JP4431566B2 - Vibration / excitation testing machine - Google Patents

Description

  The present invention relates to a vibration / vibration tester, and more particularly to a vibration / vibration tester that drives an actuator that linearly expands and contracts to vibrate a device under test with an arbitrary vibration characteristic.

  In a vibration / vibration tester (vibration tester or vibration tester, hereinafter simply referred to as a vibration tester) that vibrates a device under test with arbitrary vibration characteristics and tests its seismic performance and durability. Is required to suppress the distortion of the excitation waveform and correctly reproduce the target excitation waveform.

  As a conventional method for improving the distortion of the excitation waveform, as shown in Japanese Utility Model Laid-Open No. 5-14879, it is improved by changing the control system, Japanese Utility Model Laid-Open No. 4-104550 and Japanese Patent Publication No. 56-72326. As described above, there is known a method of improving by adding a correction pulse to a control input.

As described above, various methods for improving the waveform distortion of the vibration tester have been proposed, but all of them are common in that the waveform distortion is improved by correcting the control input.
Japanese Utility Model Publication No. 5-14879 Japanese Utility Model Publication No. 4-104550 Japanese Patent Publication No. 56-72326

  However, in such a conventional method for improving the vibration waveform distortion, the waveform distortion caused by friction at the sliding portion of the actuator cannot be sufficiently improved.

  FIG. 24 shows a schematic configuration of a vibration testing machine using a conventional electrohydraulic servomechanism. This vibration testing machine is for expanding and contracting the cylinder rod 7 of the hydraulic cylinder 3 to vibrate the test object mounted between the mounting jigs 9 and 10, and supplying hydraulic pressure to the hydraulic cylinder 3. A hydraulic source 4 to be controlled, a servo valve 5 and a voltage / current converter 6 are provided.

  FIG. 25 shows the linear model, where K1 to K4 are constant gain parameters, A is the pressure receiving area of the cylinder rod 7, M is the cylinder rod mass (the total mass of the movable part including the piston, mounting jig, etc.), and s is Laplace operator. Further, since the dynamic characteristics of the servo valve 5 are not directly related to this explanation, it is assumed that the spool of the servo valve 5 has no response delay.

  The acceleration waveform when the vibration testing machine is subjected to sinusoidal vibration is shown in FIG. 26 (1) (the circled number 1 in the figure is represented in this way. The same applies hereinafter). Friction acts on the piston sliding portion and the like, and in particular, at the piston stroke end (dead point), the piston speed becomes zero and the static friction force acts, so that the acceleration waveform is greatly distorted stepwise in this way. The distortion caused by this friction becomes particularly significant when the cylinder rod is lightened. In order to completely cancel this distortion by the control input, it is necessary to change the acceleration waveform stepwise by the control input.

  FIG. 27 shows the transfer characteristic from the control input to the acceleration of the linear model shown in FIG. The gain characteristic diagram of the transfer characteristic shown in FIG. 27 is as shown in FIG. 28, and has a delay characteristic in the high frequency region. For this reason, even if a step-like signal is applied to the control input, the acceleration waveform cannot change stepwise as shown in FIG.

  For this reason, the correction of the control input can speed up the convergence after the acceleration waveform is distorted as shown in (2) of FIG. 26, but the stepwise distortion a cannot be reduced. The distortion of the acceleration waveform could not be improved sufficiently.

  The present invention has been made in view of the above-described problems of the prior art, and it is an object of the present invention to improve stepwise acceleration waveform distortion due to friction that cannot be sufficiently improved only by correcting control inputs. And

  The present invention provides a vibration / vibration tester that drives an actuator that linearly expands and contracts to vibrate a device under test with an arbitrary vibration characteristic. A heavier weight member is attached to the component within the allowable range of the thrust of the testing machine, and the heavier the testing machine, the larger the thrust of the testing machine is, and according to the mass of the component of the device under test. A control device that has a plurality of control systems designed to improve deterioration of responsiveness due to an increase in the mass of the component of the device under test by the weight member by changing the control system. It is characterized by.

