Evaluation method for performance specifications of pressure motor type mass dampers

Abstract

To provide an evaluation method of performance specifications of a pressure motor type mass damper in which the performance specifications of the mass damper can be accurately evaluated and easily adjusted.SOLUTION: An evaluation method of performance specifications of a pressure motor type mass damper includes: not allowing a hydraulic fluid HF to flow in a communication passage 4; setting up a test piece 1E (first test piece) of a mass damper 1, configured so that a gear motor 5 is inoperable, onto a testing device 51 via a reaction jig 59 to vibrate; calculating relief characteristics cr and rigidity kb of contraction spring elements obtained by contracting the hydraulic fluid HF of the master damper 1, spring elements other than the hydraulic fluid HF, and a reaction jig 59, on the basis of the detected input vibration force and displacement (damper force Fd and displacement x); allowing the hydraulic fluid HF to flow in the communication passage 4; setting up an original mass damper 1 (second test piece) where the gear motor 5 is operable onto the testing device 51 to vibrate; and calculating an equivalent mass md and an attenuation coefficient cd of the mass damper 1, on the basis of the detected input vibration force and displacement, and of the rigidity kb of the contraction spring element and the relief characteristics cr.SELECTED DRAWING: Figure 1

Description

Technical Field

[0001] The present invention relates to a pressure motor type mass damper that suppresses the vibration of a structure by means of the rotational inertia mass effect of a rotating mass driven by a pressure motor, and to a method for evaluating performance specifications representing its performance.

Background Art

[0002] The applicant has disclosed, for example in Patent Document 1, a test apparatus for evaluating the performance specifications of a mass damper (rotational inertia mass damper). This mass damper is of the ball screw type, and the relative displacement between the inner cylinder and the screw shaft generated during vibration is converted into the rotational motion of the rotating mass, whereby the rotational inertia mass effect (inertial force) by the rotating mass and the viscous damping effect (viscous force) by the viscous body disposed between the rotating mass and the inner cylinder are exerted, and a vibration suppression effect is obtained.

[0003] The test apparatus is configured such that an actuator, a mass damper, and a connecting member (reaction force jig) are sequentially arranged in a test frame assembled by assembling steel materials in a lattice shape, and load cells and displacement sensors are provided. Then, from the relationship between the damper resistance force and the damper displacement detected when a steady excitation force is input from the actuator to the mass damper, an equivalent mass (rotational inertia mass) and a damping coefficient are calculated (evaluated) as the performance specifications of the mass damper.

Prior Art Documents

Patent Documents

[0004]

Patent Document 1

Summary of the Invention

Problems to be Solved by the Invention

[0005] However, if the performance specification evaluation method for the conventional ball screw type mass damper described above is applied to a pressure motor type mass damper, the following problems arise. As is well known, a pressure motor type mass damper includes, for example, a cylinder filled with working fluid, a piston slidably mounted inside the cylinder and having a relief valve, and a communication passage that bypasses the piston and communicates with the first and second fluid chambers of the cylinder. A pressure motor (e.g., a gear motor) is provided in the communication passage, and a rotating mass is connected to the output shaft of the pressure motor.

[0006] In the pressure motor type mass damper with the above configuration, as the piston reciprocates relative to the cylinder during vibration, the flow of the working fluid in the communication passage is converted into rotational motion of the rotating mass by the pressure motor, thereby exhibiting the rotational inertia mass effect of the rotating mass, and simultaneously the communication passage and The viscous damping effect of the working fluid flowing within the pressure motor provides vibration suppression. Furthermore, when the pressure of the working fluid in either the first or second fluid chamber reaches a predetermined pressure, the relief valve opens, releasing the working fluid into the other fluid chamber and preventing excessive damper reaction force.

[0007] When conducting vibration tests on a pressure motor-type mass damper as described above and evaluating its performance parameters, the elements of the performance parameters to be evaluated are six in total: the compressive stiffness of the working fluid of the mass damper, the stiffness of spring elements other than the working fluid, the relief characteristics of the relief valve (relief load, damping coefficient before and after relief), the equivalent mass and damping coefficient of the mass damper, and the stiffness of the reaction force fixture of the test apparatus. This is considerably more than in the case of a ball screw type. For this reason, in the conventional test apparatus described above, where the mass damper is incorporated into the test frame and the performance parameters are evaluated from the relationship between the damper resistance force and damper displacement detected when a steady excitation force is input to the mass damper from the actuator, it is difficult to accurately calculate and evaluate each of the six elements of performance parameters.

[0008] Furthermore, if the calculated performance specifications, such as the relief characteristics, fall outside the acceptable range, it is necessary to temporarily remove the piping (such as connecting passages), pressure motor, rotating mass, and piston from the mass damper, make adjustments in that state, and then reassemble them. This results in low efficiency in the performance specification adjustment process and increases the manufacturing cost of the mass damper.

