JP6490563B2 - Mass damper testing equipment - Google Patents

Description

  The present invention relates to a mass damper testing apparatus that configures an additional vibration system together with a rigid transmission member, has a rotating mass, and tests the performance of a mass damper for suppressing vibration of a structure.

  Conventionally, as an apparatus for testing this type of mass damper, one disclosed in, for example, Patent Document 1 has been proposed by the present applicant. This conventional test apparatus includes a test frame in which upper, lower, left, and right frames made of steel are combined in a cross beam shape, an actuator provided in the test frame, and a load cell. The mass damper has a cylindrical portion, a screw shaft that can move in the axial direction relative to the cylindrical portion, and a rotating mass that rotates as the screw shaft moves relative to the cylindrical portion. The cylinder portion of the mass damper is attached to the right frame via a connecting member, and the screw shaft is attached to the actuator via a load cell. The resistance force of the mass damper is detected by the load cell. The actuator is attached to the left frame and generates an excitation force having a predetermined excitation frequency.

  Due to the relatively high rigidity of the entire right frame and the connecting member, the natural frequency of the vibration system determined by the overall rigidity of the right frame and the connecting member and the rotational inertia mass of the mass damper is the excitation frequency of the actuator. When the frequency is considerably higher than the above, among the frequency components included in the resistance force of the mass damper detected by the load cell, the component corresponding to the natural frequency of the vibration system becomes large. As a result, irregular undulations appear in the detected resistance force of the mass damper. As a result, performance parameters representing the performance of the mass damper such as the rotational inertial mass of the rotating mass cannot be appropriately calculated based on the resistance force. For details of the swell, refer to FIG.

  Therefore, in this conventional test apparatus, a disc spring unit as an adjustment member is provided in series between the screw shaft and the load cell, and the overall rigidity of the adjustment member, the right frame, and the connecting member, and the rotational inertia mass of the mass damper The rigidity of the adjustment member is adjusted so that the natural frequency of the vibration system determined by the above approaches the excitation frequency of the actuator. As a result, among the frequency components included in the detected resistance force of the mass damper, the component corresponding to the excitation frequency is increased, thereby suppressing the swell as described above and appropriately calculating the performance parameter. In addition, the performance of the mass damper is appropriately evaluated based on the performance parameter.

Japanese Patent Application No. 2015-157949

  As described above, in the conventional test apparatus, the operator performs the installation work for installing the adjustment member and the adjustment work for adjusting the rigidity of the adjustment member so that the natural frequency of the vibration system approaches the vibration frequency. It must be done manually, and the test becomes complicated accordingly. In addition, since the adjustment member is provided, the test apparatus becomes complicated.

  The present invention has been made to solve the above-described problems, and the performance of the mass damper can be achieved even when the vibration frequency of the actuator is relatively low and the rigidity of the support member is relatively high. It is an object of the present invention to provide a mass damper test apparatus capable of appropriately calculating parameters, performing performance test confirmation work relatively easily, and simplifying the configuration.

  In order to achieve the above object, the invention according to claim 1 is configured to test the performance of a mass damper for constituting an additional vibration system together with a rigid transmission member, having a rotating mass, and suppressing vibration of a structure. Mass damper test apparatus 1, 41 having rigidity and supporting member MS connected in series to a test target mass damper 21 to be tested (right frame 5 in the embodiment (hereinafter the same in this section)) , Connecting members 7, 9), an actuator 6 for inputting an excitation force to the vibration system S including the test target mass damper 21 and the support member MS, and the test target mass damper 21 whose specifications are predetermined ideal values. A virtual adjustment member MVA for bringing the natural frequency of the first virtual vibration system VS1 in which the ideal mass damper 21I and the support member MS are connected in series with each other to a predetermined frequency. The predicted value of the displacement of the virtual adjustment member MVA when assuming that an excitation force of a predetermined frequency is input to the second virtual vibration system VS2 connected in series to the first virtual vibration system VS1 (virtual adjustment member prediction) Virtual displacement acquisition means (control device 15, steps 2 to 9 in FIG. 5) for acquiring the displacement xb1pr) based on the vibration equation (formulas (1), (11), (22)) of the second virtual vibration system VS2. Step 41 in FIG. 8, steps 52 to 55 in FIG. 10, basic input displacement (basic value xbase) for inputting an excitation force of a predetermined frequency to the vibration system S, and displacement of the acquired virtual adjustment member The control means (control device 15, steps 10 and 11 in FIGS. 5, 8, and 10) that controls the excitation force of the actuator 6 according to the predicted value of the actuator, and the damper that is the resistance force of the test target mass damper 21 Dunn that detects resistance P The performance of the mass damper 21 to be tested in accordance with the resistance force P detected when the excitation force of the actuator 6 controlled by the resistance detection means (load cell 11) and the control means is input to the vibration system S. Performance parameter calculating means (control device 15, steps 31 to 33 in FIG. 7) for calculating performance parameters (maximum speed resistance force Qv, damping coefficient Ceq, rotational inertia mass md).

  According to this configuration, the rigid support member is connected in series to the test target mass damper to be tested, and the excitation force of the actuator is input to the vibration system including the test target mass damper and the support member. . In addition, a virtual adjustment member for bringing the natural frequency of the first virtual vibration system in which the ideal mass damper as the test target mass damper whose specifications are predetermined ideal values and the support member are connected in series to a predetermined frequency. However, when it is assumed that an excitation force of a predetermined frequency is input to the second virtual vibration system connected in series to the first virtual vibration system, the predicted value of the displacement of the virtual adjustment member is the second virtual vibration system. Is acquired by the virtual displacement acquisition means based on the vibration equation. Furthermore, the excitation force of the actuator is controlled by the control means in accordance with the basic input displacement for inputting the excitation force of a predetermined frequency to the vibration system and the obtained predicted value of the displacement of the virtual adjustment member. .

  As is clear from the definition of the virtual adjustment member and the definition of the predicted value of the displacement of the virtual adjustment member described above, the excitation force of the actuator is determined according to the basic input displacement and the predicted value of the displacement of the virtual adjustment member. The test object when it is assumed that the excitation force of a predetermined frequency is input to the vibration system in which the virtual adjustment member is connected in series as the resistance force of the test object mass damper during the input of the excitation force by controlling A resistance force equivalent to that of the mass damper can be obtained.

  According to the above-described configuration, the test object is tested by the performance parameter calculation unit according to the damper resistance force of the test target mass damper detected when the excitation force of the actuator controlled by the control unit is input to the test target mass damper. A performance parameter representing the performance of the mass damper is calculated. As described above, among the frequency components included in the detected damper resistance force, the component corresponding to the frequency of the excitation force of the actuator can be increased, so that the swell as described above can be suppressed. It is possible to appropriately calculate the performance parameter according to the resistance force.

  Further, the virtual adjustment member is a virtual member and is not actually provided in the test apparatus, and the adjustment member of the conventional test apparatus described above is unnecessary. Accordingly, it is possible to relatively easily perform the performance test confirmation and to simplify the configuration.