  According to the present invention, the influence of friction on acceleration is reduced by increasing the mass of the component of the device under test. Thereby, it is possible to significantly reduce the distortion of the excitation waveform caused by friction without modifying the excitation tester, and to keep the distortion of the excitation waveform within an allowable range. In addition, a decrease in responsiveness caused by increasing the mass of the component of the device under test can be improved by changing the control system.

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 1 shows a schematic configuration of a vibration testing machine to which the present invention is applied. This vibration testing machine includes a portal frame 2 installed on the ground 1, a hydraulic cylinder 3 fixed to the portal frame 2, a hydraulic source 4 that supplies hydraulic pressure to the hydraulic cylinder 3, and hydraulic pressure from the hydraulic source 4. The servo valve 5 controls the supply of hydraulic pressure to the cylinder 3. A control signal from a controller (not shown) is input to the servo valve 5 via the voltage / current converter 6, and the expansion / contraction amount of the cylinder rod 7 of the hydraulic cylinder 3 is controlled. Here, the hydraulic cylinder 3 is a double rod type, and the cylinder rod 7 penetrates vertically.

  A cross beam 8 connecting two arms 2a and 2b extending in parallel upward from the portal frame 2 is provided, and mountings are provided for attaching a test object to the upper end of the cylinder rod 7 and the lower surface of the cross beam 8, respectively. The jigs 9 and 10 are fixed. If the lower part of the object to be tested (not shown) is connected to the mounting jig 9 and the upper part to the mounting jig 10, and the cylinder rod 7 is expanded and contracted, compression and tensile forces act alternately on the object to be tested. The body is vibrated up and down.

  Under such a configuration, in the present invention, the weight absorbing member 11 for absorbing disturbance is attached to the movable portion of the hydraulic cylinder 3 to increase the mass of the movable portion. Here, a disk-shaped weight member 11 is rigidly coupled between the cylinder rod 7 and the mounting jig 9. With such a structure, the weight member 11 can be easily attached to an existing testing machine.

  In addition, the attachment position of the weight member 11 is not limited to this position, and may be another position as long as it is a movable part. For example, you may attach to the lower end of the cylinder rod 7, as shown in FIG. In this case, since the force in the tensile direction acts on the cylinder rod 7 due to the mass of the weight member 11, even if the cylinder rod 7 is eccentric with respect to the vertical axis due to the bending force when the test object is compressed, the action of returning the cylinder rod 7 is restored. There is an advantage that the burden on the seal is reduced.

  Further, a configuration in which the weight member 11 and the mounting jig 10 are integrated, that is, a configuration in which the mounting jig 10 is eliminated and the weight member 10 is used as a jig for mounting the test object may be employed. In this case, in addition to the advantage that the configuration of the testing machine can be simplified, there is also an advantage that it is possible to surely prevent the weight member 11 from being attached or dropped.

  Next, the operation will be described. Here, it is first shown that the stepwise vibration waveform distortion due to friction cannot be improved by the conventional method of changing the control system.

  The block diagram shown in FIG. 3 is obtained by adding a control system to the linear model of the vibration testing machine shown in FIG. 1 or FIG. In the figure, K1 to K4 are constant gain parameters, A is the pressure receiving area of the cylinder rod 7, M is the mass of the cylinder rod (the total mass of the movable part including the piston, mounting jig, etc.), and s is the Laplace operator. C1 to C4 are arbitrary controllers, and the detection signals are the differential pressure between the cylinder chambers of the hydraulic cylinder 3, and the acceleration, speed, and displacement of the cylinder rod 7.

Here, the block diagram shown in FIG.
Since the force a before the disturbance is equivalent to “differential pressure × A”, it can be represented by differential pressure feedback.
The force b after the disturbance is applied is equivalent to “acceleration × M” and can be represented by acceleration feedback.
The flow rate c can be expressed by addition / subtraction of the signals d, e, and f. Since d is equivalent to “differential pressure × K3”, it can be represented by differential pressure feedback, and e is equivalent to “speed × K4”, so speed feedback F is equivalent to “control input × K1”, and feedback of the control input can be represented by adjusting these controllers in advance based on the arbitraryness of the controllers C1 to C4.
Since the signal g is equivalent to “flow rate × K2”, it can be represented by another signal in the same manner as the other signals as described above.
Therefore, all control systems for the electrohydraulic servomechanism can be expressed.