[0009] The present invention was made to solve these problems, and aims to provide a method for evaluating the performance specifications of a pressure motor type mass damper that can accurately evaluate the performance specifications of the pressure motor type mass damper and can easily adjust the performance specifications. [Means for solving the problem]

[0010] To achieve this objective, the invention according to claim 1 comprises a cylinder filled with working fluid, a piston slidably mounted within the cylinder and dividing the inside of the cylinder into first and second fluid chambers, a relief valve built into the piston that opens when the pressure in one of the first and second fluid chambers reaches a predetermined pressure and releases the pressure to the other of the first and second fluid chambers, a communication passage that bypasses the piston and communicates with the first and second fluid chambers through a communication port in the cylinder, a pressure motor provided in the communication passage, and a rotating mass connected to the pressure motor, wherein when the structure vibrates, the piston reciprocates relative to the cylinder, and the pressure motor is activated by the flow of working fluid in the communication passage, and the rotational inertia mass effect of the rotating mass and the communication passage andA method for evaluating performance parameters representing the performance of a pressure motor type mass damper for suppressing structural vibrations by the viscous damping effect of a working fluid flowing within the pressure motor, comprising: a first test specimen preparation step of preparing a mass damper configured such that the working fluid does not flow through the communication passage and the pressure motor is inoperable as the first test specimen; a first vibration step of installing the first test specimen in a test apparatus via a reaction force jig, exciting it, and detecting the input excitation force and displacement; a first performance parameter calculation step of calculating the stiffness of the working fluid of the mass damper, the spring elements other than the working fluid, and the compressed spring elements obtained by compressing the reaction force jig, and the relief characteristics of the relief valve, based on the excitation force and displacement detected during the first vibration step; a second test specimen preparation step of preparing a mass damper in its original configuration, such as one in which the working fluid flows through the communication passage and the pressure motor is operable, as the second test specimen; and installing the second test specimen in a test apparatus. via a reaction force jig The system is characterized by comprising: a second excitation step of installing and exciting the system and detecting the input excitation force and displacement; and a second performance specification calculation step of calculating the equivalent mass and damping coefficient of the mass damper based on the excitation force and displacement detected during the second excitation step and the stiffness and relief characteristics of the compressed spring element calculated in the first performance specification calculation step.

[0011] The mass damper to which the present invention is applied is a pressure motor type having the above configuration, and when a structure vibrates, the piston reciprocates relative to the cylinder, and the pressure motor is activated by the flow of the working fluid in the communication passage, resulting in the rotational inertia mass effect due to the rotating mass and the communication passage and The system is configured to suppress structural vibrations by utilizing the viscous damping effect of the working fluid flowing within the pressure motor.

[0012] Furthermore, in the evaluation method for the performance specifications of the pressure motor type mass damper of the present invention, first, a first test specimen is prepared in which the working fluid does not flow through the communication passage and the pressure motor is inoperable, compared to the original configuration of the mass damper. Next, in the first excitation step, this first test specimen is placed in the test apparatus via a reaction force jig and excited, and the excitation force and displacement input to the first test specimen are detected. Then, in the first performance specification calculation step, the stiffness of the working fluid of the mass damper, the spring elements other than the working fluid, and the compressed spring elements obtained by compressing the reaction force jig, as well as the relief characteristics of the relief valve, are calculated based on the excitation force and displacement detected during the first excitation step.

[0013] As described above, in the first excitation process, by exciting the first test specimen with the above configuration, the excitation force and displacement input to the first test specimen are detected while the flow of the working fluid in the connecting passage and the operation of the pressure motor are prevented. As a result, the detected excitation force and displacement include the rotational inertia mass effect due to the rotating mass and the connecting passage and The viscous damping effect of the working fluid in the pressure motor has no influence whatsoever. Furthermore, the three spring elements related to the mass damper and test apparatus—namely, the spring element consisting of the working fluid of the mass damper, the spring element other than the working fluid, and the spring element consisting of the reaction force jig, all connected in series—are reduced to a single equivalent compressed spring element, and its stiffness is evaluated. From the above, the stiffness of the three spring elements can be easily and accurately calculated as the stiffness of a single compressed spring element, and the relief characteristics of the relief valve can be calculated accurately.

[0014] Next, a mass damper in its original configuration, in which the working fluid flows through the communication passage and the pressure motor is operable, is prepared as a second test specimen, and in the second excitation process, the second test specimen is... The test apparatus is installed via a reaction force jig, The second test specimen is subjected to vibration, and the vibration force and displacement applied to it are detected. Then, in the second performance specification calculation process, the equivalent mass and damping coefficient of the mass damper are calculated based on the vibration force and displacement detected during the second vibration process, and the stiffness and relief characteristics of the compressed spring element calculated in the first performance specification calculation process.

[0015] As described above, in the second excitation process, the mass damper in its original configuration is excited as the second test specimen, causing the working fluid to flow through the communication passage and the excitation force and displacement input to the second test specimen to be detected while the pressure motor is operating. As a result, the detected excitation force and displacement are affected by the rotational inertia mass effect due to the rotating mass and the communication passage. and The viscous damping effect caused by the working fluid flowing within the pressure motor is well reflected. Therefore, based on these excitation forces and displacements, and the stiffness and relief characteristics of the compressed spring element calculated in the first performance specification calculation process, the equivalent mass and damping coefficient of the mass damper can be calculated with high accuracy.

[0016] The invention according to claim 2 is a method for evaluating the performance specifications of a pressure motor type mass damper as described in claim 1, further comprising: a first determination step, after the first performance specification calculation step, determining whether the calculated stiffness and relief characteristics of the compression spring element fall within a predetermined allowable range; and a first adjustment step, if, as a result of the first determination, at least one of the stiffness and relief characteristics of the compression spring element falls outside the allowable range, adjusting that at least one to fall within the allowable range.

[0017] With this configuration, if the stiffness and / or relief characteristics of the compressed spring element calculated in the first performance specification calculation step are determined to be outside a predetermined tolerance range, the determined performance specifications can be adjusted to fall within the tolerance range.