  The invention according to claim 2 is the mass damper test apparatus 41 according to claim 1, wherein the test target mass damper 21 includes a main body (inner cylinder 22) and a movable part (screw shaft 23a) movable relative to the main body. ) And a transmission mechanism (ball screw 23) for transmitting the displacement of the movable part relative to the main body part to the rotary mass 24 in a state where the displacement is converted into a rotational motion, and the excitation force of the actuator 6 controlled by the control means is Based on the damper resistance force P detected when input to the vibration system S, based on the vibration equation of the second virtual vibration system VS2, the excitation force of the actuator 6 controlled by the control means is the second virtual vibration. Damper predicted resistance force for calculating the resistance force of the ideal mass damper 21I predicted to be generated by being input to the system VS2 as a damper predicted resistance force Ppr that is a predicted value of the damper resistance force P An output means (control device 15, steps 52, 3-6, steps 53, 54 in FIG. 10) and setting means (control device 15, step 21 in FIG. 6) for setting the rigidity kb1 of the virtual adjustment member MVA. Further, the virtual displacement acquisition means calculates a predicted value of the displacement of the virtual adjustment member according to the calculated damper predicted resistance force Ppr and the set rigidity kb1 of the virtual adjustment member (step 55 in FIG. 10). It is characterized by.

  According to this configuration, the test target mass damper includes a main body part, a movable part movable with respect to the main body part, and a transmission mechanism that transmits the displacement of the movable part with respect to the main body part to the rotary mass in a state where the displacement is converted into a rotational motion. have. The actuator is controlled by the controller based on the vibration equation of the second virtual vibration system according to the damper resistance detected when the excitation force of the actuator controlled by the controller is input to the vibration system. The resistance force of the ideal mass damper that is predicted to be generated when the excitation force of the actuator is input to the second virtual vibration system is assumed to be a predicted damper resistance value that is a predicted value of the damper resistance force by the predicted damper resistance calculation means. Calculated. In other words, this damper predicted resistance force is controlled by the control means in the second virtual vibration system when it is assumed that the test target mass damper is provided in the second virtual vibration system instead of the ideal mass damper. This corresponds to a predicted value of the damper resistance force that is predicted to be generated by inputting the excitation force of the actuator.

  Further, the rigidity of the virtual adjustment member is set by the setting means, and the predicted value of the displacement of the virtual adjustment member is calculated according to the calculated damper predicted resistance force and the set rigidity of the virtual adjustment member. As described above, since the virtual adjustment member is connected in series with the first virtual vibration system, the predicted value of the displacement of the virtual adjustment member according to the damper predicted resistance force and the rigidity of the virtual adjustment member as described above. By calculating, the calculation can be performed appropriately. In this case, the damper predicted resistance force according to the damper resistance force of the mass damper to be tested detected during the input of the excitation force of the actuator controlled by the control means is used to calculate the predicted value of the displacement of the virtual adjustment member. Since it uses, the predicted value of the displacement of the virtual adjustment member commensurate with the test target mass damper can be calculated appropriately.

  The invention according to claim 3 is the mass damper test device 41 according to claim 2, wherein the movable portion acceleration detecting means (control device 15, step of FIG. 11) for detecting the acceleration of the movable portion relative to the main body portion (damper acceleration ad). 61), and the damper predicted resistance calculation means further includes a damper that further responds to the acceleration of the movable part detected when the excitation force of the actuator 6 controlled by the control means is input to the vibration system S. The predicted resistance force Ppr is calculated (steps 53 and 54, equation (22)).

  According to this configuration, the acceleration of the movable part with respect to the main body is detected by the movable part acceleration detecting means. Further, the damper predicted resistance force is calculated further according to the acceleration of the movable part detected when the excitation force of the actuator controlled by the control means is input. Since the acceleration of the movable part relative to the main body of the test target mass damper has a close correlation with the predicted damper resistance of the test target mass damper, the predicted damper resistance can be calculated more appropriately.

BRIEF DESCRIPTION OF THE DRAWINGS It is a figure which shows schematically the test apparatus by 1st Embodiment of this invention with the test object mass damper to which this is applied. It is a block diagram which shows the control apparatus of a test apparatus, etc. (A) It is a model figure which shows an ideal mass damper, a virtual adjustment member, a 1st virtual vibration system, and a 2nd virtual vibration system, (b) It is a model figure which shows the rotation mass and viscous body of a whole spring element, an ideal mass damper. . It is sectional drawing of a test object mass damper. It is a flowchart which shows the vibration control process performed by the control apparatus of FIG. It is a flowchart which shows the rigidity setting process performed by the control apparatus of FIG. It is a flowchart which shows the evaluation process performed by the control apparatus of FIG. It is a flowchart which shows the modification of an excitation control process. It is a block diagram which shows the control apparatus etc. of the test apparatus by 2nd Embodiment of this invention. It is a flowchart which shows the vibration control process performed by the control apparatus of FIG. It is a flowchart which shows the continuation of FIG.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. FIG. 1 schematically shows a test apparatus 1 according to a first embodiment of the present invention, together with a test target mass damper 21 which is a mass damper to which the test apparatus 1 is applied. This test target mass damper 21 constitutes an additional vibration system together with a rigid transmission member (not shown), and suppresses vibration of a structure such as a building. The configuration is similar to the mass damper described in FIG. First, the configuration and operation of the test target mass damper 21 will be briefly described.

  As shown in FIGS. 1 and 4, the test target mass damper 21 includes an inner cylinder 22, a ball screw 23, a rotating mass 24, and a limiting mechanism 25. The inner cylinder 22 is made of a cylindrical steel material. One end of the inner cylinder 22 is open, and the other end is attached to the first flange 26 via a universal joint.

  The ball screw 23 includes a screw shaft 23a and a nut 23c that is rotatably engaged with the screw shaft 23a via a large number of balls 23b. One end of the screw shaft 23a is accommodated in the opening of the inner cylinder 22 described above, and the other end of the screw shaft 23a is attached to the second flange 27 via a universal joint. Further, the nut 23 c is rotatably supported by the inner cylinder 22 via a bearing 28. In FIG. 1, for convenience, the symbol of the screw shaft 23a is omitted. The screw shaft 23a is in a predetermined neutral position shown in FIG. 4 when an excitation force from an actuator 6 described later is not input.

  The rotary mass 24 is made of a material having a large specific gravity, for example, iron, and is formed in a cylindrical shape. The rotating mass 24 covers the inner cylinder 22 and the ball screw 23, and is rotatably supported by the inner cylinder 22 via a bearing 29. A pair of ring-shaped seals 30 and 30 are provided between the rotary mass 24 and the inner cylinder 22. A space formed by the seals 30, 30, the rotation mass 24 and the inner cylinder 22 is filled with a viscous body 31 made of silicon oil.

  In the test target mass damper 21 configured as described above, when a relative displacement occurs between the inner cylinder 22 and the screw shaft 23 a, the limiting mechanism 25 is operated in a state where the relative displacement is converted into a rotational motion by the ball screw 23. The rotation mass 24 is rotated by being transmitted to the rotation mass 24 through the rotation mass 24.

  The limiting mechanism 25 includes a ring-shaped rotary sliding member 25a, a plurality of screws 25b, and springs 25c (only two are shown). The rotating mass 24 rotates integrally with the nut 23c until the load acting in the axial direction of the test target mass damper 21 (hereinafter referred to as “axial load”) reaches a limit load determined according to the tightening degree of the screw 25b. . On the other hand, when the axial load of the test target mass damper 21 reaches the limit load, slippage occurs between the rotary sliding member 25 a and the nut 23 c or the rotary mass 24.

