JP6918415B2 - Vibration suppression device for structures - Google Patents

Description

The present invention relates to a structure vibration suppressing device for suppressing the vibration of a structure, and more particularly to a structure vibration suppressing device configured as an additional vibration system.

Conventionally, as a vibration suppression device for this type of structure, for example, one disclosed in Patent Document 1 is known. This vibration suppression device is applied to the piloti type building, and is between the additional rigid frame frame erected on the floor of the piloti floor of the building and the beam of the building and the additional rigid frame frame due to the vibration of the building. It is equipped with a rotary inertial mass damper and an additional damping element that operate as a relative displacement occurs in the horizontal direction. In the vibration suppression device, the ramen frame and the rotational inertia mass damper constitute an additional vibration system, and the horizontal rigidity of the ramen frame and the inertial mass of the rotational inertia mass damper are determined by the natural frequencies of the additional vibration system. It is adjusted to be tuned to (almost equal to) the first natural frequency of.

Patent No. 5234432

The characteristics of vibrations input to a building, such as seismic motion, are not constant, and their frequencies differ from time to time, and the way the building sways (responsiveness) changes accordingly. On the other hand, in the conventional vibration suppression device described above, the magnitude of the inertial mass of the rotary inertial mass damper cannot be controlled (cannot be changed), and the natural frequency of the additional vibration system is always set to the natural frequency of the building. It is only tuned to the first natural frequency. Therefore, when the frequency of the predominant frequency component of the vibration input to the building is close to the natural frequency of the second or third order other than the first order of the building, the vibration of the building is appropriately absorbed by the additional vibration system. It may not be possible to suppress it.

The present invention has been made to solve the above problems, and appropriately adjusts the inertial mass of the variable rotational inertial mass damper that constitutes the additional vibration system according to the occasional vibration input to the structure. It is an object of the present invention to provide a vibration suppression device for a structure that can be controlled and thereby appropriately suppress the vibration of the structure with an additional vibration system.

In order to achieve the above object, the invention according to claim 1 is a structure vibration suppression device for suppressing vibration of a structure erected on a foundation, and an elastic member connected to the structure. And, an additional vibration system is formed together with the elastic member, and the displacement of the structure transmitted through the elastic member due to the vibration of the structure is converted into the rotational motion of the rotating mass, and as the rotating mass rotates. A variable rotational inertial mass damper configured so that the inertial mass generated can be changed, a dominant frequency acquisition means for acquiring the dominant frequency, which is the frequency of the dominant frequency component of the vibration input to the structure, and the structure. The natural frequency of the additional vibration system is tuned to the nearby natural frequency, which is the natural frequency closest to the acquired dominant frequency among the predetermined multiple natural frequencies of the structure. It is characterized by comprising a first control means for controlling the inertial mass of the variable rotation inertial mass damper.

According to this configuration, the additional vibration system is composed of the variable rotation inertial mass damper and the elastic member, and the natural frequency of such an additional vibration system is the inertial mass of the variable rotation inertial mass damper and the rigidity of the elastic member. It is decided according to. Displacement due to vibration of the structure is transmitted to the variable rotational inertia mass damper via the elastic member, and the transmitted displacement of the structure is converted into the rotational motion of the rotating mass, which results in the rotation of the rotating mass. The additional vibration system including the variable rotational inertia mass damper and the elastic member vibrates. In the variable rotary inertial mass damper, the inertial mass generated by the rotation of the rotating mass can be changed.

Further, the predominant frequency, which is the frequency of the predominant frequency component of the vibration input to the structure, is acquired by the predominant frequency acquisition means, and during the vibration of the structure, among a plurality of predetermined natural frequencies of the structure. The inertial mass of the variable rotation inertial mass damper is controlled by the first control means so that the natural frequency of the additional vibration system is tuned to the near natural frequency, which is the natural frequency closest to the acquired dominant frequency. NS. As a result, by synchronizing the natural frequency of the added vibration system with the natural frequency in the vicinity of the structure, the vibration of the structure at the strongest dominant frequency included in the occasional vibration input to the structure is added. It can be appropriately absorbed and suppressed by the vibration system. As described above, according to the present invention, the inertial mass of the variable rotary inertial mass damper constituting the additional vibration system can be appropriately controlled according to the occasional vibration input to the structure, whereby the vibration of the structure can be appropriately controlled. Can be appropriately suppressed by the additional vibration system.

In the "tuning" in the present invention, in addition to adjusting the natural frequency of the additional vibration system so as to be the same as the near natural frequency, it is also adjusted so as to be substantially the same as the near natural frequency. Is included.

The invention according to claim 2 is the vibration suppression device for a structure according to claim 1, wherein the variable rotation inertial mass damper is a cylinder in which displacement due to vibration of the structure is transmitted via an elastic member and a cylinder. Is divided into a first fluid chamber and a second fluid chamber filled with a working fluid, and the displacement due to the vibration of the structure is transmitted via the elastic member so that the inside of the cylinder slides in the axial direction. The first continuous passage, which bypasses the piston and communicates with the first and second fluid chambers, and is filled with the working fluid, and the first communication passage, which is provided in the first communication passage, is provided in the first communication passage. A flow conversion mechanism that converts the flow of the working fluid into the rotational motion of the rotating mass is provided in parallel with the first communication passage, bypasses the piston, communicates with the first and second fluid chambers, and is filled with the working fluid. The first control means is a structure having a second passage and a flow adjustment mechanism for adjusting the flow rate of the working fluid flowing in the second passage, which is provided in the second passage. By adjusting the flow rate of the working fluid flowing through the second communication passage through the flow rate adjustment mechanism during vibration, the variable rotational inertia is adjusted so that the natural frequency of the additional vibration system is synchronized with the natural frequency in the vicinity of the structure. It is characterized by controlling the inertial mass of the mass damper.

According to this configuration, the displacement due to the vibration of the structure is transmitted to the cylinder and the piston via the elastic member, so that the piston slides in the cylinder and moves to one side of the first and second fluid chambers. When it moves, the working fluid in one of the fluid chambers is pushed out by the piston into the first communication passage, so that the working fluid flows in the first communication passage toward the other fluid chamber side. The flow of this working fluid is converted into the rotational motion of the rotating mass by the flow conversion mechanism.

Further, a second passage is provided in parallel with the first passage, and the flow rate of the working fluid flowing through the second passage is adjusted by the flow rate adjusting mechanism. Since this second passage bypasses the piston and is provided in parallel with the first passage, when the flow rate of the working fluid in the second passage changes, the first movement of the piston accompanies the change. The flow rate of the working fluid in the communication passage changes. Therefore, by adjusting the flow rate of the working fluid in the second passage through the flow rate adjusting mechanism, the flow rate of the working fluid in the first passage can be changed, and the amount of rotation of the rotating mass can be changed. Thereby, the inertial mass of the variable rotary inertial mass damper generated with the rotation of the rotary mass can be changed.

Further, according to the above-described configuration, the first control means adjusts the flow rate of the working fluid flowing through the second communication passage through the flow rate adjusting mechanism during the vibration of the structure, thereby adjusting the natural vibration of the additional vibration system. The inertial mass of the variable rotation inertial mass damper is controlled so that the number is synchronized with the natural frequency in the vicinity of the structure. As described above, the effect of the invention according to claim 1, that is, the inertial mass of the variable rotation inertial mass damper can be appropriately controlled according to the occasional vibration input to the structure, whereby the vibration of the structure is added to the vibration system. It is possible to effectively obtain the effect that it can be appropriately suppressed.

The invention according to claim 3 is the vibration suppression device for the structure according to claim 2, wherein the flow rate adjusting mechanism includes an electric pump for delivering the working fluid in the second continuous passage.

According to this configuration, by driving the electric pump, the flow of the working fluid can be forcibly generated in the second connecting passage. Further, by controlling the electric pump, the flow rate of the working fluid in the second passage is directly adjusted, and the flow rate of the working fluid in the first passage changes accordingly. Therefore, the variable rotation inertial mass damper The inertial mass of can be changed appropriately.

According to the fourth aspect of the present invention, in the vibration suppression device for the structure according to the second or third aspect, the piston is opened when the pressure of the working fluid in the first fluid chamber reaches the first predetermined pressure. , The first relief valve that communicates the first and second fluid chambers with each other and the valve opens when the pressure of the working fluid in the second fluid chamber reaches the second predetermined pressure, and the second and first fluid chambers are opened to each other. It is characterized in that a second relief valve for communication is provided.

According to this configuration, the piston is provided with the first and second relief valves, and the first relief valve opens when the pressure of the working fluid in the first fluid chamber reaches the first predetermined pressure. The second relief valve opens when the pressure of the working fluid in the second fluid chamber reaches the second predetermined pressure, and the opening of the first or second relief valve causes the first and second fluid chambers to open each other. Can be communicated. As a result, it is possible to prevent the pressure of the working fluid in the first and second fluid chambers from becoming excessive, and appropriately limit the axial force acting on the cylinder and the piston.

The invention according to claim 5 is the vibration suppression device for a structure according to claim 1, wherein the variable rotation inertial mass damper is a cylinder in which displacement due to vibration of the structure is transmitted via an elastic member and a cylinder. Is divided into a first fluid chamber and a second fluid chamber filled with a working fluid, and the displacement due to the vibration of the structure is transmitted via the elastic member so as to slide in the cylinder in the axial direction. The pressure energy of the first communication passage filled with the working fluid and the pressure energy of the working fluid flowing in the first communication passage, while communicating with the first and second fluid chambers by bypassing the piston and the piston. A variable capacitance type fluid pressure motor configured to be able to change the displacement volume while converting to rotational energy, an actuator for changing the displacement volume of the fluid pressure motor, and the rotational energy converted by the fluid pressure motor are transmitted. The first control means has a rotating mass that rotates due to the displacement, and the first control means changes the displacement volume of the fluid pressure motor via the actuator during the vibration of the structure, so that the natural frequency of the additional vibration system can be changed. It is characterized in that the inertial mass of the variable rotational inertial mass damper is controlled so as to be tuned to the near natural frequency of the structure.

According to this configuration, the displacement due to the vibration of the structure is transmitted to the cylinder and the piston via the elastic member, so that the piston slides in the cylinder and moves to one side of the first and second fluid chambers. When it moves, the working fluid in one of the fluid chambers is pushed out by the piston into the first communication passage, so that the working fluid flows in the first communication passage toward the other fluid chamber side. The pressure energy of the working fluid is converted into rotational energy by a variable displacement fluid pressure motor, and the converted rotational energy is transmitted to the rotary mass to rotate the rotary mass. Along with this, an inertial mass corresponding to the rotational inertial mass of the rotating mass is generated. In this case, the amount of rotation of the rotating mass with respect to the pressure energy of the same working fluid is changed by changing the push-out volume (geometric volume push-out per rotation of the fluid pressure motor) of the variable capacitance type fluid pressure motor. This makes it possible to change the inertial mass of the variable rotational inertial mass damper generated with the rotation of the rotating mass.

Further, according to the above-described configuration, in the first control means, the natural frequency of the additional vibration system is unique to the vicinity of the structure by changing the push-out volume of the fluid pressure motor via the actuator during the vibration of the structure. The inertial mass of the variable rotation inertial mass damper is controlled so as to be synchronized with the frequency. As described above, the effect of the invention according to claim 1, that is, the inertial mass of the variable rotation inertial mass damper can be appropriately controlled according to the occasional vibration input to the structure, whereby the vibration of the structure is added to the vibration system. It is possible to effectively obtain the effect that it can be appropriately suppressed.

The invention according to claim 6 is the vibration suppression device for a structure according to claim 1, wherein the variable rotation inertial mass damper rotates the displacement of the structure transmitted through the elastic member with the vibration of the structure. A conversion mechanism that converts to power and a stepped speed change mechanism that transmits the rotational power converted by the conversion mechanism to the rotating mass in a state where the speed is changed at one speed ratio selected from a plurality of predetermined gear ratios. The plurality of gear ratios are a plurality of natural frequencies of the additional vibration system obtained corresponding to each of the plurality of gear ratios when each of the plurality of gear ratios is selected during the vibration of the structure. However, it is set so as to synchronize with each of a plurality of predetermined natural frequencies of the structure, and the first control means corresponds to the near natural frequencies of the structure from the plurality of gear ratios of the speed change mechanism. It is characterized in that the inertial mass of the variable rotation inertial mass damper is controlled so that the natural frequency of the additional vibration system is synchronized with the near natural frequency by selecting the gear ratio to be used.

According to this configuration, the displacement of the structure transmitted through the elastic member due to the vibration of the structure is converted into a rotational motion by the conversion mechanism. The converted rotational power is transmitted to the rotary mass by a stepped speed change mechanism in a state where the gear is changed at one gear ratio selected from a plurality of predetermined gear ratios, whereby the rotary mass rotates. In this way, the rotational power from the conversion mechanism is transmitted to the rotating mass in a state where the speed is changed by one of these plurality of gear ratios, and the rotational amount of the rotating mass is increased by selecting the gear ratio of the shifting mechanism. By changing it, the inertial mass of the variable rotary inertial mass damper generated with the rotation of the rotating mass can be changed. Further, as described above, the natural frequency of the additional vibration system including the variable rotational inertial mass damper is determined according to the inertial mass of the variable rotational inertial mass damper.

In this case, the plurality of gear ratios are the plurality of natural frequencies of the additional vibration system obtained corresponding to each of the plurality of gear ratios when each of the plurality of gear ratios is selected during the vibration of the structure. Is set to tune corresponding to each of a plurality of predetermined natural frequencies of the structure. Further, the first control means selects a gear ratio corresponding to the near natural frequency of the structure from a plurality of gear ratios so that the natural frequency of the additional vibration system is synchronized with the near natural frequency. Variable rotation inertial mass The inertial mass of the damper is controlled. As described above, the effect of the invention according to claim 1, that is, the inertial mass of the variable rotation inertial mass damper can be appropriately controlled according to the occasional vibration input to the structure, whereby the vibration of the structure is added to the vibration system. It is possible to effectively obtain the effect that it can be appropriately suppressed.

According to the invention of claim 7, in the structure vibration suppression device according to claim 1, the variable rotation inertial mass damper rotates the displacement of the structure transmitted through the elastic member with the vibration of the structure. It has a conversion mechanism that converts it into power, and a stepless speed change mechanism that transmits the rotational power converted by the conversion mechanism to the rotating mass in a speed-shifted state and allows the gear ratio to be changed steplessly. However, the first control means has a variable rotational inertia mass so that the natural frequency of the additional vibration system is synchronized with the natural frequency in the vicinity of the structure by setting the gear ratio of the speed change mechanism during the vibration of the structure. It is characterized by controlling the inertial mass of the damper.

According to this configuration, the displacement of the structure transmitted through the elastic member due to the vibration of the structure is converted into a rotational motion by the conversion mechanism. The converted rotational power is transmitted to the rotating mass in a state of being shifted by the stepless speed change mechanism, whereby the rotating mass rotates. The gear ratio of the transmission mechanism can be changed steplessly. In this way, the rotational power from the conversion mechanism is transmitted to the rotating mass in a state where the speed is changed steplessly, and the rotational mass is changed by changing the rotational amount of the rotating mass by setting the gear ratio of the transmission mechanism. The inertial mass of the variable rotary inertial mass damper generated by the rotation of the damper can be continuously changed.

Further, during the vibration of the structure, the first control means sets the gear ratio of the speed change mechanism so that the natural frequency of the additional vibration system is synchronized with the natural frequency in the vicinity of the structure. The inertial mass of the mass damper is controlled. As described above, the effect of the invention according to claim 1, that is, the inertial mass of the variable rotation inertial mass damper can be appropriately controlled according to the occasional vibration input to the structure, whereby the vibration of the structure is added to the vibration system. It is possible to effectively obtain the effect that it can be appropriately suppressed.

The invention according to claim 8 is the vibration suppression device for the structure according to claim 6 or 7, wherein the speed change mechanism has a rotation shaft in which the rotational power converted by the conversion mechanism is transmitted in a state of being changed. A fitting hole is formed in the rotating mass, and a friction material to be fitted between the rotating shaft and the inner peripheral surface of the fitting hole is further provided. The rotating mass is connected to the rotating shaft via the friction material. It is characterized by being.

According to this configuration, the rotational power converted by the conversion mechanism is transmitted to the rotary shaft of the transmission mechanism in a speed-shifted state, and the rotary mass transmits a friction material that fits between the fitting hole and the rotary shaft. It is connected to the rotating shaft via. Therefore, when the rotational torque of the rotating mass becomes very large due to the extremely large vibration of the structure, the friction coefficient of the friction material is set so that the rotating mass and / or the rotating shaft slides with respect to the friction material. By setting, it is possible to limit the inertial force due to the inertial mass generated by the rotation of the rotating mass.

The invention according to claim 9 comprises a variable damping damper configured to dampen the vibration of the additional vibration system and change the damping coefficient in the vibration suppressing device for the structure according to any one of claims 1 to 8. During vibration of the structure, the maximum value of the response magnification of the structure is minimized according to the near natural frequency and the inertial mass of the variable rotation inertial mass damper controlled by the first control means according to the fixed point theory. As described above, a second control means for controlling the damping coefficient of the variable damping damper is further provided.

According to this configuration, the vibration of the additional vibration system is damped by a variable damping damper whose damping coefficient can be changed. Further, during the vibration of the structure, the response magnification of the structure is increased according to the near natural frequency according to the fixed point theory by the second control means and the inertial mass of the variable rotation inertial mass damper controlled by the first control means. The damping coefficient of the variable damping damper is controlled so that the maximum value of is minimized.

In recent years, in the field of vibration damping of structures using an additional vibration system (tuned mass damper), by setting the damping factor of the damping element that damps the vibration of the additional vibration system to the optimum damping coefficient according to the fixed point theory by Den Hartog. , It is known that the vibration of the structure can be suppressed more appropriately. In this case, as disclosed in, for example, "Displacement control design of buildings, Norio Inoue / Yuki Igoko, Maruzen Publishing", the mass ratio, that is, the additional vibration system, is used as a parameter for setting the damping coefficient to the optimum damping coefficient. The ratio of the inertial mass to the mass of the structure and the natural frequency of the structure corresponding to the vibration of the structure to be suppressed (hereinafter referred to as "object natural frequency") are mainly used. Here, while the mass of the structure is a predetermined fixed value, as described above in the present invention, the inertial mass of the variable rotation inertial mass damper of the additional vibration system and the target natural frequency of the structure are different from time to time. It changes according to the vibration.

On the other hand, according to the present invention, the near natural frequency, that is, the natural frequency of the structure closest to the dominant frequency of the vibration input to the structure, and the inertial mass of the controlled variable rotational inertia mass damper. Since the damping coefficient is controlled as described above, the vibration of the structure can be suppressed more appropriately.