[0018] The invention according to claim 3 is a method for evaluating the performance specifications of a pressure motor type mass damper as described in claim 2, characterized in that the first test specimen is constructed by omitting the communication passage, pressure motor and rotating mass compared to the original mass damper configuration, and by sealing the communication port of the cylinder with a cover material.

[0019] In the first test body of this configuration, there are no communication passages, pressure motors, and rotating masses in the mass damper of the original configuration, and the communication ports of the cylinder are sealed with a lid material. Therefore, the adjustment in the first adjustment step for keeping the rigidity and / or relief characteristics of the reduction spring element within the allowable range can be easily performed while directly accessing the spring element and the relief valve without requiring removal work of the communication passage, pressure motor, and rotating mass, and the adjustment work can be minimized.

[0020] The invention according to claim 4 is a method for evaluating the performance specifications of a pressure motor type mass damper according to any one of claims 1 to 3, further comprising: a second determination step of determining whether the calculated equivalent mass and damping coefficient of the mass damper are within their respective predetermined allowable ranges after the second performance specification calculation step; and a second adjustment step of adjusting at least one of the equivalent mass and damping coefficient to be within the allowable range when, as a result of the second determination, at least one of the equivalent mass and damping coefficient is not within the allowable range.

[0021] According to this configuration, when it is determined that the equivalent mass and / or damping coefficient calculated in the second performance specification calculation step are not within their respective predetermined allowable ranges, the performance specifications can be appropriately adjusted to be within the allowable range.

Brief Description of the Drawings

[0022] [Figure 1] It is a diagram schematically showing a test device for evaluating the performance specifications of a pressure motor type mass damper together with the mass damper. [Figure 2] It is a cross-sectional view showing the mass damper. [Figure 3] It is a block diagram showing a control device of the test device and the like. [Figure 4] (a) Mass damper, (b) Mass damper and test device, and (c) Mass damper and test device in which a plurality of spring elements in (b) are reduced to a single spring element, each modeled and shown in the figure. [Figure 5]This is a flowchart showing the procedure for evaluating and adjusting the performance specifications of a mass damper. [Figure 6] This figure, similar to Figure 2, shows a test specimen of the mass damper used in the first step of Figure 5. [Figure 7] This figure, corresponding to Figure 4(c), shows a model of the test specimen and test apparatus in Figure 6. [Figure 8] This figure shows the actuator displacement and damper force obtained from the first step of the vibration test, as well as the calculated performance specifications. [Figure 9] This diagram illustrates the method for calculating (a) the damping coefficient and (b) the equivalent mass of a mass damper. [Figure 10] This diagram schematically shows a test apparatus different from that in Figure 1, along with the mass damper, for evaluating the performance specifications of a pressure motor type mass damper. [Modes for carrying out the invention]

[0023] Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. Figure 1 shows a mass damper (rotational inertia mass damper) 1 installed in a test apparatus 51. This mass damper 1 is a gear motor type, using a gear motor 5 as the pressure motor, and is installed in structures such as buildings, for example, to suppress vibrations of the structure by generating a damping force when the structure vibrates. First, the configuration and operation of this gear motor type mass damper 1 will be described below.

[0024] As shown in Figure 2, the mass damper 1 comprises a cylindrical cylinder 2, a piston 3 slidably mounted inside the cylinder 2, a communication passage 4 that bypasses the piston 3 and communicates with the inside of the cylinder 2, a gear motor 5 acting as a pressure motor located in the communication passage 4, and a flywheel 9 connected to the output shaft 8 of the gear motor 5.

[0025] The cylinder 2 has a circumferential wall 2a and first and second end walls 2b and 2c provided at both ends of the circumferential wall 2a. A projection 2d having a rod housing chamber 2g is concentrically provided on the first end wall 2b, and a first mounting fixture FL1 is provided at its end via a universal joint.

[0026] The piston 3 is slidably mounted within the cylinder 2, dividing the internal space of the cylinder 2 into a first fluid chamber 2e and a second fluid chamber 2f. The first and second fluid chambers 2e and 2f and the connecting passage 4 are filled with hydraulic fluid HF. The hydraulic fluid HF is a conventional type with appropriate viscosity.

[0027] A piston rod 10 is integrally mounted concentrically on the piston 3. The piston rod 10 extends on both sides of the piston 3, and on the side of the second end wall 2c, it passes through its rod guide hole in a liquid-tight manner and extends outward. A second mounting fixture FL2 is provided at the outer end of the piston rod 10 via a universal joint. On the side of the first end wall 2b, the piston rod 10 also passes through its rod guide hole in a liquid-tight manner and extends into the rod housing chamber 2g of the protruding portion 2d, and a piston accumulator 31 is provided at its end.

[0028] The piston accumulator 31 is for storing pressure caused by the thermal expansion of the hydraulic fluid HF, and has a hollow casing portion 32 formed at the end of the piston rod 10, a piston 34 slidably mounted within the casing portion 32 and defining an oil chamber 33 on the piston 3 side, and a set spring 35 that biases the piston 34 toward the oil chamber 33. In addition, a rod communication hole 10a is formed along the piston rod 10. The rod communication hole 10a communicates with the oil chamber 33 at one end and extends to the center of the piston 3 at the other end.