  Next, the test apparatus 1 will be described with reference to FIGS. 1 and 2. In the following description, for the sake of convenience, the upper side and the lower side in FIG. 1 are respectively referred to as “upper” and “lower”, the left side and the right side are “left” and “right”, respectively, and the front side and the rear side are “front” and “ "After". As shown in FIG. 1, the test apparatus 1 inputs an excitation force to a vibration system S (described later) including upper, lower, left and right frames 2, 3, 4, 5 and a test target mass damper 21 that are integrally provided in a cross beam shape. And a guide member 8 for supporting the connecting member 7 on the lower frame 3 so as to be movable in the left-right direction. The connecting member 7, the guide mechanism 8, and the test target mass damper 21 are arranged in a space defined by upper, lower, left and right frames 2 to 5.

  The actuator 6 is composed of, for example, a solenoid, and includes a main body 6a attached to the left frame 4 and a plunger 6b connected to a connecting member 7 via a load cell 11 described later. The actuator 6 is controlled by a control device 15 (see FIG. 2), which will be described later, whereby an excitation force is output from the plunger 6b.

  Each of the frames 2 to 5 and the connecting member 7 are made of steel. The first flange 26 of the test target mass damper 21 described above is connected to the right frame 5 via a connecting member 9 made of steel, and the axis of the test target mass damper 21 extends in the left-right direction. The connecting member 9 may be omitted and the first flange 26 may be directly connected to the right frame 5. Further, the connecting member 7 is disposed at the center in the left-right direction of the lower frame 3, and the second flange 27 of the test target mass damper 21 is attached to the right surface of the connecting member 7, that is, the surface opposite to the actuator 6. It has been.

  The guide mechanism 8 is attached to the center of the upper surface of the lower frame 3 in the left-right direction, and extends in the left-right direction. The slide member 8b engages with the rail 8a via a plurality of balls (not shown). have. The slide member 8b can move only in the left-right direction with respect to the rail 8a and cannot rotate. A connecting member 7 is attached to the upper surface of the slide member 8b.

  The test apparatus 1 further includes a load cell 11 provided between the connecting member 7 and the actuator 6, a first displacement sensor 12 attached to the right surface of the connecting member 7, and the control device 15. The load cell 11 is of a strain gauge type, for example, and detects a load acting on the connecting member 7 as a resistance force P (hereinafter referred to as “damper resistance force”) P of the test target mass damper 21, and the detection signal is transmitted to the control device 15. Output to. The load cell 11 may be a capacitance type or other appropriate type. Further, the connecting member 7 may be omitted, and the plunger 6 b of the actuator 6 may be connected to the second flange 27 via the load cell 11.

  The first displacement sensor 12 is, for example, a laser type sensor that detects the displacement xd of the screw shaft 23a with respect to the inner cylinder 22 of the test-target mass damper 21 (hereinafter referred to as “damper displacement”) and sends the detection signal to the control device 15. Output. The first displacement sensor 12 may be a contact type or other appropriate type. The control device 15 is configured by a combination of a power source for driving the actuator 6, a rectifier, a CPU, a RAM, a ROM, an I / O interface, and the like.

  In the test apparatus 1 having the above configuration, an excitation force (for example, a sinusoidal excitation force) is output from the actuator 6, and the excitation force of the actuator 6 (hereinafter referred to as “actuator excitation force”) is coupled. A vibration system S including the member 7, the test target mass damper 21, the connecting member 9, and the right frame 5 is input. The controller 15 calculates various performance parameters representing the performance of the test target mass damper 21 according to the damper resistance force P and the damper displacement xd detected during the input of the actuator excitation force.

  In this case, as described in the problem and the means for solving the problem of the present invention, the natural vibration of the vibration system S determined by the overall rigidity of the right frame 5 and the connecting member 7 and the rotational inertial mass of the mass damper 21 to be tested. When the number is considerably higher than the vibration frequency of the actuator excitation force, the component corresponding to the natural frequency of the vibration system S among the vibration frequency components included in the damper resistance force P detected by the load cell 11 increases. As a result, irregular undulations appear in the detected damper resistance force P, so that the performance parameter of the test target mass damper 21 cannot be appropriately calculated based on the damper resistance force P.

  Therefore, in the present embodiment, the actuator excitation force is controlled by the control device 15 as described below in order to appropriately calculate the performance parameter. First, the technical viewpoint of the control method will be described.

  When it is assumed that the above-described virtual adjustment member MVA and ideal mass damper 21I of the present invention are connected in series to the right frame 5 and the connecting members 7 and 9, the entire right frame 5 and the connecting members 7 and 9 are supporting members. Assuming MS, a model diagram showing the entirety of the virtual adjustment member MVA, the ideal mass damper 21I and the support member MS is expressed as shown in FIG. 3A, for example. The ideal mass damper 21I corresponds to the test target mass damper 21 in which each of the specifications such as the rotational inertia mass of the rotary mass 24 and the damping coefficient of the viscous body 31 is a predetermined ideal value described in the catalog. In FIG. 3A, MDS is a spring element (hereinafter referred to as “ideal damper spring element”) including an inner cylinder of the ideal mass damper 21I and a ball screw. Further, the virtual adjustment member MVA is for bringing the natural frequency of the first virtual vibration system VS1 in which the support member MS and the ideal mass damper 21I are connected in series to a predetermined frequency described later.

  Further, assuming that the entire virtual adjustment member MVA, the support member MS, and the ideal damper spring element MDS are the whole spring element MAS, a model diagram showing the whole spring element MAS, the rotation mass 24I of the ideal mass damper 21I and the viscous body 31I is, for example, FIG. It is expressed as (b). Hereinafter, the rotating mass 24I and the viscous body 31I of the ideal mass damper 21I are collectively referred to as “ideal damper elements”.

  Here, assuming that the rigidity (spring constant) of the virtual adjustment member MVA is kb1, the rigidity of the ideal damper spring element MDS is kb2, the rigidity of the support member MS is kb3, and the rigidity of the overall spring element MAS is kb, kb is kb1 · It is expressed as kb2 · kb3 / (kb1 · kb2 + kb1 · kb3 + kb2 · kb3). Here, an input displacement (hereinafter simply referred to as “input displacement”) due to an actuator excitation force input to the overall spring element MAS and the ideal damper element is assumed to be x, and a displacement of the overall spring element MAS due to the actuator excitation force (hereinafter “ X = xdi + xb is established, where xb is the “total spring element displacement”. This xdi is the displacement of the screw shaft relative to the inner cylinder of the ideal mass damper 21I (hereinafter referred to as “ideal damper displacement”).

Further, a change speed of the ideal damper displacement xdi (a differential value of xdi; hereinafter referred to as “ideal damper speed”) is represented by vdi, and a nonlinear viscous damping coefficient cdi · (vdi) due to the shear resistance of the viscous body 31I is represented by | vdi | Assuming that it is proportional to the power of i.e., cdi · (vdi) = cv · | vdi | ^α-1 , the vibration when the actuator excitation force is input to the whole spring element MAS and the ideal damper element The equation is expressed by the following equation (1).
PI = cv · | vdi | ^α−1 · vdi + mdi · adi = kb (x−xdi)
...... (1)
Here, PI is the resistance force of the ideal mass damper 21I (hereinafter referred to as “ideal damper resistance force”), and mdi is the rotational inertial mass (equivalent mass) of the rotary mass 24I. Further, adi is the change acceleration of the ideal damper displacement xdi (two-time differential value of xdi; hereinafter referred to as “ideal damper acceleration”), and the other parameters are as described above.