The invention according to claim 10 is the vibration suppression device for the structure according to claim 9, which is subordinate to any one of claims 2 to 5, wherein the variable damping damper includes a cylinder, a piston, a first communication passage, and a working fluid. The second control means is provided in the first communication passage and has an adjusting valve for adjusting the flow resistance of the working fluid flowing in the first communication passage by changing the opening degree, and the second control means is in the vicinity during the vibration of the structure. By changing the opening of the adjusting valve according to the natural frequency and the inertial mass of the variable rotation inertial mass damper controlled by the first control means, the maximum value of the response magnification of the structure is minimized. It is characterized by controlling the damping coefficient of the variable damping damper.

According to this configuration, the flow resistance of the working fluid in the first communication passage is adjusted by the adjusting valve. As described in the description of the invention according to claims 2 and 5, the displacement due to the vibration of the structure is transmitted to the cylinder and the piston via the elastic member, so that the piston slides in the cylinder, and accordingly. , The working fluid pressed by the piston flows between the first communication passage, the first and second fluid chambers, and a pressure difference of the working fluid is generated between the first and second fluid chambers, and this pressure is generated. The difference acts as a resistance to the piston. Further, the flow (pressure energy) of the working fluid flowing in the first continuous passage is converted into the rotational motion of the rotating mass of the additional vibration system. As is clear from the above operation, the variable damping damper having a cylinder, a piston, a first communication passage, a working fluid, and a regulating valve functions to dampen the displacement of the elastic member and dampen the vibration of the additional vibration system. However, the damping coefficient can be changed by adjusting the flow resistance of the working fluid in the first communication passage with the adjusting valve.

Further, according to the above-described configuration, the opening degree of the adjusting valve is changed according to the natural frequency in the vicinity and the inertial mass of the controlled variable rotation inertial mass damper during the vibration of the structure. The damping coefficient of the variable damping damper is controlled so that the maximum value of the response magnification is minimized. Therefore, the effect of the invention according to claim 9, that is, the effect of suppressing the vibration of the structure more appropriately can be effectively obtained. Further, since the cylinder, the piston, the first communication passage, and the working fluid are also used as the components of the variable rotation inertial mass damper and the variable damping damper, the vibration suppression device can be miniaturized accordingly.

The invention according to claim 11 is the vibration suppressing device for a structure according to claim 9, which is subordinate to any one of claims 6 to 8. In the variable damping damper, the displacement due to the vibration of the structure is transmitted through an elastic member. The cylinder and the inside of the cylinder are divided into two viscous fluid chambers filled with viscous fluid, and the displacement due to the vibration of the structure is transmitted through the elastic member, so that the inside of the cylinder is axially oriented. A piston configured to slide on the surface and a communication passage that bypasses the piston and communicates with two viscous fluid chambers, and is provided in a communication passage filled with a viscous fluid and a communication passage, and the communication passage is provided by changing the opening degree. It has a regulating valve that adjusts the flow resistance of the viscous fluid flowing inside, and the second control means is the proximity natural frequency during the vibration of the structure and the variable rotational inertia mass controlled by the first control means. By changing the opening degree of the adjusting valve according to the inertial mass of the damper, the damping coefficient of the variable damping damper is controlled so that the maximum value of the response magnification of the structure is minimized.

According to this configuration, the displacement due to the vibration of the structure is transmitted to the piston via the elastic member, so that when the piston moves in the cylinder to one side of the two viscous fluid chambers, the viscous fluid of the one is transferred. The viscous fluid in the chamber is pushed out into the communication passage by the piston, flows through the communication passage, and the viscous fluid in the communication passage flows into the other viscous fluid chamber. Along with this, a pressure difference of the viscous fluid is generated between the two viscous fluid chambers, and this pressure difference acts as a resistance force on the piston. Further, the flow resistance of the viscous fluid flowing in the communication passage is adjusted by the adjusting valve. As is clear from the above operation, the variable damping damper having a cylinder, a piston, a communication path, a viscous fluid, and a regulating valve functions to dampen the displacement of the elastic member and dampen the vibration of the additional vibration system. The damping coefficient can be changed by adjusting the flow resistance of the viscous fluid flowing in the communication passage by changing the opening degree of the adjusting valve.

Further, according to the above-described configuration, the opening degree of the adjusting valve is changed according to the natural frequency in the vicinity and the inertial mass of the controlled variable rotation inertial mass damper during the vibration of the structure. The damping coefficient of the variable damping damper is controlled so that the maximum value of the response magnification is minimized. Therefore, the effect of the invention according to claim 9, that is, the effect of suppressing the vibration of the structure more appropriately can be effectively obtained.

It is sectional drawing which shows the variable rotation inertial mass damper of the vibration suppression apparatus by 1st Embodiment of this invention. It is sectional drawing which follows the line II-II of FIG. It is sectional drawing which follows the line III-III of FIG. It is a block diagram which shows the control device of the vibration suppression device by 1st Embodiment. It is a figure which shows roughly the vibration suppression device by 1st Embodiment together with a part of a building to which this is applied. It is sectional drawing which shows the operation state in the 1st operation mode of the variable rotation inertial mass damper. It is sectional drawing which shows the operation state in the 2nd operation mode of the variable rotation inertial mass damper. It is a flowchart which shows a part of the tuning control processing executed by the control device of FIG. It is a flowchart which shows the continuation of FIG. (A) The relationship between the frequency of the seismic motion input to the building to which the vibration suppression device of the present invention is applied and the acceleration response magnification of the building to the input, and (b) the relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the seismic motion. , (C) The relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the building response is shown in the case where the predominant frequency of the seismic motion is equal to the primary natural frequency of the building, respectively, and (d) the frequency of the seismic motion and The relationship between the acceleration response magnification of the building to the input, (e) the relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the seismic motion, (f) the relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the building response, the predominance of the seismic motion. It is a figure which shows the case where the frequency is equal to the secondary natural frequency of a building, respectively. (A) The relationship between the frequency of the seismic motion input to the building to which the conventional vibration suppression device is applied and the acceleration response magnification of the building to the input, (b) The relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the seismic motion. (C) The relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the building response is shown for each case where the predominant frequency of the seismic motion is equal to the secondary natural frequency of the building, and (d) the secondary mode of the building. The relationship between the frequency of the seismic motion input to the building to which the vibration suppression device (additional vibration system) configured to synchronize with the natural frequency of the building is applied and the acceleration response magnification of the building to the input is (e) of the seismic motion. The relationship between the frequency and the Fourier amplitude spectrum of the seismic motion is shown, and (f) the relationship between the frequency of the seismic motion and the Fourier amplitude spectrum of the building response is shown for the case where the predominant frequency of the seismic motion is equal to the first-order natural frequency of the building. It is a figure. It is sectional drawing which shows the variable rotation inertial mass damper of the vibration suppression apparatus by 2nd Embodiment of this invention. It is sectional drawing which shows the variable damping damper of the vibration suppression apparatus by 2nd Embodiment. It is a block diagram which shows the control device of the vibration suppression device by 2nd Embodiment. It is a figure which shows roughly the vibration suppression device by 2nd Embodiment together with a part of a building to which this is applied. It is a flowchart which shows the tuning control processing executed by the control device of FIG. It is sectional drawing which shows the variable rotation inertial mass damper of the vibration suppression apparatus by the 3rd Embodiment of this invention. It is sectional drawing which shows the variable rotation inertial mass damper of the vibration suppression apparatus according to 4th Embodiment of this invention. It is a block diagram which shows the control device of the vibration suppression device according to 4th Embodiment. It is a figure which shows roughly the vibration suppression device by 4th Embodiment together with a part of a building to which this is applied. FIG. 5 is a flowchart showing a tuning control process executed by the control device of FIG. It is sectional drawing which shows the variable rotation inertial mass damper of the vibration suppression apparatus according to 5th Embodiment of this invention. It is a block diagram which shows the control apparatus of the vibration suppression apparatus according to 5th Embodiment. It is a figure which shows roughly the vibration suppression device by 5th Embodiment together with a part of a building to which this is applied. FIG. 5 is a flowchart showing a tuning control process executed by the control device of FIG. 24.

Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. The vibration suppression device according to the first embodiment of the present invention includes the variable rotation inertial mass damper 1 shown in FIG. 1, and is applied to a high-rise building B (see FIG. 5). This building B has a rigid frame structure in which a plurality of columns and beams are combined in a grid shape, and is erected on a foundation (not shown). The variable rotary inertial mass damper 1 is configured so that the inertial mass and the damping coefficient generated as the rotary mass 21, which will be described later, rotates can be continuously changed. It includes a piston 3 slidably provided in the cylinder 2 in the axial direction, a piston rod 4 integrated with the piston 3, and a first-series passage 5 and a second-series passage 6 connected to the cylinder 2.

The cylinder 2 integrally has a cylindrical peripheral wall 2a and a disk-shaped first end wall 2b and a second end wall 2c provided at both ends of the peripheral wall 2a in the axial direction, respectively. The space defined by the peripheral walls 2a, the first and second end walls 2b, and 2c is divided into the first fluid chamber 2d and the second fluid chamber 2e by the piston 3. The first and second fluid chambers 2d and 2e are filled with the viscous fluid HF. The viscous fluid HF is composed of a suitable viscous fluid, such as silicone oil.

Further, the first end wall 2b of the cylinder 2 is integrally provided with convex portions 2f protruding outward in the axial direction concentrically, and the convex portions 2f are provided with a first convex portion 2f via a universal joint. A fitting FL1 is provided. Further, a rod guide hole 2g is formed at the center of the second end wall 2c. One end of the piston rod 4 is integrally connected to the piston 3, extends in the axial direction in the cylinder 2, and is liquidtightly inserted into the rod guide hole 2g via a seal, and is outside the second end wall 2c. It extends in the direction. A second attachment FL2 is provided at the other end of the piston rod 4 via a universal joint.

Further, the outer peripheral surface of the piston 3 is in liquid-tight contact with the inner peripheral surface of the peripheral wall 2a of the cylinder 2 via a seal, and a plurality of axially penetrating outer end portions of the piston 3 in the radial direction. The first communication hole 3a and the second communication hole 3b (only one of each is shown) are formed. A first relief valve 11 is provided in the first communication hole 3a, and a second relief valve 12 is provided in the second communication hole 3b.

The first relief valve 11 is composed of a valve body and a spring that urges the valve body in the valve closing direction, and when the pressure of the viscous fluid HF in the first fluid chamber 2d is smaller than a predetermined upper limit value, The first communication hole 3a is closed, and when the upper limit is reached, the first communication hole 3a is opened. As a result, the first and second fluid chambers 2d and 2e communicate with each other through the first communication hole 3a, and the pressure in the first fluid chamber 2d is released to the second fluid chamber 2e side.

Similarly, the second relief valve 12 is composed of a valve body and a spring that urges the valve body in the valve closing direction, and the pressure of the viscous fluid HF in the second fluid chamber 2e is higher than the above upper limit value. When it is small, the second communication hole 3b is closed, and when the upper limit is reached, the second communication hole 3b is opened. As a result, the first and second fluid chambers 2d and 2e communicate with each other through the second communication hole 3b, and the pressure in the second fluid chamber 2e is released to the first fluid chamber 2d side. The upper limit values of the first and second relief valves 11 and 12 may be set to different values.

The first and second communication passages 5 and 6, respectively, are connected to the cylinder 2 so as to bypass the piston 3 and communicate with the first and second fluid chambers 2d and 2e, and are provided in parallel with each other. There is. Further, the cross-sectional areas of both 5 and 6 are set to a value smaller than the cross-sectional area of the cylinder 2, and the first and second fluid chambers 2d and 2e are set in the first and second passages 5 and 6. Similarly, the viscous fluid HF is filled. In FIG. 1, for convenience, the reference numerals of the viscous fluid HF in the first and second passages 5 and 6 are omitted.

Further, the variable rotary inertial mass damper 1 uses the adjusting valve 15 for adjusting the flow resistance of the viscous fluid HF flowing in the first continuous passage 5 and the flow of the viscous fluid HF in the first continuous passage 5 as a rotary motion. A gear motor M to be converted, a rotating mass 21 connected to the gear motor M, and a gear pump P for adjusting the flow rate of the viscous fluid HF flowing through the second communication passage 6 are further provided. The regulating valve 15 is composed of, for example, a normally open type solenoid valve, and its opening degree can be continuously changed. The gear motor M is of the circumscribed gear type, and has a casing 22 and a first gear 23 and a second gear 24 housed in the casing 22. An inscribed gear type may be used as the gear motor M.

The casing 22 is integrally provided in the central portion of the first communication passage 5, and communicates with the first communication passage 5 via two entrances 22a and 22b facing each other. Further, the first and second gears 23 and 24 are respectively composed of spur gears, are integrally provided on the first and second rotating shafts 25 and 26, and mesh with each other. The first and second rotating shafts 25 and 26 extend horizontally in the direction orthogonal to the first communication passage 5, and are rotatably supported by the casing 22, and the first rotating shaft 25 projects to the outside of the casing 22. (See Fig. 2). Further, the meshing portions of the first and second gears 23 and 24 face the inlets and outlets 22a and 22b of the casing 22. Further, the above-mentioned rotating mass 21 is coaxially and integrally provided on the portion of the first rotating shaft 25 protruding from the casing 22. The rotating mass 21 is made of a material having a relatively large specific gravity, for example, iron, and is formed in a disk shape.

The gear pump P is of the external gear type, and has a flow rate adjusting motor 31 as a power source, a casing 32, and a first gear 33 and a second gear 34 housed in the casing 32. An inscribed gear type gear pump P may be used. The casing 32 is integrally provided in the central portion of the second communication passage 6, and communicates with the second communication passage 6 via two entrances 32a and 32b facing each other. Further, the first and second gears 33 and 34 are respectively composed of spur gears, are integrally provided on the first and second rotating shafts 35 and 36, and mesh with each other. The first and second rotating shafts 35 and 36 each extend horizontally in a direction orthogonal to the second communication passage 6 and are rotatably supported by the casing 32, and the first rotating shaft 35 projects to the outside of the casing 32. (See Fig. 3). Further, the meshing portions of the first and second gears 33 and 34 face the inlets and outlets 32a and 32b of the casing 32.

The flow rate adjusting motor 31 is composed of, for example, a DC motor capable of forward / reverse rotation. The rotor (not shown) of the flow rate adjusting motor 31 is coaxially connected to the portion of the first rotating shaft 35 protruding from the casing 32, and integrally rotates the first rotating shaft 35 and the first gear 33.

Further, the vibration suppression device further includes a control device 51, a power supply 52, and a seismograph 53 shown in FIG. The control device 51 is composed of a combination of a CPU, RAM, ROM, I / O interface, DC / DC converter, etc., the power supply 52 is composed of, for example, a battery, and the regulating valve 15 and the flow rate adjusting motor 31 are It is connected to the power supply 52 via the control device 51. The operation of the adjusting valve 15 and the flow rate adjusting motor 31 is controlled by the control device 51. The seismograph 53 is composed of, for example, an acceleration sensor, and is provided on a foundation (not shown) on which the building B is erected. The seismograph 53 measures the seismic motion input to the building B, and the measurement signal is used as a control device 51. Enter in.

Further, the vibration suppression device includes a first elastic member EM1 and a second elastic member EM2 shown in FIG. The first and second elastic members EM1 and EM2 are made of elastic steel materials having relatively low rigidity, and their rigidity is set to a predetermined value. Further, the first and second elastic members EM1 and EM2 are fixed to the lower beam BD and the upper beam BU of the building B, respectively, and extend upward from the lower beam BD and downward from the upper beam BU, respectively. The first and second attachments FL1 and FL2 are attached to the first and second elastic members EM1 and EM2, respectively. As described above, the cylinder 2 and the piston rod 4 of the variable rotary inertial mass damper 1 are connected to the lower beam BD and the upper beam BU via the first and second elastic members EM1 and EM2, respectively, and both BD and BU. It extends horizontally between. Further, the variable rotation inertial mass damper 1, the first and second elastic members EM1 and EM2 constitute an additional vibration system.

Further, the variable rotary inertial mass damper 1, the first and second elastic members EM1 and EM2 are provided in one set for each of all the layers of the building B, and FIG. 5 shows one set thereof. In FIG. 5, for convenience, some components such as the first and second passages 5 and 6 are not shown. By the way, it goes without saying that the variable rotary inertial mass dampers 1, 1st and 2nd elastic members EM1 and EM2 may be provided not only in all the layers of the building B but only in some layers.

Next, the operation of the variable rotation inertial mass damper 1 will be described. When a relative displacement in the horizontal direction occurs between the upper and lower beams BU and BD as the building B vibrates, the relative displacement is caused by the cylinder 2 and the piston via the first and second elastic members EM1 and EM2. By being transmitted to the rod 4 as an external force, the cylinder 2 and the piston rod 4 move relatively in the axial direction, and the piston 3 slides in the cylinder 2.

In this case, when the piston 3 moves to the first fluid chamber 2d side (left side in FIG. 1), a part of the viscous fluid HF in the first fluid chamber 2d is pushed out into the first continuous passage 5 by the piston 3. As a result, the viscous fluid HF flows to the second fluid chamber 2e side (right side) in the first communication passage 5. On the contrary, when the piston 3 moves to the second fluid chamber 2e side (right side), a part of the viscous fluid HF in the second fluid chamber 2e is pushed out by the piston 3 into the first continuous passage 5. As a result, the viscous fluid HF flows to the first fluid chamber 2d side (left side) in the first communication passage 5.

The flow of the viscous fluid HF is converted into rotary motion by the gear motor M, the first and second gears 23 and 24 are rotated, and the first rotating shaft 25 and the rotating mass 21 integrated with the first gear 23 are rotated. As a result, the additional vibration system including the variable rotary inertial mass damper 1, the first and second elastic members EM1 and EM2 vibrates.

Further, as the rotating mass 21 rotates as described above, the inertial mass Md of the variable rotating inertial mass damper 1 is generated. This inertial mass Md is the inertial mass in the axial direction with respect to the external force due to the vibration input to the cylinder 2 and the piston 3, and is the inertial mass by the rotating mass 21 (inertial mass based on the rotating inertial mass of the rotating mass 21 (equivalent mass)). Mr and the inertial mass Mh due to the viscous fluid HF in the first communication passage 5 are included. As described below, the inertial mass Mr by the rotating mass 21 at this time controls the operation of the gear pump P via the flow rate adjusting motor 31 to adjust the flow rate of the viscous fluid HF in the second communication passage 6. By doing so, it changes continuously. FIG. 6 shows that when the piston 3 moves to the first fluid chamber 2d side, the flow rate adjusting motor 31 is rotationally driven in the clockwise direction in the figure, so that the viscous fluid HF in the second communication passage 6 is driven by the piston 3. It shows the situation (hereinafter referred to as "first operation mode") in which the fluid is flowing to the first fluid chamber 2d side, which is the same as the above. Note that in FIG. 6, for convenience, the reference numerals of some of the components are omitted.