[0029] On the other hand, the piston 3 has first and second communication holes that penetrate axially and communicate with the first and second fluid chambers 2e and 2f, and a third communication hole that extends vertically connecting the first and second communication holes and communicates with the rod communication hole 10a. Check valves 36, 36 are provided on both sides of the third communication hole in the first communication hole. Each check valve 36 is configured to allow the flow of hydraulic fluid HF only from the third communication hole side to the first or second fluid chamber 2e and 2f side. In addition, orifices 37, 37 are provided on both sides of the third communication hole in the second communication hole.

[0030] In the above configuration, when the pressure of the hydraulic fluid HF in the cylinder 2 increases due to a rise in temperature of the hydraulic fluid HF, the hydraulic fluid HF slowly flows from the first and second fluid chambers 2e and 2f into the oil chamber 33 of the piston accumulator 31 via the second communication hole of the piston 3, the orifices 37, 37, the third communication hole, and the rod communication hole 10a. As a result, the set spring 35 is compressed via the piston 34, and the pressure of the hydraulic fluid HF is stored in the piston accumulator 31, thereby preventing malfunctions caused by the rise in the pressure of the hydraulic fluid HF due to temperature rise or other factors.

[0031] Furthermore, the piston 3 has a first relief passage 3d and a second relief hole 3e that penetrate in the axial direction. The first and second relief passages 3d and 3e are provided with a first relief valve 11 and a second relief valve 12, respectively. The first and second relief valves 11 and 12 have the same configuration as each other and are configured as normally closed valves, and each has a valve body and a spring that biases the valve body in the closing direction.

[0032] The first relief valve 11 closes the first communication passage 3d until the pressure of the hydraulic fluid HF in the first fluid chamber 2e reaches a predetermined pressure, and then opens the first communication passage 3d when the predetermined pressure is reached. As a result, the pressure in the first fluid chamber 2e is released to the second fluid chamber 2f side via the first communication passage 3d, and is limited to a pressure below the predetermined pressure plus the effect of the damping coefficient after relief. Similarly, the second relief valve 12 closes the second communication passage 3e until the pressure in the second fluid chamber 2f reaches a predetermined pressure, and then opens the second communication passage 3e when the predetermined pressure is reached. As a result, the pressure in the second fluid chamber 2f is released to the first fluid chamber 2e side via the second communication passage 3e, and is limited to a pressure below the predetermined pressure plus the effect of the damping coefficient after relief.

[0033] The gear motor 5 is, for example, an internal gear motor and is positioned in the center of the communication passage 4. The gear motor 5 has a housing 6 that communicates with the communication passage 4 via two inlets 6a, 6a, a rotatable input gear and an output gear (neither shown) housed in the housing 6 and meshing with each other, and an output shaft 8 integrally provided with the output gear. The housing 6 is supported by the peripheral wall 2a of the cylinder 2. A drain passage (not shown) for discharging the hydraulic fluid HF is also provided inside the housing 6. The output shaft 8 is liquid-tightly supported by the housing 6 via a seal (not shown).

[0034] One end of a drain pipe 41 is connected to the drain passage of the housing 6, and the other end of the drain pipe 41 passes through the protruding portion 2d of the cylinder 2, the casing portion 32 of the piston accumulator 31, and the piston 34, and communicates with the oil chamber 33. In this configuration, when the pressure of the hydraulic fluid HF inside the housing 6 rises, the hydraulic fluid HF inside the housing 6 is discharged from the drain passage through the drain pipe 41 to the oil chamber 33 of the piston accumulator 31, thereby releasing the pressure and preventing excessive pressure buildup inside the housing 6.

[0035] The flywheel 9 is made of a material with a relatively high specific gravity, such as steel, and is formed, for example, in the shape of a disc, and is integrally mounted coaxially on the output shaft 8.

[0036] The mass damper 1, configured as described above, is installed, for example, between two relative displacing parts within a structure via first and second mounting fixtures FL1 and FL2, and is used as a seismic damping device. When the structure vibrates during an earthquake or other event, the piston 3 moves within the cylinder 2 in accordance with the relative displacement between the two parts. Consequently, the hydraulic fluid HF in the first or second fluid chambers 2e and 2f is pushed out by the piston 3, flows into the communication passage 4, flows through the housing 6 of the gear motor 5, and then flows into the second or first fluid chambers 2f and 2e.

[0037] The pressure generated by the flow of the hydraulic fluid HF is converted into rotational motion of the input and output gears of the gear motor 5, and the flywheel 9, which is integrated with the output shaft 8, is rotated, thereby exerting a rotational inertia mass effect (inertial force). In addition, the viscous damping effect (viscous force) due to the viscous resistance when the hydraulic fluid HF flows through the communication passage 4 and the gear motor 5 is exerted, and together with the rotational inertia mass effect, a vibration suppression effect on the structure is obtained.

[0038] Furthermore, based on the above-described configuration and operation, the mass damper 1 is modeled as shown in Figure 4(a). That is, it is a model in which (a) an inertial mass element with equivalent mass md consisting of a flywheel 9 and hydraulic fluid HF, and (b) a viscous element with damping coefficient cd consisting of hydraulic fluid HF flowing through the communication passage 4 and gear motor 5 are connected in series to (c) a spring element with compressive stiffness kds consisting of the hydraulic fluid HF of the mass damper 1, (d) a spring element with stiffness kdo other than the hydraulic fluid HF of the mass damper 1, and (e) a relief element with relief characteristic cr consisting of the first and second relief valves 11 and 12 (viscous element with damping coefficient c1 before relief, limiting element for relief load Fr, and viscous element with damping coefficient c2 after relief).