  In FIG. 3, in order to evaluate the distortion of the acceleration waveform due to friction, the transfer characteristics from disturbance (friction) to acceleration are obtained as shown in FIG. Here, the characteristic change when the controllers C1 to C4 are adjusted in the characteristic of FIG. 4 (when the control system is changed) will be described.

  FIG. 5 shows (1) when C4 is increased with respect to (2) when C1 = C2 = C3 = 0 and C4 = positive constant value (hereinafter referred to as this characteristic). (3) when it is made smaller is shown. FIG. 6 shows (1) when C3 is a negative constant value with respect to the reference (2), and (3) when C3 is a positive constant value. FIG. 7 shows (1) when C2 is a negative constant value with respect to the reference (2), and (3) when C2 is a positive constant value. FIG. 8 shows (1) when C1 is a negative constant value with respect to the reference (2), and (3) when C1 is a positive constant value.

  As shown in FIGS. 5 to 8, it is understood that the resonance frequency, the attenuation characteristic of the resonance point, and the characteristic in the low frequency region below the resonance frequency can be changed by feeding back each detection signal. However, no matter what the detection signal is fed back, the characteristics higher than the resonance frequency cannot be changed. This is because the characteristics in the high frequency region are determined by the coefficient ratio of the highest order of the denominator numerator of the transfer function shown in FIG. That is, in the example shown in FIG. 4, the coefficient ratio is 1 / M, and the coefficient ratio cannot be adjusted even if the controllers C1 to C4 are adjusted.

  In the example of FIGS. 5 to 8, the case where the controllers C1 to C4 are constants has been described. However, even if it is assumed that the controllers C1 to C4 have arbitrary frequency characteristics, the coefficient ratio described above should be adjusted. I can't. In general, a controller with frequency characteristics can be described as follows (a and b are real numbers, n is zero or a positive integer):

When this equation is applied to the controller C1 in the transfer function shown in FIG.

And the highest order of the denominator numerator is n + 3. However, their coefficient ratio is, after all,

Thus, even if the controller has an arbitrary frequency characteristic, the coefficient ratio cannot be changed. Although the case where the controller C1 has frequency characteristics is shown here, it can be confirmed by similar calculation that the coefficient ratio cannot be changed even when applied to the controllers C2 to C4.

  Thus, in the characteristics from disturbance to acceleration, the characteristics in the high frequency region cannot be changed by adjusting the control system. In other words, since the characteristic of the high frequency region is determined by the coefficient ratio 1 / M of the highest order of the denominator of the transfer characteristic, the characteristic of the high frequency region can be changed only by changing the movable part mass M. It can be said.

  Based on this idea, in the present invention, the weight member 11 is attached to the actuator movable portion, and the movable portion mass M is increased. FIG. 9 shows the characteristic change when the movable part mass M is changed. (1) in FIG. 9 shows a case where the mass M of the movable part is increased with respect to the reference waveform (2) when C1 = C2 = C3 = 0 and C4 = positive constant value. Indicates a case where the mass M of the movable part is reduced. In this way, by changing the mass M of the movable part, it is possible to change the characteristics of the high frequency region that could not be changed by changing the control system.

  When this is evaluated by time response, it becomes as shown in (3) of FIG. Thus, by increasing the mass M of the movable part, the stepwise distortion a due to friction can be reduced to b, and the acceleration waveform distortion can be greatly improved.

  5 to 9 show the damping of the resonance point in a considerably bad state so that the characteristic change can be easily seen. FIG. 10 shows the waveform when the damping is improved for simplicity of description. The portion distorted stepwise by friction is the same regardless of the resonance point damping.

  By the way, if the weight member 11 is attached to the movable part of the actuator and the mass of the movable part is increased, distortion of the acceleration waveform is greatly improved. However, if the mass of the movable part is increased, (a) Problems such as deterioration in responsiveness and (b) increase in required thrust of the testing machine arise. These points will be described below.

(A) Deterioration of responsiveness First, the problem that responsiveness deteriorates with an increase in mass of the movable part will be described.