  Next, in order to use the vibration equation represented by the equation (1) in the time history response analysis, the change amount per unit time of the ideal damper resistance force PI (hereinafter referred to as “ideal damper resistance force change amount”) ΔPI It is expressed by the function of

First, the derivative of the viscous resistance force cv · sgn (vdi) | vdi | ^α (= cv · | vdi | ^α-1 · vdi) of the ideal mass damper 21I is the ideal damper speed in relation to the velocity and the viscous resistance force. A tangential attenuation coefficient cdt (vdi) with respect to vdi, which is expressed by the following equation (2). The tangential mass related to the inertial force is always the rotational inertial mass mdi of the rotary mass 24I.
cdt (vdi) = cv · α · | vdi | ^α-1 (2)

Using the tangential damping coefficient cdt (vdi) and the rotational inertia mass mdi of the rotational mass 24I, the ideal damper resistance force change ΔPI is expressed by the following equation (3).
ΔPI = cdt (vdi) · Δvdi + mdi · Δadi
= Cv · α · | vdi | ^α-1 · Δvdi + mdi · Δadi (3)
Here, Δvdi is a change amount of the ideal damper speed vdi per step in the time history response analysis (hereinafter referred to as “ideal damper speed change amount”), and Δadi is a change amount of the ideal damper acceleration adi per step. (Hereinafter referred to as “ideal damper acceleration change amount”).

The ideal damper speed change amount Δvdi is expressed by the following equation (4) using the average acceleration method (β = 0.25).
Δvdi = (Δadi + 2 · adiz) Δt / 2 (4)
Here, Δt is the time per step in the time history response analysis (time between each step; hereinafter referred to as “step time”), and adiz is the previous value of the ideal damper acceleration adi, that is, in the previous step. The ideal damper acceleration adi.

Further, when this equation (4) is transformed, the ideal damper acceleration change amount Δadi is expressed by the following equation (5).
Δadi = (2 · Δvdi / Δt) −2 · adiz (5)

The following formula (6) is obtained from the above formulas (3) to (5).
ΔPI = cr · Δvdi−2 · mdi · adiz (6)

The variable cr in the equation (6) is expressed by the following equation (7).
cr = cdt (vdi) + 2 · mdi / Δt
= Cv · α · | vdi | ^α-1 + 2 · mdi / Δt (7)

Further, the ideal damper resistance change amount ΔPI is obtained by using the rigidity kb of the overall spring element MAS and the change amount of the overall spring element displacement xb per step (hereinafter referred to as “total spring element displacement change amount”) Δxb. It is represented by the following formula (8).
ΔPI = kb · Δxb (8)

Further, the total spring element displacement change amount Δxb is expressed by the following equation (9) using an average acceleration method (β = 0.25). Here, Δvb is a change speed of the overall spring element displacement change amount Δxb (a single differential value of Δxb, hereinafter referred to as “total spring element speed change amount”). In other words, the total spring element speed change amount Δvb is a change amount per step of the change speed of the total spring element displacement xb (a single differential value of xb, hereinafter referred to as “total spring element speed vb”). Further, vbz is the previous value of the overall spring element speed vb, that is, the overall spring element speed vb in the previous step.
Δxb = (Δvb + 2 · vbz) Δt / 2 (9)

Further, when this equation (9) is modified, the total spring element speed change amount Δvb is expressed by the following equation (10).
Δvb = (2 · Δxb / Δt) −2 · vbz (10)

From the formulas (6) to (10), the ideal damper resistance change amount ΔPI is represented by the following formula (11). In Expression (11), Δv is a change amount per step (hereinafter referred to as “input speed change amount”) of a change speed of the input displacement x (a single differential value of x; hereinafter referred to as “input speed v”). .
ΔPI = cr (Δv−Δvb) −2 · mdi · adiz
= (Cr · Δv / κ) + (2 · cr · vbz / κ)
-(2 · mdi · adiz / κ) (11)
Here, κ = 1 + {(2 · cr) / (Δt · kb)}.

The overall spring element speed vb, the ideal damper acceleration adi, and the ideal damper speed vdi are expressed by the following equations (12), (13), and (14), respectively.
vb = vbz + Δvb = (2 · Δxb / Δt) −vbz (12)
adi = a−ab
= A − {(4 · Δxb / Δt ² ) − (4 · vbz / Δt) −abz}
(13)
vdi = v−vb = v − {(2 · Δxb / Δt) −vbz}
= V − {(2 · ΔPI) / (Δt · kb) −vbz} (14)
Here, a is the change acceleration of the input displacement x (double differential value of x; hereinafter referred to as “input acceleration”), and ab is the change acceleration of the overall spring element displacement xb (double differential value of xb. It is called “overall spring element acceleration”). Abz is the previous value of the overall spring element acceleration ab, that is, the overall spring element acceleration ab in the previous step.

Further, the predicted value (hereinafter referred to as “ideal damper predicted resistance force”) PIpr of the ideal damper resistance force PI this time is expressed by the following equation (15).
PIpr = PIprz + ΔPI (15)
Here, PIprz is the previous value of the ideal damper predicted resistance force PIpr, that is, the ideal damper predicted resistance force PIpr in the previous step.

Further, the predicted displacement value (hereinafter referred to as “virtual adjustment member predicted displacement”) xb1pr of the virtual adjustment member MVA this time is expressed by the following equation (16). As described above, kb1 is the rigidity of the virtual adjustment member MVA.
xb1pr = PIpr / kb1 (16)

  The control device 15 applies a displacement equivalent to the input displacement input to the virtual vibration system (FIG. 3A) in which the virtual adjustment member MVA is connected in series to the vibration system S including the test target mass damper 21 to the actuator 6. In order to input to the vibration system S, the target value xobj of the input displacement due to the actuator excitation force is calculated according to the basic value xbase and the virtual adjustment member predicted displacement xb1pr, and based on the calculated target value xobj The actuator excitation force is controlled.

  Specifically, the control device 15 executes an excitation control process for controlling the actuator 6 shown in FIG. This process is repeatedly executed every predetermined time (for example, 10 msec) as the step time Δt, and is completed (stopped) when a predetermined execution time (for example, 60 sec) elapses from the start. First, in step 1 of FIG. 5 (illustrated as “S1”, the same applies hereinafter), various parameters obtained at that time are shifted as their previous values. These parameters include a basic value xbase, an input speed v, an ideal damper predicted resistance force PIpr, an overall spring element speed vb, an input acceleration a, an ideal damper acceleration adi, and an ideal damper speed vdi calculated as described later. included.

Next, the variable cr is calculated by the equation (7) (cr = cv · α · | vdi | ^α-1 + 2 · mdi / Δt) (step 2). In this case, as the variables cv and α in the equation (7), predetermined values obtained in advance by experiments in accordance with predetermined ideal values of the specifications of the test target mass damper 21 described in the specification are used. In addition, a predetermined ideal value described in the specification is used as the rotational inertia mass mdi of the rotary mass 24I, and the predetermined time is used as the step time Δt. This also applies to the calculation of other parameters described later. Further, the previous value vdiz of the ideal damper speed shifted in step 1 is used as the ideal damper speed vdi. Note that at the first time of this process, the ideal damper speed vdi is set to a value of 0 in the calculation.