[0039] Next, the test apparatus 51 will be described with reference to Figures 1 and 3. For convenience, in the following description, the upper and lower sides of Figure 1 will be referred to as "upper" and "lower," the left and right sides as "left" and "right," and the front and back sides as "front" and "back." As shown in Figure 1, the test apparatus 51 consists of upper, lower, left, and right frames 52, 53, 54, and 55 assembled integrally in a grid pattern, and includes a test frame 56 in which the mass damper 1 is housed, an actuator 57 for inputting excitation force to the mass damper 1, a connecting member 58 for connecting the actuator 57 to the mass damper 1, a reaction force jig 59 connected to the right frame 55 to receive the reaction force of the mass damper 1, and a guide mechanism 60 provided on the lower frame 53 to guide the movement of the connecting member 58 in the left-right direction.

[0040] The upper and lower left and right frames 52-55, the connecting member 58, and the reaction force jig 59 are made of steel materials such as H-beams. The mass damper 1 is positioned so that its axial direction extends in the left-right direction, and is connected to the connecting member 58 via the first mounting fixture FL1 and to the reaction force jig 59 via the second mounting fixture FL2.

[0041] The actuator 57 is composed of, for example, a solenoid and has a main body 57a attached to the left frame 54 and a plunger 57b connected to a connecting member 58 via a load cell 61. The actuator 57 is controlled by a control device 63 (see Figure 3), which outputs an excitation force from the plunger 57b.

[0042] As shown in Figure 3, the test apparatus 51 further includes the load cell 61, a displacement sensor 62 provided on the mass damper 1, and a control device 63 that controls the actuator 57. The load cell 61 is, for example, a strain gauge type, which detects the load acting on the connecting member 58 as the force acting on the mass damper 1 (hereinafter referred to as "damper force") Fd, and outputs the detection signal to the control device 63.

[0043] The displacement sensor 62 is, for example, a laser type, and detects the displacement of the plunger 57b relative to the main body 57a of the actuator 57 as the displacement x input from the actuator 57 to the mass damper 1 (hereinafter referred to as "actuator displacement"), and outputs the detection signal to the control device 63. The control device 63 is composed of a combination of a power supply for driving the actuator 57, a rectifier, a CPU, RAM, ROM, I / O interface, etc. (none of which are shown).

[0044] Based on the above configuration of the test apparatus 51, the model when the mass damper 1 is installed in the test apparatus 51 is represented as shown in Figure 4(b). That is, the relief element of the mass damper 1 with relief characteristic cr shown in Figure 4(a) is connected to the stationary body via a spring element with rigidity kbj consisting of a reaction force jig 59, and the actuator 57 is connected to the opposite side of the mass damper 1.

[0045] Furthermore, since the three spring elements in Figure 4(b) (compression stiffness kds, stiffness kdo, and stiffness kbj) are in series with respect to each other, it is possible to reduce (combine) them into a single spring element. In that case, the model in Figure 4(b) is represented as shown in Figure 4(c). The stiffness kb of the reduced spring element (hereinafter referred to as the "reduced spring element") is expressed by the following equation (1). 1 / kb = 1 / kds+1 / kdo+1 / kbj (1)

[0046] Furthermore, the relationship between the displacements of the elements in the models of Figures 4(b) and (c) is expressed by the following equation (2). x = xbj+xiHGD = xb+xr+xd ···(2) Here, x: actuator displacement xbj: Displacement of spring element in reaction force jig xiHGD: Damper displacement xb: Displacement of the compressed spring element xr: Displacement of the relief element xd: Displacement of the inertial mass element (viscous element)

[0047] Furthermore, the force (load) relationships between elements in the models of Figures 4(b) and (c) are expressed by the following equations (3) and (4), depending on the relief conditions of the first and second relief valves 11 and 12. When |Fd| ≤ Fr (condition before relief pitching) Fd = kbj·xbj = kb·xb = c1·vr = md·αd+cd·vd ···(3) When |Fd|>Fr (condition after relief) Fd = kbj·xbj = kb·xb = sgn(vr)·Fr+c2·(vr-sgn(vr)·Fr / c1) = md·αd+cd·vd ···(4) Here, Fd: Actuator excitation force (= damper force) Fr: Relief load c1: Damping coefficient before relief c2: Damping coefficient after relief VR:XR speed vd:xd speed Acceleration of αd:xd md: Equivalent mass of the inertial mass element cd: Damping coefficient of the viscous element

[0048] Next, the method for evaluating the performance characteristics of the mass damper 1 will be explained with reference to Figures 5 to 9. This evaluation method involves conducting a vibration test of the mass damper 1 using the test apparatus 51, and calculating the performance characteristics of the mass damper 1, such as the stiffness kb of the compressed spring element, the relief characteristics cr (relief load Fr, damping coefficients c1 and c2 before and after relief), the equivalent mass md, and the viscosity coefficient cd, from the relationship between the damper force and damper displacement detected at that time. Furthermore, if the calculated performance characteristics do not fall within a predetermined allowable range, the mass damper 1 is adjusted to fall within the allowable range.