  FIG. 11 shows target followability (responsiveness) corresponding to FIG. 9 (characteristics from disturbance to acceleration) when the movable part mass of the actuator is changed. As shown in this figure, it can be seen that even if the movable part mass of the actuator is increased, the responsiveness is hardly deteriorated in the bandwidth (up to about 10 Hz). Normally, the resonance frequency is set on the higher frequency region side compared to the bandwidth (the highest frequency that can be responded), so if the mass increases so that the resonance frequency does not fall within the bandwidth frequency band, the responsiveness deterioration is practical. There is no problem above.

  Further, even if the mass of the movable part is large and the resonance frequency falls within the frequency band of the bandwidth, the characteristics of the low frequency region can be adjusted by adjusting the controllers C1 to C4 as shown in FIGS. It can be improved sufficiently (FIGS. 5 to 8 are characteristics of disturbance (input) to acceleration (output), but the frequency band that can be adjusted by each controller is the same regardless of where the input and output are taken. ).

  (4) of FIG. 12 shows an example in which the resonance point has entered the band of the bandwidth shown in FIG. 11 (up to about 10 Hz) because the mass of the movable part of the actuator has been increased too much. In such a case, as shown in FIG. 6, the resonance frequency is set high with minor velocity feedback, and the resonance point damping is adjusted with minor acceleration feedback as shown in FIG. The displacement feedback gain may be adjusted as shown. Thereby, target value followability can be improved as shown to (5) of FIG. In FIG. 12 (5), since the resonance point damping is improved by acceleration feedback, the peak of the resonance point cannot be confirmed.

  As described above, since the responsiveness can be improved by changing the control system, the deterioration of the responsiveness caused by increasing the mass of the movable part of the actuator does not cause a problem in practice.

(B) Increase in Required Thrust of Test Machine Next, a problem that the thrust required for the actuator is increased by increasing the mass of the movable part will be described.

  Usually, tests are performed under various vibration conditions on the testing machine, but the maximum thrust required from these vibration conditions can be calculated (the acceleration is obtained by differentiating the target value of displacement twice, In addition, if the mass of the moving part is applied, the thrust required for the testing machine alone can be calculated, so the reaction force from the DUT can be added to this), so the maximum thrust required for these excitations and the testing machine It is possible to calculate how much the thrust of the testing machine has a margin by comparing with the maximum thrust.

  At this time, if the thrust of the testing machine is sufficiently large with respect to the maximum thrust required for excitation, the mass of the weight member to be added can be determined with a considerable degree of freedom, but the stepped shape of the acceleration waveform due to friction can be determined. When it is desired to sufficiently reduce the strain, it is necessary to at least double the mass of the movable part. In addition, considering the thrust loss due to friction and internal leakage of oil, adding a weight member will reduce the weight of the weight member so that the maximum thrust required for vibration does not exceed 90% of the maximum thrust of the testing machine. When it is determined, it is necessary to allow a margin for the thrust of the testing machine.

  Also, if the thrust of the testing machine is not large enough for the maximum thrust required for excitation, increasing the mass of the moving part will result in insufficient testing machine thrust, so the mass of the moving part will increase almost. I can't. Further, if the mass of the movable part is not increased, the maximum thrust required for excitation can be kept small, but this time it is necessary for excitation from the formula “mass M × acceleration a = force F + disturbance W”. When the thrust F is small, the disturbance W becomes relatively large with respect to the thrust F required for excitation, and the influence of the disturbance appearing in the acceleration a becomes large.

  Therefore, when the thrust of the testing machine is not sufficiently large with respect to the maximum thrust required for the vibration as described above, the weight member 11 can be attached and detached, and the weight member 11 is attached within a range where the thrust of the testing machine is sufficient. What should I do?

  Thereby, even when there is no margin in the thrust of the testing machine, the weight member 11 is within the allowable range of the thrust of the testing machine under the excitation condition where the distortion of the acceleration waveform appears to be large (when the thrust required for the excitation is small). And the distortion of the acceleration waveform can be sufficiently reduced.

  Hereinafter, an example in which the weight member 11 can be removed and the mass of the movable part can be changed to two types in the vibration tester shown in FIG. Note that the numerical value described in parentheses on the right side of the thrust required for each excitation condition is the influence of a disturbance of 1 [N] on the acceleration waveform distortion.