In this first operation mode, the volume (flow rate) V1 of the viscous fluid HF that flows with the movement of the piston 3 and the volume (flow rate) of the viscous fluid HF that is sent out by the gear pump P in the second communication passage 6 and flows. ) The sum with V2 becomes the volume (flow rate) V of the viscous fluid HF flowing in the first continuous passage 5 and flowing into the gear motor M, so that the amount of rotation of the rotating mass 21 becomes large.

Further, the inertial mass Mr by the rotating mass 21 at this time is obtained as follows. First, from the above relationship, the following equation (1) holds.
V = V1 + V2 …… (1)
In the following, V1 is referred to as "piston volume", V2 is referred to as "pump volume", and V is referred to as "motor volume". The piston volume V1 is represented by the following equation (2), and the pump volume V2 is represented by the following equation (3).
V1 = Ap ・ X …… (2)
V2 = vmd ・ n …… (3)
Here, Ap is the pressure receiving area of the piston 3 with respect to the viscous fluid HF, and when the piston 3 is moving to the first fluid chamber 2d side, it becomes the cross-sectional area of the piston 3, and the piston 3 becomes the second fluid chamber 2e. When moving to the side, the value is obtained by subtracting the cross-sectional area of the piston rod 4 from the cross-sectional area of the piston 3. The pressure receiving area Ap of the piston 3 may be set to the cross-sectional area of the piston 3 or a value obtained by subtracting the cross-sectional area of the piston rod 4 from the cross-sectional area of the piston 3 regardless of the moving direction of the piston 3. .. Further, X is the movement amount of the piston 3, vmd is the delivery volume (pushing volume) per rotation of the flow rate adjusting motor 31, and n is the rotation speed of the flow rate adjusting motor 31.

From the above equations (1) and (2), the movement amount X of the piston 3 in the first operation mode is expressed by the following equation (4).
X = V1 / Ap = (V-V2) / Ap …… (4)

Further, assuming that the gear motor M makes one rotation, the motor volume V required for the one rotation is vm (pushing volume), and the pump volume V2 at this time is vm2. By substituting into V and V2 of), the movement amount Xm of the piston 3 at this time is expressed by the following equation (5).
Xm = (vm-vm2) / Ap …… (5)

The movement amount Xm of the piston 3 corresponds to the lead length (Ld) of the ball screw in the mass damper using the ball screw mechanism. Therefore, the inertial mass (equivalent mass) Mr by the rotating mass 21 in the first operation mode is expressed by the following equation (6).
^{Mr = (2π / Xm) 2} · md · D 2/8
= {(2π · Ap) / (vm-vm2)} 2 · md · D 2/8 ...... (6)
Here, md is the actual amount of the rotating mass 21, and D is the outer diameter of the rotating mass 21.

As is clear from this equation (6), the inertial mass Mr in the first operation mode increases by the amount corresponding to the pump volume V2 (= vm2). Therefore, the inertial mass Mr can be changed by changing the pump volume V2 by adjusting the rotation speed of the flow rate adjusting motor 31.

On the other hand, with respect to the above first operation mode, FIG. 7 shows a second operation by driving the flow rate adjusting motor 31 to rotate counterclockwise in the figure when the piston 3 moves to the first fluid chamber 2d side. It shows a situation in which the viscous fluid HF in the communication passage 6 is flowing to the second fluid chamber 2e side contrary to the piston 3 (hereinafter referred to as “second operation mode”). Note that in FIG. 7, for convenience, the reference numerals of some of the components are omitted. In this second operation mode, the difference between the piston volume V1 and the pump volume V2 becomes the motor volume V (the following equation (1)'), so that the amount of rotation of the rotating mass 21 becomes small.
V = V1-V2 …… (1)'
From this relationship, the movement amount X of the piston 3 in the second operation mode is expressed by the following equation (4)'by replacing (V-V2) in the equation (4) with (V + V2).
X = V1 / Ap = (V + V2) / Ap …… (4)'

Further, the movement amount Xm of the piston 3 when the gear motor M makes one rotation is expressed by the following equation (5)'by replacing (vm-vm2) in the equation (5) with (vm + vm2).
Xm = (vm + vm2) / Ap …… (5)'

Therefore, the inertial mass Mr by the rotating mass 21 in the second operation mode is expressed by the following equation (6)'.
^{Mr = (2π / Xm) 2} · md · D 2/8
= {(2π · Ap) / (vm + vm2)} 2 · md · D 2/8
…… (6)'

As is clear from this equation (6)', the inertial mass Mr by the rotating mass 21 in the second operation mode is reduced by the amount corresponding to the pump volume V2 (= vm2). Therefore, the inertial mass Mr can be changed by changing the pump volume V2 by adjusting the rotation speed of the flow rate adjusting motor 31.

The pump volume vm2 has positive and negative values, and the case where the flow direction of the viscous fluid HF in the second passage 6 is the same as the moving direction of the piston 3 (first operation mode) is set as a negative value, and the opposite case (second operation). When the mode) is set to a positive value, the above equations (6) and (6)'are integrated into the equation (6)'. Further, when this equation (6)'is expressed with respect to the pump volume vm2, the following equation (7) is obtained.
vm2 = {(2π ・ Ap) / sqrt [8Mr / (md ・ D ² )]} -vm
…… (7)

Further, the rotation speed n of the flow rate adjusting motor 31 required to flow the viscous fluid HF having the pump volume vm2 is represented by the following equation (8) from the above equation (3).
n = vm2 / vmd …… (8)

As described above, in the variable rotation inertial mass damper 1, the gear pump P is operated to forcibly generate the flow of the viscous fluid HF in the second communication passage 6, and the rotation rate n of the flow rate adjusting motor 31. The pump volume V2 (flow rate of the viscous fluid HF in the second communication passage 6) is adjusted by controlling the pump volume V2. Then, the motor volume V (flow rate of the viscous fluid HF in the first communication passage 5) changes with respect to the piston volume V1 (flow rate of the viscous fluid HF accompanying the movement of the piston 3) by the amount of the pump volume V2. By changing the amount of rotation of the rotating mass 21 accordingly, the inertial mass Mr by the rotating mass 21 is changed, and by extension, the inertial mass Md of the variable rotating inertial mass damper 1 including the inertial mass Mr. is changed.

Note that the inertial mass Mh by the viscous fluid HF of the first communication passage 5 is represented by Mh = ρ · Ae1 · l1 · α1 2, in a very small tendency compared to the inertial mass Mr by rotating mass 21 .. Here, ρ is the density of the viscous fluid HF, Ae1 and l1 are the cross-sectional area and length of the first passage 5, respectively, and α1 is the pressure receiving area of the piston 3 with respect to the cross-sectional area Ae1 of the first passage 5. It is a ratio of Ap.

Further, by switching the rotation direction of the flow rate adjusting motor 31, the flow direction of the viscous fluid HF in the second communication passage 6 is changed to the same direction as the moving direction of the piston 3 (first operation mode) or the opposite direction (second operation). By switching to (mode), the motor volume V can be increased or decreased with respect to the piston volume V1. As a result, the range of change of the inertial mass Mr by the rotating mass 21 is expanded.

Further, as is clear from the above-described operation of the variable rotation inertial mass damper 1, the cylinder 2, the piston 3, the first communication passage 5, the viscous fluid HF and the adjusting valve 15 are the variable rotation inertial mass damper 1, the first and the first. 2 It functions as a variable damping damper that damps the vibration of the additional vibration system including the elastic members EM1 and EM2 and can continuously change the damping coefficient.

Specifically, as the piston 3 slides in the cylinder 2 as described above, a pressure difference of the viscous fluid HF is generated between the first and second fluid chambers 2d and 2e, and this pressure difference is , Acts on the piston 3 as a resistance force (difficulty in flowing), and also acts to dampen the displacements of the first and second elastic members EM1 and EM2 and dampen the vibration of the additional vibration system. In this case, the damping coefficient is continuously changed by adjusting the flow resistance of the viscous fluid HF flowing in the first continuous passage 5 by changing the opening degree of the adjusting valve 15, and the damping coefficient is changed to The smaller the opening degree of the regulating valve 15, the larger the pressure difference of the viscous fluid HF between the first and second fluid chambers 2d and 2e.

Further, FIG. 8 and FIG. 9 show that the control device 51 controls the operation of the flow rate adjusting motor 31 and the opening degree of the adjusting valve 15 in order to control the inertial mass Md and the damping coefficient of the variable rotary inertial mass damper 1. The indicated tuning control process is repeatedly executed at predetermined time intervals. Hereinafter, this tuning control process will be described with reference to FIGS. 8 and 9. As described above, the inertial mass Mh due to the viscous fluid HF contained in the inertial mass Md of the variable rotational inertial mass damper 1 tends to be very small as compared with the inertial mass Mr. due to the rotating mass 21, and thus the tuning control process. Then, ignoring this inertial mass Mh and regarding the inertial mass Mr by the rotating mass 21 as the inertial mass Md of the variable rotating inertial mass damper 1, the operation of the flow rate adjusting motor 31 and the opening degree of the adjusting valve 15 are controlled.

First, in step 1 of FIG. 8 (shown as “S1”; the same applies hereinafter), it is determined whether or not the building B is vibrating. This determination is based on the measurement signal of the seismograph 53, and when the seismic motion represented by the measurement signal is larger than a predetermined value, it is determined that the building B is vibrating. When the answer is NO and the building B is not vibrating, the current process is terminated as it is, as shown in FIGS. 8 and 9.

On the other hand, when the answer in step 1 is YES and the building B is vibrating, the predominant frequency fcp, which is the frequency of the predominant frequency component of the seismic motion input to the building B, is calculated (step 2). The calculation of the predominant frequency fcp is performed as follows, for example. That is, the seismic motion measured by the seismograph 53 is frequency-analyzed by the fast Fourier transform, and the Fourier amplitude spectrum of the seismic motion is calculated for each frequency (frequency) of the seismic motion. Then, the calculated plurality of Fourier amplitude spectra are compared with each other, and the frequency corresponding to the largest Fourier amplitude spectrum among them is set (calculated) as the predominant frequency fcp. The above-mentioned predetermined time, which is the execution cycle of the tuning control process, is set to a time sufficient to execute this frequency analysis. Since the above-mentioned method for calculating the Fourier amplitude spectrum is well known, its description is omitted, but it is disclosed in, for example, "Introduction to Spectral Analysis of New Seismic Motion Author: Yorihiko Osaki Kajima Institute Publishing".

Next, the neighborhood natural frequency fn is set based on the relationship between the calculated dominant frequency fcp and a plurality of predetermined natural frequencies of the building B (step 3). Specifically, the dominant frequency fcp is compared with a plurality of predetermined natural frequencies (natural frequencies of the primary to xth modes) of the building B stored in the ROM of the control device 51, and these plurality. Of the natural frequencies of, the natural frequency closest to the dominant frequency fcp is set as the near natural frequency fn.

For example, when the predominant frequency fcp is within the first predetermined frequency range including the natural frequency of the building B in the predetermined primary mode (hereinafter referred to as "primary natural frequency"), the near natural frequency fn is the primary natural. The frequency is set and fcp is within the second predetermined frequency range (> first predetermined frequency range) including the natural frequency of the predetermined secondary mode of the building B (hereinafter referred to as "secondary natural frequency"). Occasionally, fn is set to the second natural frequency. Such setting of the near natural frequency fn is performed using a plurality of predetermined frequency ranges set corresponding to the plurality of natural frequencies. From the above, the near natural frequency fn is set to the primary natural frequency when the primary natural frequency is closest to the dominant frequency fcp, and the secondary natural frequency is set to the secondary natural frequency when the secondary natural frequency is closest to the dominant frequency fcp. Is set to.

Next, using the set near natural frequency fn and the overall rigidity (spring constant) θs of the first and second elastic members EM1 and EM2, the inertial mass Mds for this control is calculated by the following equation (9). (Step 4). In this equation (9), the equation (= sqrt (θs / Md) / 2π) expressing the natural frequency of the additional vibration system described above is developed for the inertial mass Md of the variable rotational inertial mass damper 1, and Md is set to Mds. In addition, the natural frequency of the additional vibration system is replaced with fn.
Mds = θs / (fn · 2π) ² …… (9)

Further, the calculation of the inertial mass Mds in step 4 is performed, for example, for the variable rotary inertial mass damper 1 provided in each of all the layers of the building B. Hereinafter, the inertial mass Mds of the variable rotary inertial mass damper 1 provided on the i-layer of the building B is represented by appropriately adding the number of layers i of the building B as a subscript (see the formula (11) described later for _{Mds i).} .. Further, the overall rigidity θs of the first and second elastic members EM1 and EM2 are the predetermined values described above, are stored in the ROM, and are read out from the ROM in the calculation in step 4.

Next, by substituting the calculated inertial mass Mds into Mr of the above formula (7), the pump volume vm2, which is the flow rate of the viscous fluid HF in the second passage 6, is calculated (step 5). The pressure receiving area Ap of the piston 3 in this equation (7), the actual amount md of the rotating mass 21, the outer diameter D of the rotating mass 21, and the motor volume vm required for one rotation of the gear motor M are all predetermined values. , Stored in the ROM, and read from the ROM in the calculation of step 5. Next, the rotation speed n of the flow rate adjusting motor 31 is calculated by the above formula (8) according to the calculated pump volume vm2 (step 6). The delivery volume vmd per rotation of the flow rate adjusting motor 31 in the formula (8) is a predetermined value, is stored in the ROM, and is read from the ROM in the calculation in step 6. Next, the flow rate adjusting motor 31 is driven by outputting a drive signal based on the calculated rotation speed n (step 7).

By driving the flow rate adjusting motor 31 in this way, the inertial mass Mr by the rotating mass 21 regarded as the inertial mass Md of the variable rotary inertial mass damper 1 is controlled by the inertial mass Mds calculated in step 4. As a result, the natural frequency (= sqrt (θs / Md) / 2π) of the additional vibration system determined by the inertial mass Md and the rigidity θs of the first and second elastic members EM1 and EM2 becomes the same as the neighborhood natural frequency fn. Is controlled to be. The drive signal is obtained in advance by an experiment or the like so that the rotation speed n of the flow rate adjusting motor 31 can be obtained, and is stored in the ROM.

ここで、Ｌは建物Ｂの層の総数であり、_sｕ_iは、建物Ｂのｓ次モードでのｉ層の固有ベクトル（せん断成分及び曲げ成分を含む）、ｍ_iは建物Ｂのｉ層の質量である。この場合、次数ｓは、前記ステップ３において近傍固有振動数ｆｎとして設定された建物Ｂの固有振動数の次数に設定される。例えば、近傍固有振動数ｆｎが建物Ｂの１次固有振動数に設定された場合には、次数ｓは１に設定される。これらの建物Ｂの層の総数Ｌや、建物Ｂのｓ次モードでの１層〜Ｌ層の固有ベクトル_sｕ₁〜_sｕ_L、建物Ｂの１層〜Ｌ層の質量ｍ₁〜ｍ_Lはいずれも、所定値であり、予め算出されるとともにＲＯＭに記憶されており、上記のステップ８の算出においてＲＯＭから読み出される。 In step 8 of FIG. 9 following step 7, the broad nodal mass _s M ₀ of the building B is calculated by the following equation (10). This broadly _{defined node mass s} M ₀ is the equivalent mass of the building B when the s-th order mode of the building B is represented by a one-mass analysis model.

Here, L is the total number of layers of building B, _s u _i (including shear component and bending component) i layer of eigenvectors in the s-th mode of building B, m _i is the i layer of the building B It is the mass. In this case, the order s is set to the order of the natural frequency of the building B set as the neighborhood natural frequency fn in step 3. For example, when the neighborhood natural frequency fn is set to the primary natural frequency of the building B, the order s is set to 1. _{The total number L of the layers of the building B, the eigenvectors s} u ₁ to _s u _L of the 1st layer to the L layer in the s-th order mode of the _{building B, and the mass m 1 to mL} of the 1st layer to the L layer of the building B are All of them are predetermined values, are calculated in advance and are stored in the ROM, and are read from the ROM in the calculation in step 8 above.

ここで、_sｕ_i(s)−_sｕ_i-1(s)は、建物Ｂのｓ次モードでのｉ層の固有ベクトルのせん断成分と、建物Ｂのｓ次モードでの（ｉ−１）層の固有ベクトルのせん断成分との差であり、_sｕ_1(s)は、建物Ｂのｓ次モードでの１層の固有ベクトルのせん断成分である。この場合にも、次数ｓは、近傍固有振動数ｆｎとして設定された建物Ｂの固有振動数の次数に設定される。また、建物Ｂのｓ次モードでの１層〜Ｌ層の固有ベクトルのせん断成分_sｕ_1(s)〜_sｕ_L(s)はいずれも、所定値であり、予め算出されるとともにＲＯＭに記憶されており、上記のステップ９の算出においてＲＯＭから読み出される。 Next, using the inertial mass Mds calculated in step 4, the broadly defined inertial connecting element mass _s Md is calculated by the following equation (11) (step 9). The broadly defined inertial connection element mass _s Md is the inertial mass of the variable rotational inertial mass damper 1 in the broadly defined sth order mode of the additional vibration system.

Here, _s ui _(s) − _s u _{i-1 (s)} is the shear component of the eigenvector of the i-layer in the s-th order mode of the building B and (i-1) in the s-th order mode of the building B. It is the difference from the shear component of the eigenvector of the layer, and _s u _{1 (s)} is the shear component of the eigenvector of the layer in the s-th order mode of the building B. Also in this case, the order s is set to the order of the natural frequency of the building B set as the near natural frequency fn. _{Further, the shear components s} u _{1 (s)} to _s u _{L (s)} of the eigenvectors of the 1st layer to the L layer in the s-th order mode of the building B are all predetermined values, are calculated in advance, and are stored in the ROM. It is read from the ROM in the calculation of step 9 above.

In the above equation (11), unlike the case of the equation (10), the shear components _s u _{1 (s)} to _s u _{L (s)} of the eigenvectors of the 1st layer to the L layer in the s-th order mode of the building B are set. This is because the variable rotation inertial mass damper 1 operates as the vibration of the building B causes a horizontal displacement between the upper and lower beams BU and BD.