  Next, the counter value C of the counter for counting the elapsed time from the start of this process is incremented (step 3). Next, a basic value xbase of the current input displacement from the actuator 6 is calculated by searching a predetermined map (not shown) according to the counter value C (step 4). This map is formed by mapping an input displacement based on an earthquake wave (for example, a sine vibration wave) having a predetermined frequency fref as a basic value xbase in association with the counter value C.

Next, the current input speed v is calculated by the following equation (17) using the calculated basic value xbase or the like (step 5).
v = {2 (xbase−xbasez) / Δt} −vz (17)
Here, xbasez and vz are the previous values of the basic value xbase and the input speed v shifted in the step 1, respectively. In this calculation, xbasez and vz are set to 0 at the first time of this process.

  Next, an input speed change amount Δv is calculated by subtracting the previous value vz from the calculated input speed v (step 6).

  Next, using the calculated input speed change amount Δv, the previous value vbz of the overall spring element speed shifted in step 1 and the previous value adiz of the ideal damper acceleration, and the variable cr calculated in step 2, The ideal damper resistance change amount ΔPI is calculated by the equation (11) (ΔPI = (cr · Δv / κ) + (2 · cr · vbz / κ) − (2 · mdi · adiz / κ)) (step 7). ).

In the calculation of ΔPI, the previous value vbz of the overall spring element speed and the previous value adiz of the ideal damper acceleration are set to the value 0 at the first time of this process. Further, as is apparent from FIG. 3 described above, the rigidity kb of the overall spring element MAS at κ = 1 + {(2 · cr) / (Δt · kb)} Is set.
kb = kb1, kb2, kb3
/ (Kb1 / kb2 + kb1 / kb3 + kb2 / kb3) (18)

  In this case, as the rigidity kb2 of the ideal damper spring element MDS (the spring element comprising the inner cylinder and ball screw of the ideal mass damper 21I), a predetermined value based on the specifications described in the specifications of the test target mass damper 21 is used, and the support member A predetermined value obtained in advance based on the specifications of each member is used as the stiffness kb3 of the MS (the entire right frame 5 and the connecting members 7 and 9). The virtual adjustment member MVA is for bringing the natural frequency of the first virtual vibration system VS1 closer to the predetermined frequency fref. Therefore, the rigidity kb1 of the virtual adjustment member MVA is equal to the natural frequency fvs2 of the second virtual vibration system VS2 in which the support member MS, the ideal mass damper 21I, and the virtual adjustment member MVA shown in FIG. Is set in advance so as to be the predetermined frequency fref of the input excitation waveform set as the basic value xbase. In this case, the natural frequency fvs2 of the second virtual vibration system VS2 is represented by fvs2 = sqrt {kb / mdi} / (2π).

  FIG. 6 shows a stiffness setting process for setting the stiffness kb1 of the virtual adjustment member MVA and the stiffness kb of the overall spring element MAS. This process is performed only once just before the start of the vibration control process. Executed. In the stiffness setting process, in steps 21 and 22, the stiffness kb1 of the virtual adjustment member MVA and the stiffness kb of the overall spring element MAS are set as described above. In this setting, the rigidity kb2 of the ideal damper spring element MDS, the rigidity kb3 of the support member MS, and the predetermined frequency fref are input to the control device 15 by the operator.

  Returning to FIG. 5, in step 8 following step 7, the ideal damper resistance change amount ΔPI calculated in step 7 and the previous value PIprz of the ideal damper predicted resistance shifted in step 1 are used. 15) The current ideal damper predicted resistance force PIpr is calculated by (PIpr = PIprz + ΔPI). Next, using the calculated ideal damper predicted resistance force PIpr, the virtual adjustment member predicted displacement xb1pr is calculated by the above equation (16) (xb1pr = PIpr / kb1) (step 9). Also in this case, the rigidity kb1 of the virtual adjustment member MVA set as described above is used.

  Next, the target value xobj of the input displacement is calculated by subtracting the virtual adjustment member predicted displacement xb1pr calculated in step 9 from the basic value xbase calculated in step 4 (step 10). Next, a control signal based on the calculated target value xobj is output to the actuator 6 (step 11). Thus, the actuator excitation force is controlled so that the input displacement input from the actuator 6 to the vibration system S including the test target mass damper 21, the right frame 5, and the connecting members 7 and 9 becomes the target value xobj.

  Next, by dividing the ideal damper resistance change ΔPI calculated in step 7 by the stiffness kb of the overall spring element MAS (see the above equation (8)), the overall spring element displacement change Δxb is calculated (see FIG. 8). Step 12). As is apparent from this calculation method, the total spring element displacement change amount Δxb corresponds to the total spring element displacement change amount obtained by inputting the actuator excitation force by the current control.

Next, using the calculated total spring element displacement change amount Δxb, the step time Δt, and the previous value vbz of the total spring element speed shifted in step 1, the equation (12) (vb = (2 · Δxb / The current overall spring element speed vb is calculated by [Delta] t) -vbz) (step 13). Next, using the input speed v calculated in Step 5, the previous value vz of the input speed v shifted in Step 1, the previous value az of the input acceleration a, and the step time Δt, the following equation (19) Thus, the current input acceleration a is calculated (step 14).
a = {2 (v−vz) / Δt} −az (19)

Subsequently, the calculated input acceleration a, the total spring element displacement change amount Δxb, the step time Δt, the previous value vbz of the total spring element speed, and the previous value abz of the total spring element acceleration ab shifted in the step 1 are calculated. The ideal damper acceleration adi of this time is calculated by the above equation (13) (adi = a − {(4 · Δxb / Δt ² ) − (4 · vbz / Δt) −abz})) (step 15). Next, the input speed v calculated in step 5, the ideal damper resistance change amount ΔPI calculated in step 7, the step time Δt, the rigidity kb of the overall spring element MAS set in step 22 of FIG. 6, Using the previous value vbz of the overall spring element speed shifted in step 1, the ideal damper speed vdi is obtained by the above equation (14) (vdi = v − {(2 · ΔPI) / (Δt · kb) −vbz}). Calculation is made (step 16), and this process is terminated.

  As is clear from the above calculation method, the overall spring element speed vb, the ideal damper acceleration adi, and the ideal damper speed vdi calculated in steps 13, 15 and 16 are the same as the total spring element displacement change amount Δxb. This corresponds to the overall spring element speed vb, the ideal damper acceleration adi, and the ideal damper speed vdi obtained by inputting the actuator excitation force by the control.

  As described above, in the vibration control process, the overall spring element speed vb, the ideal damper acceleration adi, and the ideal damper speed vdi calculated in Steps 13, 15, and 16 are the previous values when the next main process is executed. Are used to calculate the variable cr and the ideal damper resistance change amount ΔPI.

  Next, a process for evaluating the test target mass damper 21 executed by the control device 15 will be described with reference to FIG. This process is repeatedly executed every predetermined time following the execution of the vibration control process described above.