  In the above example, the thrust of the testing machine is 10 [N]. In the conventional structure, the thrust required under the vibration condition 1 is 8 [N], and the thrust required under the vibration condition 2 is 4 [N]. Yes, the thrust of the testing machine is sufficient under both excitation conditions.

  However, when the present invention is applied and the weight member 11 is attached to the movable part and the mass of the movable part is doubled (M → 2M), the thrust required for vibration is doubled as it is. Under condition 1, the thrust of the testing machine is insufficient.

  Therefore, in such a case where the thrust is insufficient, the weight member 11 is removed and the vibration is performed with the conventional structure. Thereby, the thrust required for excitation can be suppressed to 8 [N]. Under the vibration condition 1 where the influence of the disturbance is unlikely to appear, the influence of the disturbance is hardly (1/8) even if the weight member 11 is removed. Thus, when there is no margin in the thrust of the testing machine, the thrust of the testing machine may be secured with priority.

  On the other hand, in the vibration condition 2 with a sufficient thrust, a weight member is attached to perform vibration. Thereby, in the conventional structure, the influence of disturbance is very likely to appear (1/4) under the excitation condition 2, whereas the influence of disturbance is reduced to (1/8) by attaching the weight member 11. Can be suppressed.

  In the above example, there are two types of mass of the movable portion of the actuator depending on whether the weight member 11 is attached or removed. However, the mass of the movable portion may be changed to several types by further dividing or adding the weight member 11. . This makes it possible to efficiently reduce distortion of the acceleration waveform under more excitation conditions while changing the mass of the movable part according to the thrust required for excitation and effectively using the thrust of the testing machine. . The case where the movable part mass of the actuator can be changed to two types and the case where it can be changed to four types are shown below.

  In this way, it can be seen that the influence of disturbance can be reduced under more excitation conditions when the movable part mass of the actuator can be changed to four types than when it can be changed to two types. As a result, even when the thrust of the testing machine is not large enough for the thrust required for excitation, the actuator can change the mass of the moving part so that the acceleration waveform is within the range where the thrust of the excitation machine is sufficient. Distortion can be improved.

  The examples shown in Tables 1 and 2 are examples for easily explaining the effectiveness of the present invention, and do not limit the scope to which the present invention can be applied.

  In order to change the mass of the movable part of the actuator in the configuration in which the weight member 10 and the mounting jig 11 are integrated, a plurality of mounting jigs and weight members having different masses are prepared, The mounting jig / weight member may be replaced in accordance with the above.

  By the way, as shown in Table 1 and Table 2, when the movable part mass of the actuator can be changed into several types, as can be seen from FIG. 9, the damping of the resonance point changes according to the movable part mass of the actuator. To do. For this reason, adjusting the control system with the actuator moving part light and then increasing the actuator moving part makes the control system conservative (the state where the gain can be further increased because the resonance point is better damped). End up. Conversely, if the moving part of the actuator is lightened after adjusting the control system in a state where the moving part of the actuator is heavy, the control system appears to oscillate (the state where the gain has to be lowered because the damping of the resonance point deteriorates). turn into.

  In this way, there is a limit to handling a plurality of movable part masses with one control system, and if the movable part mass of the actuator can be changed, the control system is also changed according to the movable part mass of the actuator. There is a need to.

  Therefore, when the movable part mass is variable, a plurality of control systems adjusted according to the movable part mass of the actuator are prepared in advance, and the control system is switched according to the number of weight members 11 to be attached. To do. For example, as shown in FIG. 13, when three weight members 11 are prepared and the movable part mass can be changed to four types, each is designed to be optimal when there are 0 to 3 weight members 11. The prepared four types of controllers may be prepared, and the controllers may be switched according to the number of weight members 11 to be attached.

  As a result, a control system corresponding to the mass of the movable part of the actuator can be selected each time, and the control system is prevented from becoming conservative or oscillating, and optimal control becomes possible.