In step 10 following step 9, the mass ratio μ is calculated by dividing the _{calculated broad inertia connecting element mass s} Md by the broad nodal mass _s M _{0 calculated in step 8 (μ = s} Md /). _s M ₀ ).

Next, using the calculated mass ratio μ, the optimum tuning frequency ratio γ is calculated by the following equation (12) (step 11), and the cylinder 2, the piston 3, and the like are described above by the following equation (13). The target damping constant hobj, which is the target value of the damping constant of the variable damping damper, is calculated (step 12). Here, when the optimum tuning frequency ratio γ greatly exceeds 1, the process returns to step 4 and the neighborhood natural frequency fn in the equation (9) is replaced with γ · fn to recalculate the inertial mass Mds. Is desirable. As described above, when the optimum tuning frequency ratio γ greatly exceeds 1, the neighborhood natural frequency fn in the present specification may be replaced with γ · fn.

Next, using the calculated optimum tuning frequency ratio γ and the target damping constant hobj, and the near natural frequency fn and the inertial mass Mds calculated in steps 3 and 4, respectively, the variable damping is performed by the following equation (14). The target damping coefficient cobj, which is the target value of the damping coefficient of the damper, is calculated (step 13).
cobj = 2hobj ・ Mds ・ γ ・ fn ・ 2π …… (14)

The above equations (12) to (14) are derived so that the target attenuation coefficient cobj that minimizes the maximum value of the response magnification of the building B can be calculated according to the fixed point theory, and the derivation thereof. For the method, refer to pages 104 to 110 of "Displacement control design of buildings, co-authored by Norio Inoue / Yuki Igarashi, Maruzen Publishing".

As is clear from the calculation method of the target damping coefficient cobj according to the above steps 8 to 13, the variables used for the calculation are the near natural frequency fn and the inertial mass Mds, so that both fn and Mds and the target damping The target attenuation coefficient cobj may be calculated by obtaining the relationship with the coefficient cobj in advance and mapping it, and by searching this map according to fn and Mds. Further, although the above equations (12) to (14) are used for the calculation of the target attenuation coefficient cobj, other appropriate ones derived according to the fixed point theory so that the maximum value of the response magnification of the building B is minimized. An expression may be used.

In step 14 following step 13, the adjusting valve 15 is driven by outputting a drive signal based on the calculated target attenuation coefficient cobj, and the current process is completed. By driving the adjusting valve 15 in this way, the damping coefficient of the variable damping damper including the cylinder 2 and the piston 3 is controlled to be the target damping coefficient cobj. The drive signal is obtained in advance by an experiment or the like so that the attenuation coefficient becomes the target attenuation coefficient cobj, and is stored in the ROM. Further, the tuning control process including steps 1 to 14 is repeatedly executed at predetermined time intervals as described above.

Next, an operation example of the above-mentioned tuning control process will be described with reference to FIG. FIG. 10A shows the relationship between the seismic motion frequency (hereinafter referred to as “earthquake motion frequency”) fe input to the building B and the acceleration response magnification (hereinafter referred to as “response magnification”) RM of the predetermined layer of the building B with respect to the input. 10 (b) shows the relationship between the seismic motion frequency fe and the Fourier amplitude spectrum of the seismic motion (hereinafter referred to as “seismic motion spectrum”) SE, and FIG. 10 (c) shows the Fourier of the response between the seismic motion frequency fe and the predetermined layer of the building B. The relationship with the amplitude spectrum (hereinafter referred to as “building vibration spectrum”) SB is shown for each case where the predominant frequency fcp is equal to the first-order natural frequency of the building B. In FIG. 10, fe1 and fe2 indicate frequencies having the same magnitude as the primary and secondary natural frequencies of the building B, respectively.

Further, FIG. 10 (d) shows the relationship between the seismic motion frequency fe and the response magnification RM, FIG. 10 (e) shows the relationship between the seismic motion frequency fe and the seismic motion spectrum SE, and FIG. 10 (f) shows the relationship between the seismic motion frequency fe and the seismic motion frequency fe. The relationship with the building vibration spectrum SB is shown for each case where the predominant frequency fcp is equal to the second natural frequency of the building B.

As described with reference to FIGS. 8 and 9, in the tuning control process, the dominant frequency fcp (frequency of the dominant frequency component of the seismic motion input to the building B) is calculated (step 2), and the dominant frequency fcp is calculated. When the building B is close to the primary natural frequency, the nearby natural frequency fn is set to the primary natural frequency, while when it is close to the secondary natural frequency, fn is set to the secondary natural frequency (). Along with step 3), the inertial mass Md of the variable rotating inertial mass damper 1 including the inertial mass Mr. by the rotating mass 21 is controlled so that the natural frequency of the additional vibration system becomes the same as the nearby natural frequency fn (step 3). 4-7). Further, the target damping coefficient cobj is calculated using the equations (12) to (14) derived according to the fixed point theory according to the controlled inertial mass Md (inertial mass Mds) and the near natural frequency fn (step). 8-13), thereby cobj is calculated so that the maximum value of the response factor of the building B is minimized. Further, the damping coefficient of the variable damping damper that damps the vibration of the additional vibration system is controlled to be the calculated target damping coefficient cobj (step 14).

When the predominant frequency fcp is equal to the primary natural frequency of the building B and the seismic motion spectrum SB of the frequency fe1 equal to the primary natural frequency as shown in FIG. 10B is relatively large, the above-mentioned tuning control process is executed. As a result, the response magnification RM at fe = fe1 becomes smaller as shown in FIG. 10 (a), and the building vibration spectrum SB at fe = fe1 becomes smaller as shown in FIG. 10 (c). At any of the other frequencies, the building vibration spectrum SB is small.

Further, when the predominant frequency fcp is equal to the secondary natural frequency of the building B and the seismic motion spectrum SB of the frequency fe2 equal to the secondary natural frequency is relatively large as shown in FIG. 10 (e), the tuning control process described above is performed. As a result of the execution of FIG. 10 (d), the response magnification RM at fe = fe2 becomes smaller as shown in FIG. 10 (d), and the building vibration spectrum SB at fe = fe2 becomes smaller as shown in FIG. 10 (f). As in the case of (c), the building vibration spectrum SB is small at any other frequency of the seismic motion frequency fe.

On the other hand, FIGS. 11A to 11C show an operation example of the above-mentioned conventional vibration suppression device, and in these figures, feC, feC1, feC2, RMC, SEC and SBC are shown in FIG. These are the parameters corresponding to fe, fe1, fe2, RM, SE, and SB, respectively. As described above, in the conventional vibration suppression device, the inertial mass of the rotary inertial mass damper cannot be controlled, and the natural frequency of the additional vibration system is always tuned to the primary natural frequency of the building.

As a result, as shown in FIG. 11A, the acceleration response magnification RMC of the building at the frequency feC1 equal to the primary natural frequency is small, whereas at the frequency feC2 equal to the secondary natural frequency. The acceleration response magnification RMC of the building is relatively large. In this case, the Fourier amplitude spectrum SEC of the seismic motion of the frequency feC2 in which the predominant frequency of the seismic motion input to the building is equal to the secondary natural frequency of the building and equal to the secondary natural frequency as shown in FIG. When it is relatively large, as shown in FIG. 11C, the Fourier amplitude spectrum SBC of the building response at feC = feC2 is relatively large, and the vibration of the secondary natural frequency of the building cannot be appropriately suppressed. But I understand.

Further, FIGS. 11 (d) to 11 (f) show an operation example when the natural frequency of the additional vibration system is always tuned to the secondary natural frequency of the building, unlike the first embodiment. In these figures, fen, fc1, fc2, RMc, SEc and SBc are parameters corresponding to fe, fe1, fe2, RM, SE and SB in FIG. 10, respectively. As shown in FIGS. 11 (d) to 11 (f), when the natural frequency of the additional vibration system is always tuned to the secondary natural frequency of the building, the seismic motion input to the building When the predominant frequency is equal to the first natural frequency of the building, the Fourier amplitude spectrum SBc of the building at fc = fc1 is relatively large, and the vibration of the first natural frequency of the building cannot be suppressed appropriately. I understand.

As described above, in the comparative example shown in FIG. 11, since the natural frequency of the additional vibration system is always fixed because the inertial mass of the rotary inertial mass damper cannot be controlled (variable), it is sometimes input to the building. It can be seen that the vibration of the building cannot be properly suppressed in response to the earthquake motion of.

On the other hand, in the first embodiment described above, by executing the tuning control process using the variable rotation inertial mass damper 1, the variable rotation inertial mass damper is responded to the occasional seismic motion input to the building B. It can be seen that the inertial mass Md and the damping coefficient of 1 can be appropriately controlled, and thus the vibration of the building B can be appropriately suppressed.

As described above, according to the first embodiment, as described with reference to FIGS. 1 to 7, the additional vibration system is formed by the variable rotation inertial mass damper 1 and the first and second elastic members EM1 and EM2. It is configured. The displacement of the building B due to the vibration is transmitted to the variable rotary inertial mass damper 1 via the first and second elastic members EM1 and EM2, and the transmitted displacement of the building B is the cylinder 2, the piston 3, and the viscous fluid. The HF, the first communication passage 5, and the gear motor M convert the rotary mass 21 into a rotary motion, whereby the rotary mass 21 rotates, and as a result, the additional vibration system vibrates. The inertial mass Md of the variable rotary inertial mass damper 1 generated with the rotation of the rotary mass 21 is continuously changed by adjusting the flow rate of the viscous fluid HF in the second communication passage 6 via the gear pump P. Will be done.

Further, as described with reference to FIGS. 8 and 9, the predominant frequency fcp, which is the frequency of the predominant frequency component of the vibration input to the building B, is calculated (step 2). Further, during the vibration of the building B, the natural frequency of the additional vibration system is added to the near natural frequency fn, which is the natural frequency closest to the calculated dominant frequency fcp, among a plurality of predetermined natural frequencies of the building B. More specifically, the flow rate of the viscous fluid HF in the second communication passage 6 via the gear pump P so that the natural frequency of the additional vibration system becomes the same as the nearby natural frequency fn. The inertial mass Md of the variable rotation inertial mass damper 1 is controlled by adjusting the above (steps 4 to 7). As a result, by synchronizing the natural frequency of the additional vibration system with the natural frequency fn in the vicinity of the building B, the vibration of the building B at the strongest dominant frequency fcp included in the occasional vibration input to the building B can be obtained. , Can be appropriately absorbed and suppressed by the additional vibration system.

Further, when the pressure in the first fluid chamber 2d reaches the upper limit value, the first relief valve 11 is opened, and when the pressure in the second fluid chamber 2e reaches the upper limit value, the second relief valve is opened. When the valve 12 is opened, the increased pressure of one of the first and second fluid chambers 2d and 2e is released to the other, so that the pressure is prevented from being excessive, and the cylinder 2 and the piston 3 of the variable rotary inertial mass damper 1 are prevented. The axial force acting on the can be appropriately limited.

Further, the vibration of the additional vibration system is damped by the variable damping damper including the cylinder 2, the piston 3, the first communication passage 5, the viscous fluid HF, and the adjusting valve 15, and the opening degree of the adjusting valve 15 is changed. The damping coefficient of the variable damping damper is continuously changed. Further, during the vibration of the building B, the response of the building B according to the near natural frequency fn and the inertial mass Md (inertial mass Mds) of the variable rotation inertial mass damper 1 controlled as described above according to the fixed point theory. Since the damping coefficient of the variable damping damper is controlled so that the maximum value of the magnification is minimized (steps 8 to 14), the vibration of the building B can be suppressed more appropriately. Further, the variable rotation inertial mass damper 1 also has the function of a variable damping damper, and the cylinder 2, the piston 3, the first communication passage 5, and the viscous fluid HF are the components of the variable rotation inertial mass damper and the variable damping damper. Since it is also used as a vibration suppressor, the vibration suppression device can be miniaturized accordingly.

Further, the variable rotary inertial mass damper 1 basically includes a cylinder 2, a piston 3, a first and second communication passages 5 and 6, a gear motor M is provided in the first passage 5, and a gear pump P is provided. Compared with the conventional variable rotation inertial mass damper disclosed in Japanese Patent Application Laid-Open No. 2016-151287, which requires a ball screw mechanism, a weight portion holding mechanism composed of a large number of members, and the like because the configuration is provided in the double passage 6. And the configuration is simple. For the same reason, the ball screw mechanism, weight holding mechanism, and actuator are all easier to assemble, and adjustments and maintenance are performed compared to this conventional variable rotary inertial mass damper in which the actuator is housed in the casing. Workability can be improved.

In the first embodiment, the first and second passages 5 and 6 are directly connected to the cylinder 2, but both ends of the second passage 6 are placed in the middle of the first passage 5 with gears. They may be connected in parallel so as to straddle the motors M (see FIG. 7 of Japanese Patent Application No. 2016-202281 by the applicant). Further, in the first embodiment, the gear motor M is used as the flow conversion mechanism in the present invention, but another suitable mechanism capable of converting the flow of the viscous fluid HF into rotational motion, for example, the patent No. 1 by the present applicant. The screw mechanism described in FIG. 5 of No. 5191579, the ball screw in which the piston described in FIG. 2 of Patent No. 5161395 by the applicant is integrally provided on the nut, or the vane motor or plunger motor (piston). A motor) or the like may be used. When this ball screw is used as the flow conversion mechanism, it goes without saying that the portion where the piston moves in the communication passage may be formed in a cylinder shape.

Further, in the first embodiment, the regulating valve 15 composed of the solenoid valve is used, but a regulating valve of a type driven by hydraulic pressure or pneumatic pressure may be used. Further, in the first embodiment, the gear pump P is used as the electric pump constituting the flow rate adjusting mechanism in the present invention, but other suitable electric pumps such as a vane pump and a plunger pump (piston pump) are used. You may. Further, in the first embodiment, the gear pump P is used as the flow rate adjusting mechanism, but another suitable mechanism capable of adjusting the flow rate of the viscous fluid HF (working fluid) flowing in the second communication passage 6, for example. , A flow rate adjusting valve that adjusts the flow rate of the working fluid by changing the opening degree may be used. Further, in the first embodiment, the first and second relief valves 11 and 12 are provided on the piston 3, but these may be omitted.

Further, in the first embodiment, the variable damping damper in the present invention is composed of a cylinder 2, a piston 3, a first communication passage 5, a viscous fluid HF, a regulating valve 15, and the like, and the variable rotational inertia mass damper 1 is a variable damping damper. However, the variable rotation inertial mass damper having the configuration in which the adjusting valve 15 is removed from the variable rotation inertial mass damper 1 and the damping coefficient are changed while attenuating the vibration of the additional vibration system. Possible variable damping dampers may be provided in parallel. As the variable damping damper in this case, various variable damping dampers such as the variable damping damper 81 of the second embodiment described later and the variable damping damper of the type using MR fluid (Magneto-Rheological fluid) may be used. Of course. Alternatively, the variable damping damper (adjusting valve 15) may be omitted, and the processing for controlling the damping coefficient in the tuning control processing (steps 8 to 14) may be omitted.

Further, in the first embodiment, the inertial mass Md of the variable rotation inertial mass damper 1 is controlled so that the natural frequency of the additional vibration system becomes the same as the near natural frequency fn. It may be controlled so that the natural frequency becomes substantially the same as the nearby natural frequency fn. In that case, instead of the neighborhood natural frequency fn in the above equation (9), a value obtained by adding or subtracting a very small predetermined value to fn is used.

Further, in the first embodiment (synchronization control processing), the inertial mass Mh due to the viscous fluid HF is ignored, the inertial mass Mr by the rotating mass 21 is regarded as the inertial mass Md of the variable rotating inertial mass damper 1, and the flow rate adjusting motor 31 Although the operation of the above and the opening degree of the adjusting valve 15 are controlled, the control may be performed with Md = Mr + Mh without ignoring the inertial mass Mh. In this case, in the processing after step 4, the value obtained by subtracting the inertial mass Mh from the inertial mass Mds calculated by the equation (9) is used as the inertial mass Mds. In addition, it goes without saying that the variations described above with respect to the first embodiment may be appropriately combined and applied.

Next, the vibration suppression device according to the second embodiment of the present invention will be described with reference to FIGS. 12 to 16. This vibration suppression device is applied to the building B (see FIG. 15) as in the first embodiment, and includes the variable rotation inertial mass damper 61 shown in FIG. 12 and the variable damping damper 81 shown in FIG. It includes a control device 91 shown in FIG. 14, a power supply 92, a seismograph 53, and first and second elastic members EM1'and EM2' shown in FIG. The variable rotary inertial mass damper 61 is configured so that the inertial mass generated as the rotating masses 68 and 68, which will be described later, rotate can be changed in two stages, and as shown in FIG. 12, the case-shaped main body portion. 62, a cylindrical inner cylinder 63 and an outer cylinder 64 housed in the main body 62, and a screw shaft 65 movably provided in the axial direction (left-right direction in FIG. 12) with respect to the main body 62. The screw shaft 65 has a nut 66 that is rotatably screwed via a plurality of balls (not shown).

The main body portion 62 integrally has a tubular tubular wall 62a and plate-shaped first end walls 62b and second end walls 62c provided at both ends of the tubular wall 62a in the axial direction, respectively. A thrust bearing BE1 is attached and fixed to each of the surfaces of the first and second end walls 62b and 62c facing each other. These pair of thrust bearings BE1 and BE1 are arranged coaxially with each other. The first end wall 62b is integrally provided with a convex portion 62d protruding outward in the thickness direction, and the convex portion 62d does not rotate with the torque acting as the rotating masses 68 and 68 rotate. The first fitting FL1'is provided via a universal joint having a degree of friction. Further, an insertion hole 62e penetrating in the thickness direction is formed in the second end wall 62c, and the convex portion 62d and the insertion hole 62e are arranged coaxially with the thrust bearings BE1 and BE1. Further, support holes 62f penetrating in the thickness direction are formed in the first and second end walls 62b and 62c in parallel with the above insertion holes 62e, and radial bearings BE2 are formed in the support holes 62f. Is provided.

The inner cylinder 63 integrally has a cylindrical peripheral wall 63a, and disk-shaped first end walls 63b and second end walls 63c provided at both ends of the peripheral wall 63a in the axial direction, respectively, and has a main body. It is arranged coaxially with the nut 66 in the portion 62. The first end wall 63b is engaged with the bearing BE1 of the first end wall 62b of the main body 62, and the second end wall 63c is engaged with the bearing BE1 of the second end wall 62c of the main body 62, respectively. The inner cylinder 63 is rotatably supported by the main body 62 via bearings BE1 and BE1 about the axis thereof, and is immovable with respect to the main body 62. Further, an insertion hole (not shown) penetrating in the axial direction is formed in the center of the second end wall 63c in the radial direction. The outer cylinder 64 is formed in a cylindrical shape, and the inner cylinder 63 is coaxially inserted inside the outer cylinder 64, and is rotatably supported by the inner cylinder 63 via a radial bearing (not shown). It cannot move in the axial direction with respect to 63.