  First, in step 31 of FIG. 7, the maximum speed resistance force Qv is calculated as follows. That is, during input of the actuator excitation force controlled by the excitation control process, when the screw shaft 23a is in the neutral position (see FIG. 4) according to the damper resistance force P and the damper displacement xd, that is, the inner cylinder The damper resistance force P detected when the speed of the screw shaft 23a with respect to 22 is maximum is sampled as the maximum speed resistance force Qv. In this case, as apparent from the reciprocating motion of the screw shaft 23a in the left-right direction due to the actuator excitation force, the maximum speed resistance force Qv is a positive maximum speed resistance force + Qv and a negative maximum speed resistance force. -Qv is obtained. Therefore, the average value of the absolute values of each of + Qv and −Qv is calculated as the maximum speed resistance force Qv of the test target mass damper 21.

In step 32 following step 31, the attenuation coefficient Ceq of the test target mass damper 21 is calculated as follows. That is, first, during the input of the actuator excitation force controlled by the excitation control process, the damper resistance force P and the damper displacement xd are sampled in association with each other at the predetermined time, and the sampling is performed on the screw shaft 23a. Is reciprocated for one cycle, that is, until the screw shaft 23a located at the neutral position reciprocates left and right and returns to the neutral position again. Next, the history area ΔW is calculated by the following equation (20) using the plurality of damper resistance forces P and the damper displacement xd sampled. Here, i is the sampling number of the sampled damper resistance force P and damper displacement xd, and n is the number of sampled damper resistance force P and damper displacement xd.

Next, the maximum one of the plurality of sampled damper displacements xd is set as the maximum displacement xdmax, and the calculated history area ΔW, the maximum displacement xdmax and the predetermined vibration frequency fref of the actuator excitation force are used. The attenuation coefficient Ceq of the test target mass damper 21 is calculated by the following equation (21).
Ceq = ΔW / (2π ² · fref · xdmax ² ) (21)

  In step 33 following step 32, the rotational inertia mass md of the rotary mass 24 of the test target mass damper 21 is calculated as follows. That is, first, as in the calculation of the damping coefficient Ceq, during the input of the actuator excitation force controlled by the excitation control process, the damper resistance force P and the damper displacement xd are sampled in association with each other every predetermined time. At the same time, the sampling is performed until the screw shaft 23a reciprocates for one cycle. Next, among the plurality of sampled damper resistances P, when the damper displacement xd is maximized (when the test target mass damper 21 is most expanded) or when it is minimum (the test target mass damper 21 is contracted most). P) sampled at the time) is set as the maximum inertial force Qmax.

  Next, by differentiating the plurality of sampled damper displacements xd twice, the acceleration of the screw shaft 23a relative to the inner cylinder 22 at that time is calculated, and the largest one of the calculated accelerations is Set as the maximum acceleration δmax. Next, the rotary inertia mass md of the test target mass damper 21 is calculated by dividing the set maximum inertia force Qmax by the set maximum acceleration δmax.

  In step 34 following step 33, the performance of the test target mass damper 21 is evaluated as follows, and this process is terminated. That is, the condition that the maximum speed resistance force Qv calculated in step 31 is smaller than a predetermined reference resistance force, the condition that the attenuation coefficient Ceq calculated in step 32 is smaller than a predetermined reference attenuation coefficient, and step 33. When at least one of the conditions that the rotational inertia mass md calculated in (1) is smaller than the predetermined reference mass is satisfied, it is evaluated that the performance of the test target mass damper 21 is low. Hereinafter, the maximum speed resistance force Qv, the damping coefficient Ceq, and the rotational inertia mass md are collectively referred to as “performance parameters” as appropriate.

  As described above, according to the first embodiment, the support member MS including the rigid right frame 5 and the connecting members 7 and 9 is connected to the test target mass damper 21 in series, and the test target mass damper 21 and the support member MS are supported. An actuator excitation force is input to the vibration system S including the member MS. Further, the virtual adjustment member predicted displacement xb1pr that is a predicted value of the displacement of the virtual adjustment member MVA when it is assumed that the excitation force of the predetermined frequency fref is input to the second virtual vibration system VS2 is the second virtual vibration system VS2. Is calculated based on the equation (11) derived from the vibration equation (the above equation (1)) (steps 2 to 9 in FIG. 5). The second virtual vibration system VS2 includes a virtual adjustment member MVA for bringing the natural frequency of the first virtual vibration system VS1 in which the ideal mass damper 21I and the support member MS are connected in series to a predetermined frequency fref. This is a vibration system connected in series to one virtual vibration system VS1.

  Further, the actuator excitation force is controlled according to the input displacement basic value xbase for inputting the excitation force of the predetermined frequency fref and the calculated virtual adjustment member predicted displacement xb1pr (steps 10 and 11). ). By controlling the actuator excitation force as described above, a predetermined resistance is applied to the vibration system S in which the virtual adjustment member MVA is connected in series as the damper resistance force P of the test target mass damper 21 during the input of the actuator excitation force. It is possible to obtain a resistance force equivalent to the damper resistance force P of the test target mass damper 21 when it is assumed that the actuator excitation force having the frequency fref is input.

  According to the first embodiment, according to the damper resistance force P of the test target mass damper 21 detected when the actuator excitation force controlled as described above is input to the vibration system S, the test target mass damper 21. The performance parameter representing the performance is calculated (steps 31 to 33 in FIG. 7). As described above, among the frequency components included in the detected damper resistance force P, the component corresponding to the frequency of the actuator excitation force can be increased, so that the swell as described above can be suppressed. Calculation of the performance parameter according to the resistance force P can be performed appropriately.

  Further, the virtual adjustment member MVA is a virtual member, and is not actually provided in the test apparatus 1 as shown in FIG. 1, and the adjustment member of the conventional test apparatus described above is unnecessary. Accordingly, it is possible to relatively easily perform the performance test confirmation and to simplify the configuration.

  Next, an excitation control process according to a modification of the first embodiment will be described with reference to FIG. This modification is different from the first embodiment only in the method of calculating the virtual adjustment member predicted displacement xb1pr. In FIG. 8, the same step number is attached | subjected about the same execution content as FIG. As shown in FIG. 8, in the modified example, the processing by steps 1, 2, 5-8, and 12-16 is omitted, and when the basic value xbase is calculated in step 4, The virtual adjustment member predicted displacement xb1pr is calculated by searching a predetermined map (not shown) according to the counter value C incremented in the step 3. This map is obtained by mapping the virtual adjustment member predicted displacement xb1pr calculated in advance by the equations (7) and (11) to (18) in association with the counter value C.

  Subsequent to Step 41, Steps 10 and 11 are executed, and the present process is terminated. Thus, the target value xobj of the input displacement is calculated by subtracting the virtual adjustment member predicted displacement xb1pr from the basic value xbase of the input displacement, and a control signal based on the calculated target value xobj is output to the actuator 6. The

  As described above, according to the modified example described above, the effects of the first embodiment described above can be obtained in the same manner, and the processes in steps 1, 2, 5-8, and 12-16 are omitted. Therefore, the calculation load of the control device 15 can be reduced.