  If the control system has been used for PID control of a displacement loop that has been used for a long time, and only P control, even if the gain is adjusted each time according to the mass of the movable part of the actuator, the adjustment does not take much time. Therefore, the advantage of preparing a plurality of control systems in advance is not so great. However, in the case where the control system other than the addition of the weight member 11 is used to reduce the acceleration waveform distortion, it is necessary to effectively use the detection signal other than the displacement loop (also from FIGS. 5 to 7, from the disturbance). The effectiveness of feeding back detection signals other than displacement to adjust the transfer characteristics of acceleration can be fully understood.) For this reason, the gain of the control system to be adjusted according to the mass of the movable part of the actuator increases, and when the controllers C1 to C4 shown in FIG. Making it virtually impossible to adjust.

  For this reason, when the actuator moving part mass and control system are changed to reduce acceleration waveform distortion and the actuator moving part mass is variable, the actuator moving part mass is previously determined according to the actuator moving part mass. It is a very effective means to prepare an adjusted control system so that these control systems can be selected.

  In the example shown in FIG. 13, the movable part mass of the actuator can be changed to four types, and the control system has four types corresponding to it. If it can be changed, it takes time to adjust the control system corresponding to each. In addition, the amount of memory used when storing these control systems (such as controller gain) in the control device also increases.

  Therefore, in such a case, it is possible to prepare fewer control systems than the types of change of the movable part mass of the actuator and use one control system for a plurality of movable part masses. For example, when the mass of the movable part of the actuator can be adjusted in 10 steps per kg from 1 kg to 10 kg, instead of preparing a 10-step control system adjusted for each 1 kg step, A five-stage control system adjusted every 2 kg may be prepared. By doing so, an empty container is connected to the movable part of the weight member 11, and the mass of the movable part can be changed in a stepless manner by adjusting from a state where no liquid is put to a state where the liquid is filled. It is possible to deal with cases.

  In addition, when selecting and using a control system that has been adjusted in advance according to the mass of the movable part of the actuator, the control device is provided with an adjuster for the gain of the control system and other parameters, depending on the excitation conditions and load conditions, Or when controlling with respect to the movable part mass of several actuators by one control system, you may enable it to finely adjust a control system according to the movable part mass of an actuator. As a result, the degree of freedom of control system adjustment increases, and the control system can always be controlled in an optimal state.

  The embodiment of the present invention has been described above by taking the vibration testing machine using the electro-hydraulic servo mechanism as an example, but the vibration test using the electro-mechanical servo mechanism (electric ball screw servo mechanism) instead of the electro-hydraulic type. Even if it is a machine, only the transmission mechanism to the movable part of the actuator is different, and the present invention can be applied as it is.

  Further, as shown in FIG. 14, an electrohydraulic actuator vibration testing machine in which the flow control valve (servo valve) is not used and the electrohydraulic servo mechanism is composed of a hydraulic pump 13, a motor 12, and the like. Even so, the present invention is applicable. In FIG. 14, only a portion that replaces the flow rate control valve is described in a simplified manner, but actually, other components such as an oil tank, an accumulator, and a check valve are generally used together. Furthermore, the structure of the electrohydraulic actuator has many other configurations such as a variable discharge amount of the hydraulic pump.

  Further, the present invention can be similarly applied to a vibration tester as shown in FIG. In the structure shown in FIG. 15, the weight member 11 is rigidly coupled to the vibration table 14 connected to the tip of the cylinder rod 7 of the hydraulic cylinder 3 in order to increase the movable part mass of the actuator. When the friction between the vibration table 14 and the weight member 11 is sufficiently large and they are driven as a unit, it is not necessary to rigidly couple them. The same effect can be obtained when the weight member 11 is attached to the part of the device under test where the vibration is transmitted from the vibration table 14.

  In the above embodiment, the movable part mass of the actuator is increased by attaching the weight member. Instead of attaching the weight member, the mass of the movable part is increased by increasing the mass of the components of the movable part. You may do it. For example, in the vibration testing machine shown in FIGS. 16 and 17, the mass of the movable part is increased by an amount corresponding to the weight member by increasing the diameter of the cylinder rod 7 of the hydraulic cylinder 3 or by increasing the thickness of the piston 15 in the axial direction. Increasing.