The screw shaft 65 constitutes a ball screw together with the ball and the nut 66. Further, the screw shaft 65 extends in the axial direction and is partially housed in the inner cylinder 63 so as to be movable in the axial direction in a state of being inserted into the insertion hole of the second end wall 63c thereof. It extends outward from the second end wall 62c of 62. At the end of the screw shaft 65 opposite to the inner cylinder 63, a second fitting FL2'is provided via a universal joint having friction that does not rotate with the torque acting with the rotation of the rotating masses 68 and 68. It is provided. The nut 66 is coaxially attached to the second end wall 63c of the inner cylinder 63 and inserted into the insertion hole 62e of the second end wall 62c of the main body 62. It is integrally rotatable.

With the above configuration, when the screw shaft 65 moves in the axial direction with respect to the main body 62, this movement is converted into a rotary motion by the ball screw, and as a result, the nut 66 and the inner cylinder 63 rotate.

Further, the variable rotation inertial mass damper 61 rotates in a state where the clutch 67 for connecting / disconnecting between the inner cylinder 63 and the outer cylinder 64, the pair of rotating masses 68 and 68, and the outer cylinder 64 are rotated at different speeds. Further, a transmission mechanism 69 for transmitting to the masses 68 and 68 is provided. The clutch 67 is, for example, a friction clutch, and has an inner attached to the inner cylinder 63, an outer attached to the outer cylinder 64, and an electromagnetic actuator 67a (see FIG. 14). The actuator 67a is connected to the power supply 92 via the control device 91, and the connection / disconnection of the clutch 67 is controlled by the control device 91 via the actuator 67a. When the nut 66 and the inner cylinder 63 rotate as the screw shaft 65 moves with respect to the main body 62 while the clutch 67 is engaged, the outer cylinder 64 rotates integrally with the inner cylinder 63. A hydraulic actuator 67a may be used. Further, as the clutch 67, another suitable clutch such as an electromagnetic powder clutch may be used.

Each rotary mass 68 is made of a material having a relatively large specific gravity, for example, iron, and is formed in a disk shape. The thickness thereof is relatively large, and one of the pair of rotary masses 68, 68 is a transmission mechanism 69. The other end of the rotating shaft 70, which will be described later, is provided coaxially with the other end of the rotating shaft 70. Of course, one of the pair of rotating masses 68 and 68 may be omitted.

The speed change mechanism 69 is a so-called parallel shaft type stepped speed change mechanism, and is provided in parallel with the rotary shaft 70 inserted into the above-mentioned insertion holes 62f and 62f of the main body 62 via the radial bearings BE2 and BE2. It is composed of the first gear train 71 and the second gear train 72, a synchromesh mechanism (synchronous meshing mechanism) 73, and the like, and has two gears. The rotating shaft 70 is rotatably supported by the main body 62 via radial bearings BE2 and BE2, and extends in parallel with the outer cylinder 64. The first and second gear rows 71 and 72 and the synchromesh mechanism 73 are housed in the main body 62.

The first gear row 71 is composed of a first gear 71a and a second gear 71b that mesh with each other. The former 71a is provided coaxially with the outer cylinder 64, and the latter 71b is coaxially provided with the rotating shaft 70. It is rotatably provided. Further, the second gear row 72 is composed of a third gear 72a and a fourth gear 72b that mesh with each other, the former 72a is provided coaxially with the outer cylinder 64, and the latter 72b is coaxial with the rotating shaft 70. It is rotatably provided in a shape. The setting of the number of teeth of the first to fourth gears 71a, 71b, 72a and 72b will be described later.

The synchromesh mechanism 73 is for selectively connecting / disconnecting the second gear 71b and the fourth gear 72b to the rotating shaft 70, and is a sleeve 73a formed in an annular shape and provided with internal teeth, or a sleeve. A shift fork 73b connected to the 73a, an electromagnetic actuator 73c (see FIG. 14) for driving the sleeve 73a via the shift fork 73b, and clutch gears integrally provided for each of the second and fourth gears 71b and 72b. (Not shown) and so on. Since the synchromesh mechanism 73 is a well-known one used in a stepped speed change mechanism for a vehicle, its configuration and operation will be briefly described.

The sleeve 73a is provided on the rotating shaft 70 so as to be non-rotatable and movable in the axial direction, and is arranged between the second gear 71b and the fourth gear 72b. In the synchromesh mechanism 73, when the sleeve 73a is in the neutral position shown in FIG. 12, the second and fourth gears 71b and 72b and the rotating shaft 70 are in a cut-off state. Further, when the sleeve 73a is driven from the neutral position to the second gear 71b side by the actuator 73c, the sleeve 73a moves to the second gear 71b side while synchronizing the rotations of the second gear 71b and the rotating shaft 70 with each other. When the internal teeth of the sleeve 73a mesh with the clutch gear integrated with the second gear 71b, the second gear 71b is connected to the rotating shaft 70, and the space between the fourth gear 72b and the rotating shaft 70 is cut off.

On the other hand, when the sleeve 73a is driven from the neutral position to the fourth gear 72b side by the actuator 73c, the sleeve 73a moves to the fourth gear 72b side while synchronizing the rotations of the fourth gear 72b and the rotating shaft 70 with each other. When the internal teeth of the sleeve 73a mesh with the clutch gear integrated with the fourth gear 72b, the fourth gear 72b is connected to the rotating shaft 70, and the space between the second gear 71b and the rotating shaft 70 is cut off.

The actuator 73c is connected to the power supply 92 via the control device 91, and the synchromesh mechanism 73 selectively connects / disconnects the second and fourth gears 71b and 72b to the rotating shaft 70 via the actuator 73c. It is controlled by the control device 91. As a result, as the gear train of the transmission mechanism 69 for transmitting the rotation of the outer cylinder 64 to the rotation masses 68 and 68, the first gear train 71 composed of the first and second gears 71a and 71b, and the third and fourth gear trains One of the second gear rows 72 composed of gears 72a and 72b is selected.

Further, the first and second elastic members EM1'and EM2' are made of elastic steel materials having relatively low rigidity, and their rigidity is set to a predetermined value, and the upper beam BU and the lower part of the building B are set. It is fixed to the beam BD and extends downward from the upper beam BU and upward from the lower beam BD. Further, the first and second fittings FL1'and FL2'are attached to the first and second elastic members EM1'and EM2', respectively, and the first and second fittings described later of the variable damping damper 81 are attached. FL1 and FL2 are attached respectively. As described above, the variable rotation inertial mass damper 61 and the variable damping damper 81 are connected in parallel to the upper beam BU via the first elastic member EM1'and to the lower beam BD via the second elastic member EM2'. They are connected in parallel with each other and extend horizontally between both BDs and BUs. Further, the variable rotary inertial mass damper 61 and the first and second elastic members EM1'and EM2' constitute an additional vibration system.

Further, a variable rotation inertial mass damper 61, a variable damping damper 81, and one set of the first and second elastic members EM1'and EM2' are provided for each of all the layers of the building B, and FIG. 15 shows one of them. One set is shown. Note that in FIG. 15, for convenience, some components and reference numerals are omitted. By the way, it goes without saying that the variable rotational inertial mass damper 61, the variable damping damper 81, and the first and second elastic members EM1'and EM2' may be provided not only in all the layers of the building B but only in some layers. Is.

Next, the operation of the variable rotation inertial mass damper 61 will be described. When a relative displacement in the horizontal direction occurs between the upper and lower beams BU and BD as the building B vibrates, this relative displacement is caused to the main body via the first and second elastic members EM1'and EM2'. By being transmitted as an external force to the 62 and the screw shaft 65, the screw shaft 65 moves in the axial direction with respect to the main body portion 62. The movement of the screw shaft 65 is converted into a rotary motion by the nut 66, and the nut 66 rotates with respect to the main body portion 62 together with the inner cylinder 63.

In this case, when the inner cylinder 63 and the outer cylinder 64 are connected by the clutch 67 and the first gear row 71 is selected by the synchromesh mechanism 73, the rotation of the inner cylinder 63 is the outer cylinder 64, the first. A state in which the gear is changed at a predetermined first gear ratio (number of teeth of the second gear 71b / number of teeth of the first gear 71a) based on the gear ratios of both gears 71a and 71b via the first and second gears 71a and 71b. Then, it is transmitted to the rotating shaft 70, and further transmitted to the rotating masses 68 and 68. As a result, the rotating masses 68 and 68 rotate, and as a result, an inertial force due to the inertial mass caused by the rotation of the rotating masses 68 and 68 is generated.

On the other hand, when the inner cylinder 63 and the outer cylinder 64 are connected by the clutch 67 and the second gear row 72 is selected by the synchromesh mechanism 73, the rotation of the inner cylinder 63 is the outer cylinder 64, the third and the third. Rotate in a state of being shifted at a predetermined second gear ratio (number of teeth of the fourth gear 72b / number of teeth of the third gear 72a) based on the gear ratios of both gears 72a and 72b via the four gears 72a and 72b. It is transmitted to the shaft 70, and further transmitted to the rotating masses 68 and 68. As a result, the rotating masses 68 and 68 rotate, and as a result, an inertial force due to the inertial mass caused by the rotation of the rotating masses 68 and 68 is generated.

The gear row selected by the synchromesh mechanism 73 is changed between the first gear row 71 and the second gear row 72, and the gear ratio of the transmission mechanism 69 is switched between the first gear ratio and the second gear ratio. When this is done, first, the clutch 67 shuts off between the inner cylinder 63 and the outer cylinder 64, and then the gear train is changed in that state, and the connection of the changed gear train to the gear rotation shaft 70 is completed. After that, the clutch 67 connects the inner cylinder 63 and the outer cylinder 64.

Further, when the first gear train 71 is selected, the inertial mass Md'(equivalent mass) due to the rotating masses 68 and 68 acting on the screw shaft 65 is represented by the following equation (15).
Md'= {2π / (Ld ・ R1)} ²・ md'・ D' ² /8 …… (15)
Here, R1 is the above-mentioned first gear ratio (number of teeth of the second gear 71b / number of teeth of the first gear 71a), Ld is the pitch of the screw shaft 65, and md'is the substance of the rotating masses 68 and 68. The quantity, D', is the diameter of the rotating mass 68.

On the other hand, the inertial mass Md'is represented by the following equation (16) when the second gear train 72 is selected.
Md'= {2π / (Ld ・ R2)} ²・ md'・ D' ² /8 …… (16)
Here, R2 is the second gear ratio (the number of teeth of the fourth gear 72b / the number of teeth of the third gear 72a).

As is clear from these equations (15) and (16), the inertial mass Md'of the variable rotation inertial mass damper 61 selects one of the first and second gear trains 71 and 72, and shifts the speed change mechanism 69. By changing the ratio, it is changed in two stages.

Next, the setting of the number of teeth of the first to fourth gears 71a to 72b will be described. The natural frequency fa of the additional vibration system including the rotating masses 68 and 68 and the first and second elastic members EM1'and EM2' is represented by the following equation (17).
fa = sqrt (θs'/ Md') / 2π …… (17)
Here, θs'is the overall rigidity (spring constant) of the first and second connecting members EN1'and EN2'.

The number of teeth of the first and second gears 71a and 71b is the natural frequency of the additional vibration system when the first gear row 71 is selected and the first gear ratio R1 is selected as the gear ratio of the transmission mechanism 69. The fa is set so as to be synchronized with, for example, the primary natural frequency (natural frequency of the primary mode) f1 of the building B (for example, fa = f1 or fa≈f1). More specifically, based on the above equations (15) and (17), the first gear ratio R1, the primary natural frequency f1, the pitch Ld of the screw shaft 65, the actual amount md'of the rotating masses 68 and 68, The first and second gears 71a so that the following equation (18) holds between the diameter D'of the rotating mass 68 and the overall rigidity θs' of the first and second connecting members EN1'and EN2'. , 71b has a set number of teeth.
R1 = {(2π ²・ f1 ・ D') / Ld} ・ sqrt {md'/ (2θs')}
…… (18)

The above equation (18) is an example of setting the number of teeth of the first and second gears 71a and 71b so that fa = f1, but when setting so that fa≈f1. , Instead of the primary natural frequency f1 of the above equation (18), a value obtained by adding or subtracting a very small predetermined value to f1 is used.

Further, the number of teeth of the third and fourth gears 72a and 72b is unique to the additional vibration system when the second gear row 72 is selected and the second gear ratio R2 is selected as the gear ratio of the transmission mechanism 69. The frequency fa is set so as to be synchronized with, for example, the secondary natural frequency (natural frequency in the secondary mode) f2 of the building B (for example, fa = f2 or fa≈f2). More specifically, based on the above equations (16) and (17), the second gear ratio R2, the secondary natural frequency f2, the pitch Ld of the screw shaft 65, the mass md'of the rotating masses 68 and 68, and the rotation. The third and fourth gears 72a, so that the following equation (19) holds between the diameter D'of the mass 68 and the overall rigidity θs' of the first and second connecting members EM1'and EM2'. The number of teeth of 72b is set.
R2 = {(2π ²・ f2 ・ D') / Ld} ・ sqrt {md'/ (2θs')}
…… (19)

The above equation (19) is an example of setting the number of teeth of the third and fourth gears 72a and 72b so that fa = f2, but when setting the number of teeth so that fa≈f2 , Instead of the secondary natural frequency f2 of the above equation (19), a value obtained by adding or subtracting a very small predetermined value to f2 is used.

The first and second gear ratios R1 and R2 are both set to gear ratios on the high speed side smaller than the value of 1.0, whereby the inertial mass Md'of the variable rotational inertial mass damper 61 is set. As is clear from the above equations (15) and (16), the actual amount of the rotating masses 68 and 68 is increased more than md'.

Next, the variable damping damper 81 will be described. The variable damping damper 81 is configured so that its damping coefficient can be continuously changed. As shown in FIG. 13, a cylindrical cylinder 82 and a piston provided in the cylinder 82 so as to be slidable in the axial direction. The rod 84, which is integrally provided with the 83 and the piston 83 and is partially accommodated in the cylinder 82 so as to be movable in the axial direction, the communication passage 85 connected to the cylinder 82, and the adjustment provided in the communication passage 85. It has a valve 86.

The cylinder 82 is configured in the same manner as the cylinder 2 of the first embodiment, and integrally has a peripheral wall 82a, a first end wall 82b, and a second end wall 82c. The fluid chamber defined by the peripheral walls 82a, the first and second end walls 82b, 82c is the first fluid chamber 82d on the first end wall 82b side and the second fluid on the second end wall 82c side by the piston 83. It is partitioned into chambers 82e, and both fluid chambers 82d and 82e are filled with viscous fluid HF. Further, the first end wall 82b is integrally provided with a convex portion 82f protruding outward, and the convex portion 82f is provided with a first attachment FL1 via a universal joint. The first attachment FL1 is attached to the first elastic member EM1'as described above (see FIG. 15). Further, a rod guide hole 82g penetrating in the axial direction is formed in the center of the second end wall 82c in the radial direction, and a seal is provided in the rod guide hole 82g.

The rod 84 is liquid-tightly inserted into the rod guide hole 82g via a seal, extends in the axial direction, and is movable in the axial direction with respect to the cylinder 82, and one end thereof is attached to the piston 83. Has been done. Further, a second attachment FL2 is provided at the other end of the rod 84 via a universal joint. The second attachment FL2 is attached to the second elastic member EM2'as described above (see FIG. 15).

The piston 83 is formed in a columnar shape, and its outer peripheral surface is in liquid-tight contact with the inner peripheral surface of the cylinder 82 via a seal. Further, a plurality of holes penetrating in the axial direction are formed at the outer end portion of the piston 83 in the radial direction (only two holes are shown), and the first and second relief valves 11 described above are formed in these holes. A first relief valve 87 and a second relief valve 88 having the same configurations as those of the above and 12 are provided.

The first relief valve 87 opens when the pressure of the viscous fluid HF in the first fluid chamber 82d reaches a predetermined upper limit value due to the movement of the piston 83 toward the first end wall 82b, whereby the first relief valve 87 opens. By communicating the 1 and the second fluid chambers 82d and 82e with each other, it is possible to prevent the pressure of the viscous fluid HF in the first fluid chamber 82d from becoming excessive. The second relief valve 88 opens when the pressure of the viscous fluid HF in the second fluid chamber 82e reaches the above upper limit value due to the movement of the piston 83 toward the second end wall 82c side, whereby the second relief valve 88 opens. By communicating the 2 and the first fluid chambers 82e and 82d with each other, it is possible to prevent the pressure of the viscous fluid HF in the second fluid chamber 82e from becoming excessive. The upper limit values of the first and second relief valves 87 and 88 may be set to different values.

The communication passage 85 is connected to the cylinder 82 so as to bypass the piston 83 and communicate with the first and second fluid chambers 82d and 82e, and is filled with the viscous fluid HF. The adjusting valve 86 is for adjusting the flow resistance of the viscous fluid HF flowing through the communication passage 85, and is composed of, for example, a normally open type solenoid valve, and its opening degree can be continuously changed. .. Further, the adjusting valve 86 is connected to the power supply 92 via the control device 91, and its opening degree is controlled by the control device 91.

Next, the operation of the variable damping damper 81 will be described. When a relative displacement in the horizontal direction occurs between the upper and lower beams BU and BD as the building B vibrates, this relative displacement is caused by the cylinder 82 via the first and second elastic members EM1'and EM2'. And by being transmitted to the rod 84 as an external force, the cylinder 82 and the rod 84 move relatively in the axial direction, and the piston 83 slides in the cylinder 82.

In this case, when the piston 83 moves to the first fluid chamber 82d side (left side in FIG. 13), a part of the viscous fluid HF in the first fluid chamber 82d is pushed out to the communication passage 85 by the piston 83. , The flow of the viscous fluid HF to the second fluid chamber 82e side (right side) occurs in the communication passage 85. On the contrary, when the piston 83 moves to the second fluid chamber 82e side (right side), a part of the viscous fluid HF in the second fluid chamber 82e is pushed out to the communication passage 85 by the piston 83. , The flow of the viscous fluid HF to the first fluid chamber 82d side (left side) occurs in the communication passage 85.