[0049] First, in the first step shown in Figure 5 (indicated as "S1"; the same applies hereafter), the first step is performed. In this first step, a vibration test is conducted using the test apparatus 51 on a test specimen 1E of a mass damper 1 configured so that the hydraulic fluid HF does not flow through the communication passage 4 and the gear motor 5 does not operate. From the vibration results, the stiffness kb of the compressed spring element and the relief characteristic cr are calculated. As shown in Figure 6, this test specimen 1E differs from the mass damper 1 in Figure 2 in that it does not have (is not attached) the communication passage 4, gear motor 5, flywheel 9, and drain piping 41, and each communication port 2h of the cylinder 2 that communicates with the communication passage 4 is sealed with a cover plate 71.

[0050] In this configuration, even when the test specimen 1E is vibrated, the hydraulic fluid HF does not flow into the communication passage 4, and the gear motor 5 does not operate. Therefore, the combined model of the test specimen 1E and the test apparatus 51, as shown in Figure 7, is a model in which the inertial mass element and viscous element are removed from the model in Figure 4(c), and the relief element and the compressed spring element are connected. Accordingly, the relationship of displacement between these elements is expressed by the following equation (5). x = xb + xr ···(5)

[0051] In this first step, the test specimen 1E is set in the test apparatus 51, and sinusoidal excitation (e.g., sinusoidal displacement control) is performed from the actuator 57 to the test specimen 1E under conditions of excitation level before relief reaching the relief load Fr, and conditions of excitation level after relief reaching the relief load Fr. Then, the stiffness kb of the compressed spring element and the relief characteristic cr are calculated from the relationship between the actuator displacement x and the damper force Fd, as shown in Figure 8, which is sampled and stored under each condition.

[0052] Specifically, first, based on the excitation results under the pre-relief conditions (Figures 8(a)~(c)), the stiffness kb of the compressed spring element is calculated from the relationship between the maximum value of the damper force Fd and the actuator displacement x at that time, and from equation (3). Next, the displacement xb of the compressed spring element is calculated from this stiffness kb and damper force Fd, and the displacement xr of the relief element is calculated by subtracting this displacement xb from the actuator displacement x. Then, the velocity vr of the relief element is calculated by differentiating the displacement xr, and the damping coefficient c1 of the relief element before relief is calculated from this velocity vr and damper force Fd, based on equation (3).

[0053] Subsequently, based on the excitation results corresponding to the input level of the post-relief conditions (Figures 8(d) to (f) show one such case), the relief load Fr is calculated, and the post-relief damping coefficient c2 of the relief element is calculated from the relief load Fr and the damping coefficient c1 before relief, based on equation (4). As a result, a bilinear relief characteristic cr as shown in Figure 8(g) is obtained.

[0054] As described above, in the first step, by vibrating the test specimen 1E shown in Figure 6, unlike in the case of the mass damper 1, the damper force F and actuator displacement x input to the test specimen 1E are detected while the flow of the hydraulic fluid HF in the communication passage 4 and the operation of the gear motor 5 are prevented. As a result, the detected damper force F and actuator displacement x are not affected at all by the rotational inertia mass effect due to the rotating mass or the viscous damping effect due to the working fluid in the communication passage. Furthermore, the three spring elements related to the mass damper 1, namely the spring element made of the hydraulic fluid HF of the mass damper 1, the spring element other than the hydraulic fluid HF, and the spring element made of the reaction force jig, which are connected in series with each other, are reduced to a single equivalent compressed spring element, and its stiffness is evaluated. From the above, the stiffnesses kbs, kdo, and kbj of the above three spring elements can be easily and accurately calculated as the stiffness kb of a single compressed spring element, and the relief characteristics cr of the first and second relief valves 11 and 12 can be calculated accurately.

[0055] In the next step, the second step, it is determined whether the performance parameters calculated in the first step (stiffness kb of the compressed spring element, relief characteristics cr (relief load Fr, damping coefficients c1 and c2 before and after relief)) fall within their respective predetermined allowable ranges. If the result of this determination is NO, and at least one of the stiffness kb and relief characteristics cr is outside the allowable range, the process proceeds to the third step, where the performance parameters are adjusted to fall within the allowable range. This adjustment is performed, for example, if the relief load Fr is outside the allowable range, by removing the first and second relief valves 11 and 12 from the piston 3 and adjusting the stiffness of the springs and the initial strain amount (preload). After such adjustments are made, the first step is executed again to calculate the stiffness kb of the compressed spring element and the relief characteristics cr.

[0056] As described above, in the second step, if it is determined that the stiffness kb and / or relief characteristics cr of the compressed spring element calculated in the first step are outside the predetermined allowable range, the performance specifications can be adjusted in the third step to bring them within the allowable range. Furthermore, the test specimen 1E used in the vibration test in the first step does not have the connecting passage 4, gear motor 5, flywheel 9, or drain piping 41 attached to the mass damper 1. Therefore, the adjustment in the third step to bring the stiffness kb and / or relief characteristics cr of the compressed spring element within the allowable range can be easily performed by directly accessing the spring element and the first and second relief valves 11 and 12 without requiring the removal of the connecting passage 4, gear motor 5, and flywheel 9, thereby minimizing the adjustment work.

[0057] If the result of the second step is YES, and the stiffness kb of the compressed spring element and the relief characteristic cr are within the acceptable range, the process proceeds to the fourth step, where the second step is executed. In this second step, instead of the test specimen 1E from the first step, the mass damper 1 with the original configuration shown in Figure 2, which includes the connecting passage 4 and gear motor 5, is set in the test apparatus 51 as the second test specimen, and sinusoidal excitation is performed from the actuator 57 to the mass damper 1 under the conditions of the excitation level before relief. Then, based on the actuator displacement x and damper force Fd sampled and stored from the excitation results, the displacement xb of the compressed spring element and the displacement xd of the inertial mass element (viscous element) are calculated.

