Cylindrical cam 48 Follower 50 Disc spring 51 Pressing jig 52 First group of disc spring bodies 53 Second group of disc spring bodies 55 Deflection limiting jig (displacement limiting jig)
O1 axis O2 axis

Claims (3)

  A rotary inertia mass device comprising a three-dimensional cam mechanism for converting linear motion into rotary motion, wherein the relationship between displacement and rotary inertia mass effect is nonlinear by the three-dimensional cam mechanism.   A vibration damping structure comprising the rotary inertia mass device according to claim 1, wherein the natural period changes according to a displacement. In the vibration damping structure according to claim 2,
A vibration damping structure comprising: the rotary inertia mass device according to claim 1; and a variable stiffness device whose stiffness changes according to displacement.

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