Further, in any case where the piston 83 moves to the first fluid chamber 82d side and the second fluid chamber 82e side, the viscosity between the first and second fluid chambers 82d and 82e is accompanied by the movement of the piston 83. A pressure difference in the fluid HF is generated, and this pressure difference acts as a resistance force on the piston 83. This resistance force, that is, the damping force of the variable damping damper 81 acts to damp the vibration of the additional vibration system, and the damping coefficient adjusts the opening degree of the adjusting valve 86 and the viscosity flowing through the communication passage 85. It is continuously changed by changing the flow resistance of the fluid HF.

The control device 91 and the power supply 92 are configured in the same manner as the control device 51 and the power supply 52 of the first embodiment, respectively, and the control device 91 is provided with the seismic motion measured by the seismograph 53 and input to the building B. A measurement signal representing is input. The control device 91 sets the gear ratio of the speed change mechanism 69 and controls the opening degree of the adjusting valve 86 in order to control the inertial mass Md'of the variable rotation inertial mass damper 61 and the damping coefficient of the variable damping damper 81. , The tuning control process shown in FIG. 16 is repeatedly executed at predetermined time intervals.

In FIG. 16, the same step numbers are assigned to the parts having the same execution contents as those of the tuning control process (FIGS. 8 and 9) of the first embodiment. Hereinafter, this tuning control process will be described mainly on the part of the execution content different from that of the first embodiment.

In step 21 following step 2 of FIG. 16, based on the relationship between the calculated dominant frequency fcp and the primary and secondary natural frequencies f1 and f2 of the building B stored in the ROM of the control device 91. The neighborhood natural frequency fn is set. Specifically, when the dominant frequency fcp is smaller than the value obtained by adding a predetermined value to the primary natural frequency f1, the primary natural frequency f1 of the primary and secondary natural frequencies f1 and f2 is the dominant frequency. Assuming that it is closest to the number fcp, the near natural frequency fn is set to the primary natural frequency f1. On the other hand, when the dominant frequency fcp is equal to or greater than the value obtained by adding a predetermined value to the primary natural frequency f1, the secondary natural frequency f2 of the primary and secondary natural frequencies f1 and f2 becomes the dominant frequency fcp. Assuming that it is the closest, the near natural frequency fn is set to the secondary natural frequency f2.

Next, the gear ratio of the speed change mechanism 69 is set (step 22) based on the near natural frequency fn. In this step 22, when the near natural frequency fn is set to the primary natural frequency f1, the speed change mechanism is used to synchronize the natural frequency fa of the additional vibration system with the primary natural frequency f1 of the building B. The gear ratio of 69 is set (controlled) to the first gear ratio R1. Further, when the near natural frequency fn is set to the secondary natural frequency f2, the speed change mechanism 69 is used to synchronize the natural frequency fa of the additional vibration system with the secondary natural frequency f2 of the building B. The ratio is set (controlled) to the second gear ratio R2. Along with this, the operation of the actuators 67a and 73c is controlled according to the set gear ratio.

Next, using the set near natural frequency fn and the overall rigidity θs'of the first and second elastic members EM1'and EM2', it is controlled by setting the gear ratio in step 22 by the following equation (20). The inertial mass Mds' is calculated (step 23). In this equation (20), the equation (17) (fa = sqrt (θs' / Md') / 2π) representing the natural frequency fa of the additional vibration system is expanded with respect to the inertial mass Md', and Md'is expressed. Mds'and fa are replaced with fn, respectively.
Mds'= θs' / (fn ・ 2π) ² …… (20)

Further, the calculation of the inertial mass Mds'is performed, for example, on the rotating masses 68 and 68 provided in each of all the layers of the building B. Hereinafter, the inertial mass Mds'by the rotating masses 68 and 68 provided on the i-layer of the building B is represented by appropriately adding the number of layers i as a subscript (see the formula (21) described later in _{Mds' i).} Further, the overall rigidity θs'of the first and second elastic members EM1'and EM2' are predetermined values, are stored in the ROM, and are read out from the ROM in the calculation in step 23 above.

Next, the step 8 is executed, and as described in the first embodiment, the broad nodal mass _s M ₀ of the building B is calculated. Next, using the inertial mass Mds'calculated in step 23 above, the broadly defined inertial connecting element mass _s Md'is calculated by the following equation (21) (step 24). The broadly defined inertial connection element mass _s Md'is the inertial mass of the variable rotational inertial mass damper 61 in the broadly defined sense in the sth order mode of the additional vibration system. _{The s} u _{i (s)} − _s u _{i-1 (s)} and _s u _{1 (s} ) in the equation (21) are as described in the first embodiment.

Next, the mass ratio μ'is calculated by dividing the calculated broad inertia connecting element mass _s Md'by the broad nodal mass _s M _{0 calculated in step 8 (μ'= s} Md'/ _s M). ₀ ) (step 25).

Next, using the calculated mass ratio μ', the optimum tuning frequency ratio γ'is calculated by the following equation (22) (step 26), and the attenuation constant of the variable attenuation damper 81 is calculated by the following equation (23). The target attenuation constant hobj', which is the target value, is calculated (step 27). Here, when the optimum tuning frequency ratio γ'greatly exceeds 1, the process returns to step 23, the neighborhood natural frequency fn in the equation (20) is replaced with γ'· fn, and the inertial mass Mds' is replaced again. It is desirable to calculate. As described above, when the optimum tuning frequency ratio γ'greatly exceeds 1, the neighborhood natural frequency fn in the present specification may be replaced with γ'· fn.

Next, using the calculated optimum tuning frequency ratio γ'and target attenuation constant hobj', and the near natural frequency fn and inertial mass Mds' set in steps 21 and 22, respectively, according to the following equation (24). , The target damping coefficient cobj', which is the target value of the damping coefficient of the variable damping damper 81, is calculated (step 28).
cobj'= 2hobj'・ Mds' ・ γ'・ fn ・ 2π …… (24)

The above equations (22) to (24) are similar to the equations (12) to (14) described in the first embodiment, and the maximum value of the response coefficient of the building B is minimized according to the fixed point theory. It is derived so that the target attenuation coefficient cobj'can be calculated.

As is clear from the calculation method of the target damping coefficient cobj'according to the above steps 8 and 24 to 28, the variables used for the calculation are the near natural frequency fn and the inertial mass Mds', so that both fn. , Mds'and the target attenuation coefficient cobj' may be obtained and mapped in advance, and the target attenuation coefficient cobj' may be calculated by searching this map according to fn and Mds'. Further, although the above equations (22) to (24) are used for the calculation of the target attenuation coefficient cobj', other appropriate ones derived according to the fixed point theory so that the maximum value of the response magnification of the building B is minimized. Formula may be used.

Next, the regulating valve 86 is driven by outputting a drive signal based on the calculated target attenuation coefficient cobj'(step 29), and the current process is completed. By driving the regulating valve 86 in this way, the damping coefficient of the variable damping damper 81 is controlled to be the target damping coefficient cobj'. The drive signal is obtained in advance by an experiment or the like so that the attenuation coefficient becomes the target attenuation coefficient cobj'and is stored in the ROM. Further, the tuning control process including steps 1, 2, 8 and 21 to 29 is repeatedly executed at predetermined time intervals as described above.

As described above, according to the second embodiment, as described with reference to FIGS. 12, 14 and 15, the variable rotation inertial mass damper 61 and the first and second elastic members EM1'and EM2' , The additional vibration system is configured. Displacement due to vibration of the building B is transmitted to the variable rotary inertial mass damper 61 via the first and second elastic members EM1'and EM2', and the transmitted displacement of the building B is the screw shaft 65 and the nut 66. The rotational power is converted into rotational power by a ball screw including, and the converted rotational power is changed by a stepped speed change mechanism 69 at one speed change ratio selected from predetermined first and second speed change ratios R1 and R2. Is transmitted to the rotating masses 68 and 68, whereby the rotating masses 68 and 68 rotate, and as a result, the additional vibration system vibrates. The inertial mass Md'of the variable rotational inertial mass damper 61 is changed in two stages by changing the gear ratio of the speed change mechanism 69.

Further, in the first and second gear ratios R1 and R2, the natural frequency of the additional vibration system obtained when the first and second gear ratios are selected during the vibration of the building B is the primary frequency of the building B. And the secondary natural frequencies f1 and f2 are set to be tuned to each other, respectively. Further, as described with reference to FIG. 16, the predominant frequency fcp (frequency of the predominant frequency component of the vibration input to the building B) is calculated (step 2), and the building B is vibrating and is unique to the vicinity. The inertial mass Md'so that the natural frequency of the added vibration system is tuned to the frequency fn (the natural frequency of the primary and secondary natural frequencies f1 and f2 of the building B that is closest to the predominant frequency fcp). Is controlled (steps 21 and 22). As a result, by synchronizing the natural frequency of the additional vibration system with the natural frequency fn in the vicinity of the building B, the vibration of the building B at the strongest dominant frequency fcp included in the occasional vibration input to the building B can be obtained. , Can be appropriately absorbed and suppressed by the additional vibration system.

Further, as described with reference to FIGS. 13 and 15, the variable damping damper 81 damps the vibration of the additional vibration system, and the damping coefficient is continuous by changing the opening degree of the adjusting valve 86. Is changed to. Further, as described with reference to FIG. 16, during the vibration of the building B, according to the near natural frequency fn and the inertial mass Md'(inertial mass Mds') of the controlled variable rotation inertial mass damper 61. By changing the opening degree of the regulating valve 86, the damping coefficient of the variable damping damper 81 is controlled so that the maximum value of the response magnification of the building B is minimized (steps 8 and 23 to 29). Vibration can be suppressed more appropriately.

Next, the variable rotation inertial mass damper 101 of the vibration suppression device according to the third embodiment of the present invention will be described with reference to FIG. Compared with the second embodiment described above, the variable rotary inertial mass damper 101 has a configuration of the rotary masses 102 and 102, and a friction material 103 is provided between each rotary mass 102 and the rotary shaft 70. , Different. In FIG. 17, the same components as those in the second embodiment are designated by the same reference numerals. Hereinafter, the points different from the second embodiment will be mainly described.

The rotary mass 102 is different from the rotary mass 68 of the second embodiment only in that a fitting hole 102a penetrating in the axial direction thereof is formed. The friction material 103 is made of a material having a relatively stable friction coefficient, for example, Teflon (registered trademark), is formed in an annular shape, and is attached to both ends of the rotating shaft 70. It is fitted in the fitting hole 102a of 102. By this fitting, the rotary mass 102 is connected to the rotary shaft 70. The friction coefficient of the friction material 103 is set so that the rotary mass 102 slides with respect to the friction material 103 when the rotational torque of the rotary mass 102 becomes very large due to the extremely large vibration of the building B. There is. The friction material 103 may be discontinuously attached to the peripheral surface of the rotating shaft 70 without being formed in an annular shape.

Further, flanges 70a are integrally provided on both sides of the rotating shaft 70 in the axial direction of the rotating mass 102, and the rotating mass 102 is sandwiched between the flanges 70a and 70a.

Although not shown, the variable rotation inertial mass damper 101 has the lower beam BD and the upper beam of the building B via the first and second elastic members EM1'and EM2', respectively, as in the case of the second embodiment. It is connected to the BU in parallel with the variable damping damper 81 (see FIG. 15). The variable rotary inertial mass damper 101 constitutes an additional vibration system together with the first and second connecting members EN1'and EN2'.

As described above, according to the third embodiment, the friction material 103 is attached to the rotating shaft 70, and the rotating mass 102 is formed by fitting the friction material 103 into the fitting hole 102a. It is connected to the rotating shaft 70 via 103. Further, since the friction coefficient of the friction material 103 is set as described above, when the rotational torque of the rotating mass 102 becomes very large due to the extremely large vibration of the building B, the rotating mass 102 becomes the friction material. It slides against 103, thereby limiting the inertial force due to the inertial mass generated by the rotation of the rotating masses 102, 102.

The inertial force due to the inertial mass generated by the rotation of the rotating masses 102 and 102 during the vibration of the building B may be limited by disengaging the clutch 67.

In the second and third embodiments, a single synchromesh mechanism 73 for connecting / disconnecting the second and fourth gears 71b and 72b to the rotating shaft 70 is used, but they are provided separately from each other. Synchromesh mechanisms for the second gear 71b and the fourth gear 72b may be used. Further, in the second and third embodiments, the first gear 71a is integrally provided with the outer cylinder 64, and the second gear 71b is rotatably provided on the rotating shaft 70. On the contrary, the first gear is provided. The 71a may be rotatably provided on the outer cylinder 64, and the second gear 71b may be integrally provided on the rotating shaft 70. This also applies to the third and fourth gears 72a and 72b.

Further, in the second and third embodiments, the number of gears of the speed change mechanism 69 is 2, but it may be 3 or more. In that case, a plurality of gear ratios of three or more steps correspond to each of these, and a plurality of natural frequencies of the additional vibration system obtained corresponding to each of them correspond to a plurality of natural frequencies of the building B in the third mode or higher. Is set to tune in. Further, in the second and third embodiments, the speed change mechanism 69 is a parallel shaft type stepped speed change mechanism, but other suitable stepped speed change mechanism, for example, a planetary gear device used in a vehicle or the like, or A stepped speed change mechanism composed of a combination of a brake and a clutch may be used.

Further, in the third embodiment, the friction material 103 is attached to the rotating shaft 70 and fitted into the fitting hole 102a of the rotating mass 102, but the friction material is provided on the inner peripheral surface of the fitting hole of the rotating mass 102. At the same time, the rotating shaft may be fitted inside the friction material. Alternatively, the friction material may be fitted between the rotating shaft and the inner peripheral surface of the fitting hole without being attached (fixed). In that case, the rotating mass is formed with a plurality of accommodating holes penetrating in the radial direction and communicating with the fitting holes, and each accommodating hole is provided with a spring that contacts the friction material and is outward in the radial direction of the rotating mass. The degree of compression of the spring may be changed by screwing a screw into the accommodating hole, thereby adjusting the degree of fitting (friction coefficient) of the friction material with respect to the rotating shaft and the inner peripheral surface of the fitting hole.

Next, the variable rotation inertial mass damper 111 of the vibration suppression device according to the fourth embodiment of the present invention will be described with reference to FIGS. 18 to 21. The variable rotation inertial mass damper 111 is different from the second embodiment in that the outer cylinder 64 and the clutch 67 are not provided and the configuration of the transmission mechanism 112 is different. In FIGS. 18 to 20, the same components as those in the first and second embodiments are designated by the same reference numerals. Hereinafter, the points different from those of the first and second embodiments will be mainly described.

The speed change mechanism 112 of the variable rotation inertial mass damper 111 shown in FIG. 18 is a metal belt type stepless speed change mechanism in which the gear ratio can be changed steplessly, and is well known as used in a vehicle or the like. Hereinafter, the configuration and operation thereof will be briefly described. The speed change mechanism 112 includes a drive pulley 113, a driven pulley 114, a transmission belt 115, a first solenoid valve 116, a second solenoid valve 117 (see FIG. 19), and the like, and the drive pulley 113 is a cone facing each other. It has a trapezoidal fixed portion 113a and a movable portion 113b.

The fixed portion 113a is fixed to the peripheral wall 63a of the inner cylinder 63, and the movable portion 113b is provided on the peripheral wall 63a of the inner cylinder 63 so as to be movable in the axial direction and relatively non-rotatable. Further, a V-shaped belt groove for winding the transmission belt 115 is formed between the fixed portion 113a and the movable portion 113b. A hydraulic pump (not shown) is connected to the movable portion 113b, and the oil pressure supplied from the hydraulic pump to the movable portion 113b is adjusted by changing the opening degree of the first solenoid valve 116. As a result, the effective diameter of the drive pulley 113 changes steplessly by changing the pulley width of the drive pulley 113. As shown in FIG. 19, the first solenoid valve 116 is connected to the power supply 122 via a control device 121 described later.

The driven pulley 114 is configured in the same manner as the drive pulley 113, and has a truncated cone-shaped fixed portion 114a and a movable portion 114b that face each other. The fixed portion 114a is fixed to the rotating shaft 70, and the movable portion 114b is provided on the rotating shaft 70 so as to be movable and non-rotatable in the axial direction thereof. Further, a V-shaped belt groove is formed between the fixed portion 114a and the movable portion 114b. A hydraulic pump (not shown) is connected to the movable portion 114b, and the oil pressure supplied from the hydraulic pump to the movable portion 114b is adjusted by changing the opening degree of the second solenoid valve 117. As a result, the effective diameter of the driven pulley 114 changes steplessly by changing the pulley width of the driven pulley 114. As shown in FIG. 19, the second solenoid valve 117 is connected to the power supply 122 via the control device 121.

The transmission belt 115 is formed by connecting a large number of elements made of metal plates with a band-shaped metal ring in a state of being superposed on each other, and is wound around the belt grooves of the drive pulley 113 and the driven pulley 114.

In the speed change mechanism 112 having the above configuration, the effective diameters of the drive pulley 113 and the driven pulley 114 change steplessly by controlling the opening degrees of the first and second solenoid valves 116 and 117 by the control device 121. , The gear ratio is controlled steplessly. The gear ratio of the speed change mechanism 112 can be changed steplessly between the first upper limit gear ratio and the second lower limit gear ratio, both of which are smaller than the value 1.0.

Further, as shown in FIG. 20, the variable rotation inertial mass damper 111 is provided with the upper beam BU of the building B via the first and second elastic members EM1'and EM2', respectively, as in the case of the second embodiment. And the lower beam BD are connected in parallel with the variable damping damper 81. The variable rotary inertial mass damper 111 constitutes an additional vibration system together with the first and second connecting members EN1'and EN2'. Further, the variable rotation inertial mass damper 111, the variable damping damper 81, and the first and second elastic members EM1'and EM2' are provided in one set for each of all the layers of the building B, for example, as in the second embodiment. And FIG. 20 shows one set of them. In FIG. 20, for convenience, some components and reference numerals are omitted. By the way, it goes without saying that the variable rotational inertial mass damper 111, the variable damping damper 81, and the first and second elastic members EM1'and EM2' may be provided not only in all the layers of the building B but only in some layers. Is.

Next, the operation of the variable rotation inertial mass damper 111 will be described. When a relative displacement occurs between the upper and lower beams BU and BD due to the vibration of the building B, the relative displacement is transmitted to the main body portion 62 and the screw shaft 65, so that the screw shaft 65 moves with respect to the main body portion 62. Move in the axial direction. The movement of the screw shaft 65 is converted into a rotary motion by the nut 66, and the nut 66 rotates with respect to the main body portion 62 together with the inner cylinder 63. The rotation of the inner cylinder 63 is transmitted to the rotating shaft 70 in a state where the speed is changed by the speed change mechanism 112, and is further transmitted to the rotating masses 68 and 68. As a result, as a result of the rotation of the rotating masses 68 and 68, an inertial force due to the inertial mass Md'due to the rotation of the rotating masses 68 and 68 is generated.