[0058] Specifically, the velocity vr of the relief element is calculated based on equation (3) using the damper force Fd obtained from the excitation results and the damping coefficient c1 before relief calculated in the first step, and the displacement xr of the relief element is calculated by integrating this velocity vr. Also, the displacement xb of the compressed spring element is calculated based on equation (3) using the damper force Fd and the stiffness kb of the compressed spring element calculated in the first step. Then, using the calculated displacements xr and xb, the displacement xd of the inertial mass element (viscous element) is calculated by the following equation (6), which is a transformation of equation (2). xd = x - xb - xr ... (6)

[0059] In the next fifth step, the third step is performed to calculate the damping coefficient cd of the viscous element and the equivalent mass md of the inertial mass element from the displacement xd of the inertial mass element (viscous element) obtained in the second step. Specifically, as shown in Figure 9(a), the damper force Fd when the velocity vd of the viscous element, obtained by differentiating the displacement xd, is at the maximum velocity vmax (when the displacement xd is 0) is sampled as the maximum damping force Qv. In this case, the maximum damping force Qv can be positive (+Qv) or negative (-Qv), depending on the sign of the displacement xd. Therefore, the maximum damping force Qv is calculated as the average of the absolute values ​​of +Qv and -Qv. Next, the damping coefficient cd is calculated by dividing this maximum damping force Qv by the maximum velocity vmax.

[0060] Furthermore, for the equivalent mass md of the inertial mass element, as shown in Figure 9(b), the damper force Fd when the displacement xd reaches its maximum value +δmax (when mass damper 1 is most extended) or when the displacement xd reaches its minimum value -δmax (when mass damper 1 is most compressed) is sampled as the maximum inertial force Qi. Next, the maximum value of the acceleration αd obtained by differentiating the displacement xd twice (maximum acceleration) αdmax is calculated, and the equivalent mass md is calculated by dividing the maximum inertial force Qi by the maximum acceleration αdmax.

[0061] As described above, in the second and third steps, by exciting the mass damper 1 in its original configuration as shown in Figure 2, the hydraulic fluid HF flows through the communication passage 4, and with the gear motor 5 operating, the damper force Fd and actuator displacement x input to the mass damper 1 are detected. As a result, the detected damper force Fd and actuator displacement x accurately reflect the effects of the rotational inertia mass effect due to the flywheel 6 and the viscous damping effect due to the hydraulic fluid HF flowing through the communication passage 4 and gear motor 5. Therefore, based on these damper forces Fd and actuator displacement x, and the stiffness kb and relief characteristics cr of the compressed spring element calculated in the first step, the equivalent mass md and damping coefficient cd of the mass damper 1 can be calculated with high accuracy.

[0062] In the next, sixth step, it is determined whether the calculated equivalent mass md and damping coefficient cd fall within their respective predetermined allowable ranges. If the result of this determination is NO, and at least one of the equivalent mass md and damping coefficient cd is outside the allowable range, the process proceeds to the seventh step, where the performance is adjusted to fall within the allowable range. This adjustment is performed, for example, by adjusting the size and weight of the flywheel 6 if the equivalent mass md is outside the allowable range. After such adjustments are made, the second and third steps are performed again to calculate the equivalent mass md and damping coefficient cd.

[0063] As described above, in the sixth step, if it is determined that the equivalent mass md and / or damping coefficient cd calculated in the third step are outside the predetermined allowable range, the performance specifications can be adjusted in the seventh step to bring them within the allowable range.

[0064] Then, if the result of the sixth step is YES, and the equivalent mass md and damping coefficient cd fall within the acceptable range, the calculation and verification of all performance parameters is considered complete, and the evaluation process shown in Figure 5 is terminated.

[0065] It should be noted that the present invention is not limited to the embodiments described and can be implemented in various ways. For example, the vibration test for evaluating the equivalent mass md and damping coefficient cd in the second and third steps may be performed using the test apparatus 81 shown in Figure 10. As is clear from comparison with Figure 1, the test apparatus 81 is provided with a flexible member 82 consisting of a disc spring unit and a rubber unit as a vibration jig on the right side (mass damper 1 side) of the connecting member 58 of the test apparatus 51.

[0066] In this case, the spring element made of the flexible member 82 is arranged in series with the compressed spring element (stiffness kb) described above. If its stiffness is kbs, then the stiffness kall of the entire spring element, including the compressed spring element, is expressed by the following equation (7). 1 / kall = 1 / kb + 1 / kbs ···(7)

[0067] From this relationship, the natural period of the additional vibration system, which consists of an inertial mass element with equivalent mass md and a whole spring element with stiffness kall, can be freely set to a long period by adjusting the stiffness kbs of the flexible member 82. As a result, it becomes possible to perform vibration tests on the mass damper 1 with the natural period of the additional vibration system set to, for example, near the natural period of the structure to be damped.