  Next, the test apparatus 41 according to the second embodiment of the present invention will be described with reference to FIGS. 9 to 11 focusing on differences from the first embodiment. As shown in FIG. 9, the test apparatus 41 further includes a second displacement sensor 13. The second displacement sensor 13 is, for example, a laser type similar to the first displacement sensor 12 and is provided in the right frame 5, and the displacement of the right frame 5 due to vibration is the displacement of the support member MS (hereinafter “support”). Xb3 ”(referred to as“ member displacement ”), and the detection signal is output to the control device 15. The second displacement sensor 13 may be a contact type or other appropriate type.

  FIG. 10 shows an excitation control process executed by the control device 15 of the test apparatus 41. In the figure, the same step numbers are assigned to the same execution contents as those of the vibration control process (FIG. 5) according to the first embodiment. Hereinafter, this process will be described with a focus on differences from the first embodiment.

  First, in step 51 of FIG. 10, as in step 1, the various parameters obtained at that time are shifted as their previous values. These parameters include a basic value xbase, input speed v, damper predicted resistance force Ppr, overall spring element displacement xb, overall spring element speed vb, damper speed vd, and damper acceleration ad calculated as described below. It is.

Next, using the previous value vdz of the damper speed shifted in step 51 as the ideal damper speed vdi, the variable cr is calculated by the above equation (7) (step 52). That is, the calculation formula of the variable cr in this case is cr = cv · α · | vdz | ^α-1 + 2 · mdi / Δt. Similarly to the first embodiment, as the variables cv and α in this equation, predetermined values obtained in advance by experiments in accordance with predetermined ideal values of specifications of the test target mass damper 21 described in the specifications are used. Further, the predetermined ideal value described in the specification is used as the rotational inertia mass mdi of the rotary mass 24I, and the predetermined time is used as the step time Δt.

Following Step 52, Steps 3 to 6 are executed, and in Step 53 following Step 6, a damper resistance change amount ΔP is calculated. This damper resistance change amount ΔP is a predicted value of the current change amount per predetermined time (that is, step time Δt) of the damper resistance force P of the test target mass damper 21. The previous value vbz of the shifted overall spring element speed, the previous value adz of the damper acceleration, and the variable cr calculated in step 52 are used to calculate the following equation (22) obtained by modifying the equation (11). Of course, cr calculated in step 52 is used as the variable cr included in the variable κ.
ΔP = (cr · Δv / κ) + (2 · cr · vbz / κ)
-(2 · mdi · adz / κ) (22)

  Next, the damper predicted resistance force Ppr is calculated by adding the calculated damper resistance force change amount ΔP to the detected damper resistance force P (step 54, Ppr = P + ΔP). The damper predicted resistance force Ppr is a predicted value of the current damper resistance force P of the test target mass damper 21. Next, a virtual adjustment member predicted displacement xb1pr is calculated by dividing the calculated damper predicted resistance force Ppr by the rigidity kb1 of the virtual adjustment member MVA (step 55, xb1pr = Ppr / kb1). The rigidity kb1 of the virtual adjustment member MVA is set in the same manner as in the first embodiment (FIG. 6).

  Next, Step 10 is executed, and the target value xobj is calculated by subtracting the virtual adjustment member predicted displacement xb1pr calculated in Step 45 from the basic value xbase calculated in Step 4. Next, step 11 is executed, and a control signal based on the calculated target value xobj is output to the actuator 6.

  In step 56 in FIG. 11 following step 11, the damper spring element predicted displacement xb2pr is calculated by dividing the damper predicted resistance force Ppr calculated in step 54 by the rigidity kb2 of the ideal damper spring element MDS. This damper spring element predicted displacement xb2pr is the current predicted value of the displacement of the spring element composed of the inner cylinder 22 and the ball screw 23 of the test target mass damper 21.

  Next, the total spring element displacement xb is calculated by adding the virtual adjustment member predicted displacement xb1pr and damper spring element predicted displacement xb2pr calculated in Steps 55 and 56, respectively, and the detected support member displacement xb3 to each other. (Step 57). Next, the current total spring element displacement change amount Δxb is calculated by subtracting the previous value xbz of the total spring element displacement shifted in step 51 from the calculated current total spring element displacement xb (step 51). 58).

Next, using the calculated total spring element displacement change amount Δxb, the step time Δt, and the previous value vbz of the total spring element speed shifted in step 51, the equation (12) (vb = (2 · Δxb / The current overall spring element speed vb is calculated from [Delta] t) -vbz) (step 59). Next, using the detected damper displacement xd, the previous value xdz, the step time Δt, and the previous value vdz of the damper speed shifted in step 51, the current damper speed vd is calculated by the following equation (23). (Step 60). The damper speed vd is a change speed of the damper displacement xd (a single differential value of xd).
vd = {2 (xd−xdz) / Δt} −vdz (23)

Next, using the calculated damper speed vd, the previous value vdz of the damper speed shifted in the step 51, the step time Δt, and the previous value adz of the damper acceleration shifted in the step 51, the following equation (24) is used. The current damper acceleration ad is calculated (step 61), and the present process is terminated. The damper acceleration ad is a change speed of the damper speed vd, and is a first differential value of vd, in other words, a second differential value of the damper displacement xd.
ad = {2 (vd−vdz) / Δt} −adz (24)

  As described above, in the vibration control process according to the second embodiment, the overall spring element speed vb, the damper speed vd, and the damper acceleration ad calculated in steps 59 to 61 are the same as the previous time when this process is executed. The value is used to calculate the variable cr and the damper resistance change amount ΔP. In order to use the support member displacement xb3 reflecting the current actuator excitation force controlled by the execution of the step 11, the support member displacement xb3 used for the calculation of the overall spring element displacement xb in the step 57 is as follows. The support member displacement xb3 detected when the next vibration control process is executed may be used. The same applies to the calculation of the damper speed vd and the damper acceleration ad in Steps 60 and 61. As the damper displacement xd used for the calculation, the damper displacement xd detected at the next execution of the vibration control process is used. It may be used.

  As described above, according to the second embodiment, the test target mass damper 21 rotates the inner cylinder 22, the screw shaft 23 a movable with respect to the inner cylinder 22, and the displacement of the screw shaft 23 a with respect to the inner cylinder 22. And a ball screw 23 that is transmitted to the rotary mass 24 in a state of being converted into the above. Further, according to the damper resistance force P detected when the actuator excitation force controlled by the excitation control process is input to the vibration system S, the vibration equation of the second virtual vibration system VS2 (the above formula (1)). The damper predicted resistance force Ppr, which is the predicted value of the damper resistance force P of the test target mass damper 21, is calculated based on the equation (22) derived based on () (steps 52 to 54 in FIG. 10).

  As is clear from the equation (22) and the like, the damper predicted resistance force Ppr is an ideal mass damper that is predicted to be generated when the actuator excitation force controlled by the excitation control process is input to the second virtual vibration system VS2. This corresponds to a resistance of 21I. In other words, this damper predicted resistance force Ppr is applied to the second virtual vibration system VS2 when it is assumed that the test mass damper 21 is provided in the second virtual vibration system VS2 instead of the ideal mass damper 21I. This corresponds to the predicted value of the damper resistance force P that is predicted to be generated when the actuator excitation force controlled by the vibration control process is input.