In this case, unlike the transmission mechanism 69 of the second embodiment, the transmission mechanism 112 can change the gear ratio steplessly. Therefore, the variable rotation inertial mass damper 111 changes the gear ratio of the transmission mechanism 112. The inertial mass Md'can be changed continuously.

The control device 121 and the power supply 122 are configured in the same manner as the control device 51 and the power supply 52 of the first embodiment, and a measurement signal representing the seismic motion measured by the seismograph 53 is input. As shown in FIG. 19, the control device 121 further receives a detection signal representing the rotation speed of the drive pulley 113 from the first rotation speed sensor 123 and a detection signal representing the rotation speed of the driven pulley 114 from the second rotation speed sensor 124. Is entered. The control device 121 calculates the gear ratio (rotation speed of the drive pulley 113 / rotation speed of the driven pulley 114) of the transmission mechanism 112 based on the detection signals input from the first and second rotation speed sensors 123 and 124. ..

Further, the control device 121 sets the gear ratio of the speed change mechanism 112 and controls the opening degree of the adjusting valve 86 in order to control the inertial mass Md'of the variable rotation inertial mass damper 111 and the damping coefficient of the variable damping damper 81. Therefore, the tuning control process shown in FIG. 21 is repeatedly executed at predetermined time intervals. In FIG. 21, the same step numbers are assigned to the parts having the same execution contents as the tuning control processes (FIGS. 8 and 9, 9 and 16) of the first and second embodiments.

As is clear from the comparison between FIGS. 21 and 16, the difference is that the steps 3 and 31 are executed instead of the steps 21 and 22, respectively, as compared with the second embodiment, and the neighborhood is unique. The frequency fn is set by step 3 in the same manner as in the case of the first embodiment, and the gear ratio of the speed change mechanism 112 is set by the execution of step 31 by a method different from that of the second embodiment. Therefore, only the execution contents of this step 31 will be described below.

In step 31, the gear ratio of the transmission mechanism 112 is set (controlled) based on the set near natural frequency fn. Specifically, first, the target gear ratio Rcmd is calculated by the following equation (25) based on the equations (18) and (19) based on the near natural frequency fn. Next, the opening degrees of the first and second solenoid valves 116 and 117 are controlled by outputting a drive signal based on the target gear ratio Rcmd. As a result, the gear ratio is controlled so as to be the calculated target gear ratio Rcmd.
Rcmd = {(2π ²・ fn ・ D') / Ld} ・ sqrt {md'/ (2θs')}
…… (25)

The diameter D'of the rotating mass 68, the pitch Ld of the screw shaft 65, the actual amount md'of the rotating masses 68 and 68, and the overall rigidity θs' of the first and second elastic members EM1'and EM2' in this equation (25) are All of them are predetermined values, are stored in the ROM of the control device 121, and are read from the ROM in step 31. Further, the drive signal is obtained in advance by an experiment or the like so that the gear ratio becomes the target gear ratio Rcmd, and is stored in the ROM.

As is clear from the equations (18), (19) and (25), by setting the gear ratio as described above, the natural frequency of the additional vibration system becomes the same as the near natural frequency fn. The inertial mass Md'of the variable rotation inertial mass damper 111 is controlled.

Of course, with respect to the fourth embodiment, a variation of the calculation method of the target attenuation coefficient cobj'described in the second embodiment may be adopted. That is, the relationship between the near natural frequency fn and the inertial mass Mds'and the target damping coefficient cobj'is obtained and mapped in advance, and this map is searched according to fn and Mds' to obtain the target damping coefficient cobj. 'You may calculate. Further, although the above equations (22) to (24) are used for the calculation of the target attenuation coefficient cobj', other appropriate ones derived according to the fixed point theory so that the maximum value of the response magnification of the building B is minimized. Formula may be used.

As described above, according to the fourth embodiment, as described with reference to FIGS. 18 to 20, additional vibration is added by the variable rotation inertial mass damper 111 and the first and second elastic members EM1'and EM2'. The system is composed. Displacement due to vibration of the building B is transmitted to the variable rotary inertial mass damper 111 via the first and second elastic members EM1'and EM2', and the transmitted displacement of the building B is the screw shaft 65 and the nut 66. The rotational power is converted into rotational power by a ball screw including, and the converted rotational power is transmitted to the rotary masses 68 and 68 in a state of being shifted by the stepless speed change mechanism 112, whereby the rotary masses 68 and 68 rotate. , The additional vibration system vibrates. The inertial mass Md'of the variable rotational inertial mass damper 111 is continuously changed by changing the gear ratio of the speed change mechanism 112.

Further, as described with reference to FIG. 21, the dominant frequency fcp (frequency of the dominant frequency component of the vibration input to the building B) is calculated (step 2), and the building B is vibrating during the vibration of the building B. The speed change of the speed change mechanism 112 so that the natural frequency of the additional vibration system is synchronized with the natural frequency fn (the natural frequency closest to the dominant frequency fcp among a plurality of predetermined natural frequencies of the building B). By setting the ratio, the inertial mass Md'is controlled (steps 3 and 31). As a result, by synchronizing the natural frequency of the additional vibration system with the natural frequency fn in the vicinity of the building B, the vibration of the building B at the strongest dominant frequency fcp included in the occasional vibration input to the building B can be obtained. , Can be appropriately absorbed and suppressed by the additional vibration system.

Further, as in the second embodiment, the variable damping damper 81 damps the vibration of the additional vibration system, and the damping coefficient thereof is continuously changed by changing the opening degree of the adjusting valve 86. Further, during the vibration of the building B, the opening degree of the adjusting valve 86 is changed according to the proximity natural frequency fn and the inertial mass Md'(inertial mass Mds') of the controlled variable rotary inertial mass damper 111. Since the damping coefficient of the variable damping damper 81 is controlled so that the maximum value of the response magnification of the building B is minimized (steps 8, 24 to 29), the vibration of the building B can be suppressed more appropriately.

In the fourth embodiment, the outer cylinder 64 and the clutch 67 of the second and third embodiments are omitted, but these may be adopted. In that case, the fixed portion 113a and the movable portion of the drive pulley 113 may be adopted. The 113b is provided on the outer cylinder 64. Further, with respect to the fourth embodiment, the rotary masses 68 and 68 and the rotary shaft 70 may be configured as described in the third embodiment, that is, each rotary mass may be connected to the rotary shaft via a friction material. .. In this case as well, it goes without saying that the above-mentioned variations may be adopted for the rotating mass and the friction material.

Further, in the fourth embodiment, the speed change mechanism 112 is a metal belt type continuously variable transmission mechanism, but is another suitable continuously variable transmission mechanism, for example, a traction drive type (toroidal type) used in a vehicle or the like. A continuously variable transmission mechanism or the like may be used. Further, in the fourth embodiment, the inertial mass Md'is controlled so that the natural frequency of the additional vibration system is the same as the near natural frequency fn, but the natural frequency of the additional vibration system is near-specific. It may be controlled so as to be substantially the same as the frequency fn. In that case, instead of the neighborhood natural frequency fn in the above equation (25), a value obtained by adding or subtracting a very small predetermined value to fn is used.

Further, in the second to fourth embodiments, the ball screw including the screw shaft 65 and the nut 66 is used as the conversion mechanism in the present invention, but another suitable method for converting the transmitted displacement of the structure into rotational power. For example, a mechanism composed of racks and pinions that mesh with each other, or a mechanism composed of the cylinder 2, the piston 3, the first communication passage 5, the gear motor M, and the like of the first embodiment may be used. In this case, of course, the screw mechanism of Japanese Patent No. 5191579 described above may be used instead of the gear motor M.

Further, in the second to fourth embodiments, the variable damping damper 81 is of a type using a viscous fluid HF composed of silicone oil, but may be of a type using a hydraulic oil or an MR fluid. Further, in the second to fourth embodiments, the regulating valve 86 composed of the solenoid valve is used, but a regulating valve of the type driven by hydraulic pressure or pneumatic pressure may be used. Further, in the second to fourth embodiments, the variable damping damper 81 is provided, but this is omitted and the processing for controlling the damping coefficient in the tuning control processing (steps 8 and 24 to 29) is omitted. You may. Further, in the second to fourth embodiments, the piston 83 is provided with the first and second relief valves 87 and 88, but these may be omitted.

Next, the vibration suppression device according to the fifth embodiment of the present invention will be described with reference to FIGS. 22 to 25. This vibration suppression device is applied to the building B (see FIG. 24) as in the first embodiment, and includes the variable rotation inertial mass damper 131 shown in FIG. 22, the control device 141 shown in FIG. 23, and the power supply 142. The seismograph 53 and the first and second elastic members EM1 and EM2 shown in FIG. 24 are provided. In FIGS. 22 and 24, the same components as those in the first embodiment are designated by the same reference numerals. Hereinafter, the points different from those of the first embodiment will be mainly described.

As is clear from the comparison between FIGS. 22 and 1, the variable rotation inertial mass damper 131 does not include the second communication passage 6 and the gear pump P as compared with the variable rotation inertial mass damper 1 of the first embodiment. The difference is that the variable displacement type fluid pressure motor 132 is provided instead of the gear motor M.

Since the fluid pressure motor 132 is, for example, a well-known swash plate type variable displacement hydraulic motor, its configuration and operation will be briefly described below. The fluid pressure motor 132 has a rotating shaft 132a, an actuator 132b (see FIG. 23), a swash plate, a cylinder block, a piston, a shoe (none of which are shown), and is provided in the middle of the first continuous passage 5. ing. The cylinder block is provided with an inflow port and an outflow port, and these inflow ports and outflow ports are connected to the first and second fluid chambers via the first communication passage 5 and a check valve (not shown). They are connected to 2d and 2e in parallel with each other.

In the fluid pressure motor 132, when the viscous fluid HF flows into the inflow port, the pressure energy of the inflowing viscous fluid HF is converted into the rotational energy of the rotating shaft 132a by the cylinder block, the piston, the shoe, and the swash plate. The rotating shaft 132a rotates. The tilt angle of the swash plate is configured to be continuously changed by the actuator 132b, and by changing the tilt angle of the swash plate, the push-out volume VM of the fluid pressure motor 132 (per rotation of the rotating shaft 132a). The amount of rotation of the rotating shaft 132a with respect to the pressure of the same viscous fluid HF is continuously changed by continuously changing the geometric volume).

The actuator 132b is, for example, an electromagnetic type, and is connected to the control device 141 shown in FIG. 23. Like the control device 51, the control device 141 is composed of a combination of a CPU, a RAM, a ROM, an I / O interface, and the like, is connected to the power supply 142, and tilts the swash plate via the actuator 132b. By changing the angle, the push-out volume VM of the fluid pressure motor 132 is adjusted. Of course, a fluid pressure type actuator or the like may be used as the actuator 132b.

Further, the rotating shaft 132a is integrally provided with the rotating mass 21 coaxially. When the pressure energy of the viscous fluid HF is converted into the rotational energy of the rotating shaft 132a by the fluid pressure motor 132, the rotating mass 21 rotates integrally with the rotating shaft 132a.

As shown in FIG. 24, in the variable rotation inertial mass damper 131 having the above configuration, the cylinder 2 and the piston rod 4 are the first and second elastic members EM1 as in the variable rotation inertial mass damper 1 of the first embodiment. , EM2 is connected to the lower beam BD and the upper beam BU, respectively, and extends horizontally between both BDs and BUs to form an additional vibration system together with the first and second elastic members EM1 and EM2. There is. Further, the variable rotation inertial mass damper 131, the first and second elastic members EM1 and EM2 are provided in one set for each of all the layers of the building B, and FIG. 24 shows one set thereof.

In FIG. 24, for convenience, some components such as the first passage 5 are not shown. By the way, it goes without saying that the variable rotary inertial mass damper 131 and the first and second elastic members EM1 and EM2 may be provided not only in all the layers of the building B but only in a part of the layers.

Next, the operation of the variable rotation inertial mass damper 131 will be described. When a relative displacement in the horizontal direction occurs between the upper and lower beams BU and BD as the building B vibrates, the relative displacement is caused by the cylinder 2 and the piston via the first and second elastic members EM1 and EM2. By being transmitted to the rod 4 as an external force, the cylinder 2 and the piston rod 4 move relatively in the axial direction, and the piston 3 slides in the cylinder 2.

In this case, when the piston 3 moves to the first fluid chamber 2d side (left side in FIG. 22), a part of the viscous fluid HF in the first fluid chamber 2d is pushed out into the first continuous passage 5 by the piston 3. As a result, the viscous fluid HF flows to the second fluid chamber 2e side (right side) in the first communication passage 5. On the contrary, when the piston 3 moves to the second fluid chamber 2e side (right side), a part of the viscous fluid HF in the second fluid chamber 2e is pushed out by the piston 3 into the first continuous passage 5. As a result, the viscous fluid HF flows to the first fluid chamber 2d side (left side) in the first communication passage 5.

The pressure energy of the viscous fluid HF due to this flow is converted into the rotational energy of the rotating shaft 132a by the fluid pressure motor 132, and the rotating shaft 132a rotates together with the rotating mass 21. With the flow of the viscous fluid HF and the rotation of the rotating mass 21, the inertial mass by the rotating mass 21 (inertial mass based on the rotating inertial mass of the rotating mass 21) MR and the first series are the same as in the case of the first embodiment. Inertial mass MD (inertial mass in the axial direction with respect to external force due to vibration input to the cylinder 2 and piston 3) including inertial mass Mh due to the viscous fluid HF in the passage 5 is generated.

In this case, the inertial mass MR due to the rotating mass 21 is represented by the following equation (26), and the inertial mass Mh due to the viscous fluid HF is as described in the first embodiment. Further, the inertial mass Mh due to the viscous fluid HF tends to be very small as compared with the inertial mass MR due to the rotating mass 21.
^{MR = (2π / XM) 2} · md · D 2/8
= {(2π · Ap) / VM} 2 · md · D 2/8 ...... (26)

In the formula (26), XM is the amount of movement of the piston 3 with respect to the cylinder 2 required for one rotation of the rotating shaft 132a of the fluid pressure motor 132 due to the flow of the viscous fluid HF, and is the rotational inertia using the ball screw mechanism. It corresponds to the lead length Ld of the ball screw in the mass damper. Further, the VM is the push-out volume of the fluid pressure motor 132 as described above, and the other parameters (md, D, Ap) are as described in the first embodiment.

As is clear from the above equation (26), the inertial mass MD of the variable rotary inertial mass damper 131 including the inertial mass MR by the rotary mass 21 is adjusted by adjusting the push-out volume VM of the fluid pressure motor 132 by the control device 141. Is controlled. In this case, the inertial mass MD is continuously changed by adjusting the push-out volume VM, and the smaller the push-out volume VM, the more the rotation amount of the rotating shaft 132a and the rotating mass 21 with respect to the pressure energy of the same viscous fluid HF. As it becomes larger, the inertial mass MD becomes larger. This is also clear from equation (26).

Further, as in the case of the first embodiment, the cylinder 2, the piston 3, the first communication passage 5, the viscous fluid HF and the adjusting valve 15 have the variable rotation inertial mass damper 131, the first and second elastic members EM1 and EM2. It functions as a variable damping damper that damps the vibration of the additional vibration system including it and can continuously change its damping coefficient. In this case, the damping coefficient of the variable damping damper is continuously changed and damped by adjusting the flow resistance of the viscous fluid HF flowing in the first continuous passage 5 by changing the opening degree of the adjusting valve 15. The coefficient becomes larger as the opening degree of the regulating valve 15 becomes smaller, because the pressure difference of the viscous fluid HF between the first and second fluid chambers 2d and 2e becomes larger.

Further, the control device 141 is shown in FIG. 25 in order to control the push-out volume VM of the fluid pressure motor 132 and the opening degree of the adjusting valve 15 in order to control the inertial mass MD and the damping coefficient of the variable rotary inertial mass damper 131. The tuning control process is repeatedly executed at predetermined time intervals. In FIG. 25, the same step numbers are assigned to the parts having the same execution contents as those of the tuning control process (FIGS. 8 and 9) of the first embodiment.

As is clear from the comparison between FIG. 25 and FIGS. 8 and 9, only steps 41 and 42 are executed instead of the steps 5 to 7 as compared with the first embodiment. Therefore, only the execution contents of these steps 41 and 42 will be described below. As described above, the inertial mass Mh due to the viscous fluid HF contained in the inertial mass MD of the variable rotary inertial mass damper 131 tends to be very small as compared with the inertial mass MR due to the rotating mass 21. Therefore, in the tuning control process, this inertial mass Mh is ignored, the inertial mass MR by the rotating mass 21 is regarded as the inertial mass MD of the variable rotating inertial mass damper 131, and the push-out volume VM of the fluid pressure motor 132 and the adjusting valve 15 The opening degree is controlled.

In step 41, the target push-out volume VMobj, which is the target value of the push-out volume VM, is calculated by the following equation (27) using the inertial mass Mds for this control calculated in step 4. In this equation (27), the above equation (26) is developed for the push-out volume VM, the push-out volume VM is replaced with the target push-out volume VMobj, and the inertial mass MR by the rotating mass 21 is replaced with the inertial mass Mds. be.
VMobj = (2π ・ Ap) / sqrt {8Mds / (md ・ D ² )} …… (27)

In step 42 following step 41, the actuator 132b is driven by outputting a drive signal based on the calculated target push-out volume VMobj to the actuator 132b. As a result, the push-out volume VM of the fluid pressure motor 132 is controlled to be the target push-out volume VMobj. As a result, the inertial mass MR by the rotating mass 21 regarded as the inertial mass MD of the variable rotary inertial mass damper 131 is controlled by the inertial mass Mds calculated in step 4. As a result, the natural frequency (= sqrt (θs / MD) / 2π) of the additional vibration system determined by the inertial mass MD and the rigidity θs of the first and second elastic members EM1 and EM2 becomes the same as the neighborhood natural frequency fn. Is controlled to be. The drive signal is obtained in advance by an experiment or the like so that the target push-out volume VMobj can be obtained, and is stored in the ROM.

Further, in step 8 and subsequent steps following step 42, the opening degree of the adjusting valve 15 is controlled in the same manner as in the case of the first embodiment, whereby a variable damping damper (variable rotational inertia) including the cylinder 2 and the piston 3 is controlled. The damping coefficient of the mass damper 131) is controlled to be the target damping coefficient cobj.