[0068] Furthermore, in order to perform the vibration test in the first step with the flow of hydraulic fluid HF in the communication passage 4 and the operation of the gear motor 5 blocked, the test specimen 1E shown in Figure 6 is used, which is formed by removing the communication passage 4, gear motor 5, and flywheel 6 from the mass damper 1 shown in Figure 2, and sealing the communication port 2h of the cylinder 2 with a cover plate 71. However, the test specimen is not limited to this, and other test specimens with the same configuration that can be obtained may be used.

[0069] For example, although not shown in the diagram, the configuration of the mass damper 1 in Figure 2 may be maintained, and a manual on / off valve may be installed at the communication port 2h of the cylinder 2 and kept in a closed state to prevent the hydraulic fluid HF from flowing into the communication passage 4. Alternatively, the rotation of the input gear and output gear of the gear motor 5 may be locked to prevent the hydraulic fluid HF from flowing. However, in these two examples, unlike the case of test specimen 1E in Figure 6, if it is determined that the relief characteristic cr is outside the acceptable range in the second step, for example, it will be necessary to remove the communication passage 4, gear motor 5, etc. from the mass damper 1 in order to adjust it.

[0070] Furthermore, although a gear motor is used as the pressure motor for the mass damper 1, other types of pressure motors may be used as long as they convert the flow of the working fluid into rotational motion, such as a piston motor, vane motor, or screw motor. Also, although it has been explained in the embodiment that ordinary hydraulic oil is used as the working fluid for the damper, it goes without saying that other suitable working fluids may be used. In addition, the detailed configuration can be appropriately changed within the scope of the spirit of the present invention. [Explanation of Symbols]

[0071] 1. Mass damper (2nd test specimen) 1E Test specimen (First test specimen) 2 liters 2e 1st fluid chamber 2f 2nd fluid chamber 2h communication port 3 pistons 4 passages 5. Gear motor (pressure motor) 9. Flywheel (rotating mass) 11. First relief valve (relief valve) 12. Second relief valve (relief valve) 51 Test equipment 59 Reaction jig 61 Load Cells 62 Displacement Sensor 71 Lid plate (lid material) 81 Test equipment HF hydraulic oil (working fluid) Fd Damper force (input excitation force) x Actuator displacement (input displacement) kb Stiffness of compressed spring element (performance specifications) CR relief characteristics (performance specifications) md Equivalent mass (performance specifications) of inertial mass element cd viscous element damping coefficient (performance specifications)

Claims

1. A method for evaluating performance parameters representing the performance of a pressure motor type mass damper for suppressing vibrations of a structure, comprising: a cylinder filled with working fluid; a piston slidably mounted within the cylinder and dividing the cylinder into first and second fluid chambers; a relief valve built into the piston that opens when the pressure in one of the first and second fluid chambers reaches a predetermined pressure, releasing the pressure to the other of the first and second fluid chambers; a communication passage that bypasses the piston and communicates with the first and second fluid chambers via a communication port in the cylinder; a pressure motor provided in the communication passage; and a rotating mass connected to the pressure motor, wherein when the piston reciprocates relative to the cylinder during vibration of the structure, the pressure motor is activated by the flow of the working fluid in the communication passage, and the rotational inertia mass effect of the rotating mass and the viscous damping effect of the working fluid flowing in the communication passage and the pressure motor are used to suppress vibrations of the structure. A first test specimen preparation step involves preparing the mass damper as a first test specimen, in which the working fluid does not flow through the communication passage and the pressure motor is configured to be inoperable, The first vibration step involves placing the first test specimen in the test apparatus via a reaction force jig, exciting it, and detecting the input vibration force and displacement. A first performance specification calculation step calculates the stiffness of the working fluid of the mass damper, the spring elements other than the working fluid, and the compressed spring elements obtained by compressing the reaction force jig, and the relief characteristics of the relief valve, based on the excitation force and displacement detected during the first excitation step. A second test specimen preparation step involves preparing the mass damper in its original configuration, in which the working fluid flows through the communication passage and the pressure motor is operable, as a second test specimen. The second test specimen is placed in the test apparatus via the reaction force jig, and vibration is applied, and the input vibration force and displacement are detected in the second vibration step, A method for evaluating the performance specifications of a pressure motor type mass damper, comprising: a second performance specification calculation step for calculating the equivalent mass and damping coefficient of the mass damper based on the excitation force and displacement detected during the second excitation step, and the stiffness of the compressed spring element and the relief characteristics calculated in the first performance specification calculation step.

2. After the first performance specification calculation step, a first determination step is performed to determine whether the calculated stiffness of the compressed spring element and the relief characteristics fall within predetermined allowable ranges, A method for evaluating the performance specifications of a pressure motor type mass damper according to claim 1, further comprising: a first adjustment step of adjusting at least one of the stiffness of the compression spring element and the relief characteristics so that they fall within the allowable range, if, as a result of the first determination, at least one of them falls outside the allowable range.

3. The method for evaluating the performance specifications of a pressure motor type mass damper according to claim 2, characterized in that the first test specimen is constructed by omitting the communication passage, the pressure motor, and the rotating mass compared to the mass damper with the original configuration, and by sealing the communication port of the cylinder with a cover material.

4. After the second performance specification calculation step, a second determination step is performed to determine whether the calculated equivalent mass and damping coefficient of the mass damper fall within predetermined allowable ranges. A method for evaluating the performance specifications of a pressure motor type mass damper according to any one of claims 1 to 3, further comprising: a second adjustment step of adjusting at least one of the equivalent mass and damping coefficient so that it falls within the allowable range when, as a result of the second determination, at least one of them falls within the allowable range.

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