  Furthermore, as described in the first embodiment, the rigidity kb1 of the virtual adjustment member MVA is set (step 21 in FIG. 6), and the calculated damper predicted resistance force Ppr and the set rigidity kb1 of the virtual adjustment member are set. Accordingly, the virtual adjustment member predicted displacement xb1pr is calculated (step 55 in FIG. 10). As described above, since the virtual adjustment member MVA is connected in series to the first virtual vibration system VS1, the virtual adjustment member prediction is performed according to the damper predicted resistance force Ppr and the virtual adjustment member rigidity kb1 as described above. By calculating the displacement xb1pr, the calculation can be appropriately performed.

  Further, in this case, the damper predicted resistance force Ppr corresponding to the damper resistance force P of the test target mass damper 21 detected during the input of the actuator excitation force controlled by the vibration control process is set to the virtual adjustment member predicted displacement xb1pr. Since it is used for calculation, the virtual adjustment member predicted displacement xb1pr commensurate with the test target mass damper 21 can be appropriately calculated.

  Further, a damper acceleration ad is calculated (step 61 in FIG. 11), and this damper acceleration ad is a twice differential value of the damper displacement xd, in other words, the acceleration of the screw shaft 23a with respect to the cylindrical portion 22. Further, the support member displacement xb3 is detected by the second displacement sensor 13, and a damper spring element predicted displacement xb2pr, which is a current predicted value of the displacement of the spring element composed of the inner cylinder 22 and the ball screw 23 of the test target mass damper 21, is obtained. Calculated (step 56). Further, the total of the virtual adjustment member predicted displacement xb1pr, the damper spring element predicted displacement xb2pr, and the support member displacement xb3 is calculated as the total spring element displacement xb (step 57), and the calculated total spring element displacement xb is calculated. Is used to calculate the overall spring element speed vb, which is the changing speed of the overall spring element displacement xb (the single differential value of xb) (step 59).

  Further, the damper predicted resistance force Ppr is calculated according to the damper acceleration ad and the overall spring element speed vb calculated when the actuator excitation force controlled by the excitation control process is input (FIG. 10). Steps 53 and 54, vbz and abz in equation (22). Since the damper acceleration ad and the overall spring element velocity vb have a close correlation with the damper predicted resistance force Ppr of the test target mass damper 21, the damper predicted resistance force Ppr can be calculated more appropriately.

  In the second embodiment, the virtual adjustment member predicted displacement xb1pr is calculated by dividing the damper predicted resistance force Ppr by the stiffness kb1 of the virtual adjustment member MVA. However, a predetermined map corresponding to both Ppr and kb1 is used. You may calculate by a search. In the second embodiment, the damper speed vd and the damper acceleration ad are calculated, but may be detected by a sensor.

  Further, with respect to the second embodiment, by executing Steps 2, 7, and 12 to 16 in FIG. 5 instead of Steps 52, 53, and 56 to 61, the ideal damper resistance force is substituted for the damper resistance force change amount ΔP. The change amount ΔPI may be calculated, and the calculated ΔPI may be used for calculating the damper predicted resistance force Ppr in step 54. Alternatively, instead of the detected damper resistance force P, the previous value Pprz of the predicted damper resistance force Ppr may be used for calculating the predicted damper resistance force Ppr in step 54.

In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, in the embodiment, the test target mass damper 21 includes the limiting mechanism 25 and the viscous body 31, but may be a mass damper that does not include at least one of the both 25 and 31. In the case of a mass damper that does not have the viscous body 31, the parameter related to the viscous resistance force cv · sgn (vdi) | vdi | ^α is set to a value of zero. In the embodiment, the transmission mechanism in the present invention is the ball screw 23, but it may be a transmission mechanism composed of a combination of a rack and a pinion. In addition, it is possible to appropriately change the detailed configuration within the scope of the gist of the present invention.

1 Test equipment 5 Right frame (support member)
6 Actuator 7 Connecting member (supporting member)
9 Connecting member (supporting member)
11 Load cell (Damper resistance detection means)
15 Control device (virtual displacement acquisition means, control means, performance parameter calculation means, damper predicted resistance force calculation means, setting means, movable part acceleration detection means)
21 Mass damper to be tested 22 Inner cylinder (main body)
23 Ball screw (transmission mechanism)
23a Screw shaft (movable part)
24 Rotating Mass S Vibration System 21I Ideal Mass Damper MS Support Member MVA Virtual Adjustment Member VS1 First Virtual Vibration System VS2 Second Virtual Vibration System P Damper Resistance kb1 Stiffness xbase of Virtual Adjustment Member MVA Basic Value (Basic Input Displacement)
xb1pr Virtual adjustment member predicted displacement (predicted value of displacement of virtual adjustment member)
Qv Maximum speed resistance (performance parameter)
Ceq attenuation coefficient (performance parameter)
md Rotational inertia mass (performance parameter)
41 Test equipment Ppr Damper predicted resistance ad Damper acceleration (acceleration of movable part relative to main part)

Claims (3)

A mass damper test device that configures an additional vibration system together with a rigid transmission member, has a rotating mass, and tests the performance of a mass damper for suppressing vibration of a structure,
A support member having rigidity and connected in series to a test target mass damper to be tested;
An actuator for inputting excitation force to a vibration system including the test target mass damper and the support member;
An ideal mass damper as the test target mass damper whose specifications are predetermined ideal values, and a virtual adjustment member for bringing the natural frequency of the first virtual vibration system in which the support members are connected in series to each other close to the predetermined frequency Is a predicted value of the displacement of the virtual adjustment member when it is assumed that an excitation force of the predetermined frequency is input to a second virtual vibration system connected in series to the first virtual vibration system. Virtual displacement acquisition means for acquiring based on a vibration equation of two virtual vibration systems;
Control means for controlling the excitation force of the actuator in accordance with a basic input displacement for inputting the excitation force of the predetermined frequency to the vibration system and the acquired predicted value of the displacement of the virtual adjustment member. When,
A damper resistance detecting means for detecting a damper resistance which is a resistance of the test mass damper;
A performance parameter for calculating a performance parameter representing the performance of the test target mass damper according to the damper resistance force detected when the excitation force of the actuator controlled by the control means is input to the vibration system A calculation means;
A testing apparatus for mass dampers, comprising: The test target mass damper includes a main body part, a movable part movable with respect to the main body part, and a transmission mechanism that transmits the displacement of the movable part with respect to the main body part to the rotary mass in a state where the displacement is converted into a rotational motion. In addition,
Based on the vibration equation of the second virtual vibration system according to the damper resistance force detected when the excitation force of the actuator controlled by the control means is input to the vibration system, the control means The resistance force of the ideal mass damper, which is predicted to be generated when the excitation force of the actuator controlled by the controller is input to the second virtual vibration system, is used as a damper predicted resistance force that is a predicted value of the damper resistance force. A damper predictive resistance calculating means for calculating;
Setting means for setting the rigidity of the virtual adjustment member,
The virtual displacement acquisition unit calculates a predicted value of displacement of the virtual adjustment member according to the calculated damper predicted resistance force and the set rigidity of the virtual adjustment member. The test equipment for mass damper as described in 1. Movable part acceleration detecting means for detecting acceleration of the movable part with respect to the main body part;
The damper predicted resistance force calculating means further includes the damper predicted resistance according to the acceleration of the movable part detected when the excitation force of the actuator controlled by the control means is input to the vibration system. The apparatus for testing a mass damper according to claim 2, wherein the force is calculated.

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