As described above, according to the fifth embodiment, the additional vibration system is composed of the variable rotation inertial mass damper 131 and the first and second elastic members EM1 and EM2. Further, the relative displacement between the upper and lower beams BU and BD due to the vibration of the building B is transmitted to the cylinder 2 and the piston 3 via the first and second elastic members EM1 and EM2, whereby the piston 3 becomes the cylinder 2 When it slides inside and moves to one side of the first and second fluid chambers 2d and 2e, the viscous fluid HF in the one fluid chamber is pushed out by the piston 3 into the first continuous passage 5, so that the first A flow of the viscous fluid HF to the other fluid chamber side occurs in the communication passage 5.

The pressure energy of the viscous fluid HF is converted into rotational energy by the variable displacement fluid pressure motor 132, and the converted rotational energy is transmitted to the rotary mass 21 to rotate the rotary mass 21. Along with this, an inertial mass MD corresponding to the rotational inertial mass of the rotating mass 21 is generated. In this case, by changing the push-out volume VM of the variable displacement fluid pressure motor 132, the amount of rotation of the rotating mass 21 with respect to the pressure energy of the same viscous fluid HF can be continuously changed, thereby rotating 21. The inertial mass MD of the variable rotary inertial mass damper 131 generated with the rotation of the mass can be continuously changed.

Further, as in the case of the first embodiment, the predominant frequency fcp, which is the frequency of the predominant frequency component of the vibration input to the building B, is calculated (step 2 in FIG. 25). Further, during the vibration of the building B, the natural frequency of the additional vibration system is added to the near natural frequency fn, which is the natural frequency closest to the calculated dominant frequency fcp, among a plurality of predetermined natural frequencies of the building B. More specifically, the variable rotational inertia mass is adjusted by adjusting the push-out volume VM via the actuator 132b so that the natural frequency of the additional vibration system becomes the same as the near natural frequency fn. The inertial mass MD of the damper 131 is controlled (steps 34-42). As a result, by synchronizing the natural frequency of the additional vibration system with the natural frequency fn in the vicinity of the building B, the vibration of the building B at the strongest dominant frequency fcp included in the occasional vibration input to the building B can be obtained. , Can be appropriately absorbed and suppressed by the additional vibration system.

Further, as in the case of the first embodiment, the vibration of the additional vibration system is damped by the variable damping damper including the cylinder 2, the piston 3, the first communication passage 5, the viscous fluid HF, and the adjusting valve 15, and the adjusting valve 15 By changing the opening degree of, the damping coefficient of the variable damping damper is continuously changed. Further, during the vibration of the building B, the response of the building B according to the near natural frequency fn and the inertial mass MD (inertial mass Mds) of the variable rotation inertial mass damper 131 controlled as described above according to the fixed point theory. Since the damping coefficient of the variable damping damper is controlled so that the maximum value of the magnification is minimized (steps 8 to 14), the vibration of the building B can be suppressed more appropriately. Further, the variable rotation inertial mass damper 131 also has the function of the variable damping damper, and the cylinder 2, the piston 3, the first communication passage 5, and the viscous fluid HF are the components of the variable rotation inertial mass damper and the variable damping damper. Since it is also used as a vibration suppressor, the vibration suppression device can be miniaturized accordingly. In addition, the above-mentioned effects according to the first embodiment can be obtained in the same manner.

In the fifth embodiment, the fluid pressure motor 132 is a swash plate type, but other suitable variable displacement type fluid pressure motors, for example, a sloping shaft type or a vane type, may be used. Further, although the fluid pressure motor 132 is configured so that the push-out volume VM can be continuously changed, it may be configured so that the push-out volume VM can be changed stepwise. Further, in the fifth embodiment, the regulating valve 15 composed of the solenoid valve is used, but a regulating valve of a type driven by hydraulic control or pneumatic pressure may be used. Further, in the fifth embodiment, the piston 3 is provided with the first and second relief valves 11 and 12, but these may be omitted.

Further, in the fifth embodiment, the variable damping damper in the present invention is composed of the cylinder 2, the piston 3, the first communication passage 5, the viscous fluid HF, the adjusting valve 15, and the like, and the variable rotating inertial mass damper 131 is the variable damping damper. The variable rotation inertial mass damper has a configuration in which the adjusting valve 15 is removed from the variable rotation inertial mass damper 131, and the damping coefficient is changed while attenuating the vibration of the additional vibration system. Possible variable damping dampers may be provided in parallel. Of course, as the variable damping damper in this case, various variable damping dampers such as the variable damping damper 81 and a variable damping damper of the type using MR fluid (Magneto-Rheological fluid) may be used. Alternatively, the variable damping damper (adjusting valve 15) may be omitted, and the processing for controlling the damping coefficient in the tuning control processing (steps 8 to 14) may be omitted.

Further, in the fifth embodiment, the inertial mass MD of the variable rotation inertial mass damper 131 is controlled so that the natural frequency of the additional vibration system is the same as the near natural frequency fn. It may be controlled so that the natural frequency becomes substantially the same as the nearby natural frequency fn. In that case, instead of the neighborhood natural frequency fn in the above equation (9), a value obtained by adding or subtracting a very small predetermined value to fn is used.

Further, in the fifth embodiment (synchronization control process), the inertial mass Mh due to the viscous fluid HF is ignored, the inertial mass MR due to the rotating mass 21 is regarded as the inertial mass MD of the variable rotating inertial mass damper 131, and the push-out volume VM and Although the opening degree of the adjusting valve 15 is controlled, the control may be performed with MD = MR + Mh without ignoring the inertial mass Mh. In this case, in the processing after step 4 of FIG. 25, the value obtained by subtracting the inertial mass Mh from the inertial mass Mds calculated by the formula (9) is used as the inertial mass Mds. Furthermore, it goes without saying that the variations described above with respect to the fifth embodiment may be appropriately combined and applied.

The present invention is not limited to the first to fifth embodiments described above (hereinafter, collectively referred to as "embodiments"), and can be implemented in various embodiments. For example, in the first and fifth embodiments, the first and second elastic members EM1 and EM2 made of steel are used, but other suitable elastic members such as springs and rubber may be used. .. Further, in the first embodiment, the first and second elastic members EM1 and EM2 are provided so as to extend in the vertical direction, but the former EM1 may be provided in an inverted V-shaped brace shape, and the latter EM2. May be provided in a V-shaped brace shape. These things also apply to the first and second elastic members EM1'and EM2' of the second to fourth embodiments. Further, in the embodiment, the viscous fluid HF composed of silicone oil is used, but a hydraulic oil or another suitable fluid having viscosity may be used.

Further, the variable rotary inertial mass dampers 1, 61, 101, 111, and 131 of the embodiment are merely examples, and other appropriate variable rotary inertial mass dampers capable of changing the inertial mass may be used. Of course. For example, a plurality of rotating masses that are coaxially arranged in the axial direction with each other and a displacement caused by vibration of a structure transmitted via an elastic member are converted into rotary motion, and the axial lines of the plurality of rotating masses are converted. A conversion mechanism that transmits to one rotating mass located at the end of the direction, and multiple clutches that connect / disconnect between two rotating masses that are adjacent to each other among multiple rotating masses (number of clutches = number of rotating masses) A variable rotary inertial mass damper having the number -1) may be used.

In the variable rotary inertial mass damper having the above configuration, the inertial mass of the entire rotary mass acting on the conversion mechanism is increased by connecting the two adjacent rotary masses with a clutch in order from the conversion mechanism side to the opposite side. , Gradually larger. In this case, for example, when the number of rotating squares is three, the actual amount and diameter of each rotating square are set as follows. That is, when the two sets of rotating masses are connected by a clutch (maximum inertial mass), the natural frequency of the additional vibration system is synchronized with the natural frequency of the primary mode of the structure, and the conversion mechanism When the set of rotating masses on the opposite side is cut off and the remaining set of rotating masses are connected (medium inertial mass), the natural frequency of the additional vibration system is the structure. When the natural frequency of the secondary mode of the above is synchronized and the space between the two sets of rotating masses is cut off by the clutch (minimum inertial mass), the natural frequency of the additional vibration system is the normal mode of the structure. The actual amount and diameter of each rotating mass are set so as to be synchronized with the natural frequency of.

Further, when the variable rotary inertial mass damper having the above configuration is used, the inertial mass is controlled so that the natural frequency of the additional vibration system is synchronized with the natural frequency in the vicinity by controlling the connection / disconnection of the clutch. Will be done. Of course, as the conversion mechanism in this case, various conversion mechanisms as described in the embodiment and the variations described above may be used. Alternatively, the variable rotational inertia mass damper disclosed in Japanese Patent Application Laid-Open No. 2016-151287 may be used.

Further, the control method of the tuning control process described in the embodiment is merely an example, and it goes without saying that another appropriate control method may be adopted within the scope of the gist of the present invention. Further, in the first and fifth embodiments, the variable rotation inertial mass dampers 1, 131 are used, and in the second to fourth embodiments, the variable rotation inertial mass dampers 61, 101, 111 and the variable damping damper 81 are used in the upper beam BU and the variable damping damper 81. Although connected to the lower beam BD, it may be connected to two other suitable predetermined parts of the system, including the foundation on which the structure is erected and the structure, such as the foundation and the upper end of the structure. Alternatively, it may be connected to the structure and the foundation. Further, in the embodiment, the structure in the present invention is the building B, but other suitable structures such as a bridge, a steel tower, a rack warehouse, and the like may be used. In addition, within the scope of the gist of the present invention, the detailed configuration can be changed as appropriate.

B building (structure)
1 Variable rotation inertial mass damper (variable damping damper)
2 Cylinder 2d 1st fluid chamber 2e 2nd fluid chamber 3 Piston 5 1st passage 6 2nd passage HF Viscous fluid (working fluid)
11 1st relief valve 12 2nd relief valve 15 Adjusting valve M Gear motor (flow conversion mechanism)
P gear pump (flow rate adjustment mechanism, electric pump)
21 Rotating mass EM1 1st elastic member (elastic member)
EM2 2nd elastic member (elastic member)
51 Control device (excellent frequency acquisition means, first control means, second control means)
53 Seismometer (excellent frequency acquisition means)
fcp dominant frequency fn near natural frequency Mds inertial mass (controlled variable rotational inertial mass damper inertial mass)
61 Variable rotation inertial mass damper 65 Screw shaft (conversion mechanism)
66 nut (conversion mechanism)
68 Rotating mass 69 Transmission mechanism 70 Rotating shaft 81 Variable damping damper 82 Cylinder 82d First fluid chamber (two viscous fluid chambers)
82e Second fluid chamber (two viscous fluid chambers)
83 Piston 85 Continuous passage 86 Adjusting valve EM1'First elastic member (elastic member)
EM2'Second elastic member (elastic member)
91 Control device (excellent frequency acquisition means, first control means, second control means)
Mds' Inertial Mass (Controlled Variable Rotational Inertial Mass Damper Inertial Mass)
101 Variable rotation inertial mass damper 102 Rotation mass 102a Fitting hole 103 Friction material 111 Variable rotation inertial mass damper 112 Speed change mechanism 121 Control device (excellent frequency acquisition means, first control means, second control means)
131 Variable rotation inertial mass damper (variable damping damper)
132 Fluid pressure motor 132b Actuator 141 Control device (excellent frequency acquisition means, first control means, second control means)

Claims (11)

It is a structure vibration suppression device for suppressing the vibration of the structure erected on the foundation.
An elastic member connected to the structure and
Together constitute an additional vibration system with the elastic member, and converts the displacement of the structure that is transmitted through the elastic member with the vibration of the structure to rotary motion of the rotary mass, the rotational mass rotates A variable rotary inertial mass damper configured to change the inertial mass generated as a result of the operation,
A predominant frequency acquisition means for acquiring a predominant frequency, which is a frequency of a predominant frequency component of vibrations input to the structure,
During the vibration of the structure, the natural frequency of the additional vibration system is added to the near natural frequency, which is the natural frequency closest to the acquired dominant frequency among a plurality of predetermined natural frequencies of the structure. The first control means for controlling the inertial mass of the variable rotation inertial mass damper so that the two are synchronized with each other.
A vibration suppression device for a structure, which comprises. The variable rotary inertial mass damper is
A cylinder in which displacement due to vibration of the structure is transmitted via the elastic member, and
The inside of the cylinder is divided into a first fluid chamber and a second fluid chamber filled with a working fluid, and the displacement due to the vibration of the structure is transmitted through the elastic member, so that the inside of the cylinder is aligned with the axis. With a piston configured to slide in the direction,
A first communication passage filled with working fluid while bypassing the piston and communicating with the first and second fluid chambers.
A flow conversion mechanism provided in the first passage and converting the flow of the working fluid in the first passage into the rotational motion of the rotating mass.
A second communication passage that is provided in parallel with the first communication passage, bypasses the piston, communicates with the first and second fluid chambers, and is filled with a working fluid.
It has a flow rate adjusting mechanism provided in the second passage and for adjusting the flow rate of the working fluid flowing in the second passage.
The first control means adjusts the flow rate of the working fluid flowing in the second communication passage through the flow rate adjusting mechanism during the vibration of the structure, so that the natural frequency of the additional vibration system is adjusted. The vibration suppression device for a structure according to claim 1, wherein the inertial mass of the variable rotation inertial mass damper is controlled so as to be synchronized with the natural frequency in the vicinity of the structure. The vibration suppression device for a structure according to claim 2, wherein the flow rate adjusting mechanism includes an electric pump for delivering a working fluid in the second continuous passage. The piston has a first relief valve that opens when the pressure of the working fluid in the first fluid chamber reaches a first predetermined pressure and communicates the first and second fluid chambers with each other, and the second. A second relief valve is provided, which opens when the pressure of the working fluid in the fluid chamber reaches a second predetermined pressure and communicates the second and first fluid chambers with each other. The structure vibration suppression device according to 2 or 3. The variable rotary inertial mass damper is
A cylinder in which displacement due to vibration of the structure is transmitted via the elastic member, and
The inside of the cylinder is divided into a first fluid chamber and a second fluid chamber filled with a working fluid, and the displacement due to the vibration of the structure is transmitted through the elastic member, so that the inside of the cylinder is aligned with the axis. With a piston configured to slide in the direction,
A first communication passage filled with working fluid while bypassing the piston and communicating with the first and second fluid chambers.
A variable capacitance type fluid pressure motor configured to convert the pressure energy of the working fluid flowing in the first continuous passage into rotational energy and to change the push-out volume.
An actuator for changing the push-out volume of the fluid pressure motor and
It has a rotating mass that rotates by transmitting the rotational energy converted by the fluid pressure motor.
In the first control means, the natural frequency of the additional vibration system is changed to the natural frequency in the vicinity of the structure by changing the push-out volume of the hydraulic motor via the actuator during the vibration of the structure. The vibration suppression device for a structure according to claim 1, wherein the inertial mass of the variable rotation inertial mass damper is controlled so as to be synchronized with the above. The variable rotary inertial mass damper is
A conversion mechanism that converts the displacement of the structure transmitted via the elastic member with the vibration of the structure into rotational power, and
It has a stepped speed change mechanism that transmits the rotational power converted by the conversion mechanism to the rotary mass in a state of shifting at one speed ratio selected from a plurality of predetermined gear ratios.
The plurality of gear ratios are a plurality of natural modes of the additional vibration system obtained corresponding to each of the plurality of gear ratios when each of the plurality of gear ratios is selected during vibration of the structure. The frequency is set so as to be tuned corresponding to each of a plurality of predetermined natural frequencies of the structure.
The first control means selects a gear ratio corresponding to the near natural frequency of the structure from the plurality of gear ratios of the speed change mechanism, so that the natural frequency of the additional vibration system becomes the near natural vibration. The vibration suppression device for a structure according to claim 1, wherein the inertial mass of the variable rotation inertial mass damper is controlled so as to be synchronized with a number. The variable rotary inertial mass damper is
A conversion mechanism that converts the displacement of the structure transmitted via the elastic member with the vibration of the structure into rotational power, and
While transmitted to the rotary mass while shifting the rotational power that has been converted in the conversion mechanism, anda transmission mechanism of a continuously variable type that is configured to be able to change the speed change ratio continuously,
The first control means sets the gear ratio of the speed change mechanism during vibration of the structure so that the natural frequency of the additional vibration system is synchronized with the natural frequency in the vicinity of the structure. The vibration suppression device for a structure according to claim 1, wherein the inertial mass of the variable rotation inertial mass damper is controlled. The speed change mechanism has a rotation shaft in which the rotational power converted by the conversion mechanism is transmitted in a state of being changed.
A fitting hole is formed in the rotating mass.
Further provided with a friction material to be fitted between the rotating shaft and the inner peripheral surface of the fitting hole,
The vibration suppression device for a structure according to claim 6 or 7, wherein the rotating mass is connected to the rotating shaft via the friction material. With dampen vibrations of the additional vibration system, and a variable attenuation damper which is capable of changing the number of damping factor,
During the vibration of the structure, according to the fixed point theory, the maximum value of the response magnification of the structure is obtained according to the near natural frequency and the inertial mass of the variable rotational inertial mass damper controlled by the first control means. A second control means for controlling the damping coefficient of the variable damping damper so that
The vibration suppression device for a structure according to any one of claims 1 to 8, further comprising. The variable damping damper
With the cylinder, the piston, the first passage and the working fluid,
It has a regulating valve provided in the first passage and which adjusts the flow resistance of the working fluid flowing in the first passage by changing the opening degree.
The second control means adjusts the opening degree of the adjusting valve according to the near natural frequency and the inertial mass of the variable rotary inertial mass damper controlled by the first control means during the vibration of the structure. A claim that depends on any of claims 2 to 5, wherein the damping coefficient of the variable damping damper is controlled so that the maximum value of the response magnification of the structure is minimized by the modification. 9. The structure vibration suppression device according to 9. The variable damping damper
A cylinder in which displacement due to vibration of the structure is transmitted via the elastic member, and
The inside of the cylinder is divided into two viscous fluid chambers filled with viscous fluid, and displacement due to vibration of the structure is transmitted via the elastic member, so that the inside of the cylinder slides in the axial direction. With a piston configured to
Bypassing the piston, communicating with the two viscous fluid chambers, and communicating with a viscous fluid-filled passageway,
It has a regulating valve provided in the communication passage and adjusting the flow resistance of the viscous fluid flowing in the communication passage by changing the opening degree.
The second control means adjusts the opening degree of the adjusting valve according to the near natural frequency and the inertial mass of the variable rotary inertial mass damper controlled by the first control means during the vibration of the structure. A claim that depends on any of claims 6 to 8, wherein the damping coefficient of the variable damping damper is controlled so that the maximum value of the response magnification of the structure is minimized by the modification. 9. The structure vibration suppression device according to 9.

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