  Even with such a configuration, the distortion of the excitation waveform can be greatly reduced as in the case where the weight member is attached to the cylinder rod, and the distortion of the excitation waveform can be within an allowable range. In particular, as compared with the method of attaching the weight member to the cylinder rod, the center of gravity of the movable part is located near the center of the movable part, so that the effect of improving the stability with respect to the shake in the front-rear and left-right directions is obtained. In addition, it is possible to prevent the attachment failure and dropout of the weight member.

  Here, the mass of the movable part is increased by increasing the cylinder rod 7 or the piston 15, but the mass of the movable part may be increased by changing the material thereof to a material having a higher specific gravity.

  In general, an anti-vibration tester using a hydraulic cylinder has a non-rotating mechanism rigidly coupled to the movable part for preventing the rotation of the cylinder rod. The mass of the components of the detent mechanism is equivalent to the weight member. The mass of the movable part may be increased by making it heavier only or by attaching a weight member to the detent mechanism. In this case, the same effect as that obtained by attaching the weight member to the cylinder rod can be obtained. For example, in the vibration testing machine shown in FIG. 18, the mass of the movable part is increased by attaching the weight member 11 to the detent mechanism 16 that is rigidly coupled to the attachment jig 9.

  Since the detent mechanism is often attached to the side surface of the actuator, it is effective when the weight member cannot be attached either above or below the actuator due to space restrictions.

  The vibration testing machine is equipped with various sensors such as a displacement sensor, speed sensor, acceleration sensor, and load sensor. The mass of the sensor or jig for mounting the sensor rigidly connected to the movable part is used as the weight member. Even when the weight member is rigidly coupled to the sensor or the sensor mounting jig, the same effect as when the weight member is attached to the cylinder rod can be obtained.

  For example, in the vibration testing machine shown in FIG. 19, the jig 18 for mounting the displacement sensor 19 is rigidly coupled to the lower end of the cylinder rod 7. The mass of the sensor mounting jig 18 is increased by the mass of the weight member. Thus, the same effect as that obtained when the weight member is attached to the lower end of the cylinder rod shown in FIG. 2 can be obtained.

  Depending on the sensor mounting method, it may be easier to manufacture the sensor mounting jig by making the sensor mounting jig itself larger and heavier than attaching a new weight member, or mounting the weight member on the sensor mounting jig. There is an advantage that it can be reduced.

  Further, even when the weight member is attached to the actuator movable portion, the attachment position does not have to be outside the movable portion constituent element. For example, the cylinder rod 7 is attached as in the vibration testing machine shown in FIG. A hollow cylindrical structure may be used so that the weight member 11 can be fixed at an arbitrary position in the cylinder rod 7.

  In this case, compared to the method of attaching the weight member from the outside of the movable component, the position of the center of gravity of the movable portion can be set freely, so that the stability against lateral shake in the front-rear and left-right directions can be improved.

  In the vibration testing machine shown in FIG. 20, three weight members 11 are inserted, but the vibration testing machine can be prepared by changing the number of inserted weight members or preparing several weight members having different masses. Therefore, when the thrust of the testing machine is not sufficient, the mass of the weight member can be reduced.

  Further, the weight member may be connected to the movable part via a link or the like. For example, in the vibration testing machine shown in FIG. 21, the weight member 11 is attached to the lower end of the cylinder rod 7 via the lever 20. In this example, the fulcrum 21 of the lever 20 is fixed (free in the rotation direction), and the lower end of the cylinder rod 7 and the connecting portion 22 of the lever 20 are slidable, and the weight member 11 is adapted to the expansion and contraction of the cylinder rod 7. Is going up and down. In addition, what is necessary is just to set it as the structure which can slide the fulcrum 21 of the lever 20 when fixing the connection part of the Linder rod 7 and the lever 20 (it is free in a rotation direction).

  In such a configuration, by adjusting the mounting position of the weight member 11 connected to the lever 20, the same effect as the case where the movable part mass of the vibration testing machine shown in FIG. Is obtained. That is, if the weight member 11 is attached closer to the fulcrum 21, the same effect as when the weight member 11 is lightened can be obtained, and if it is attached away from the fulcrum 21, the same effect as when the weight member 11 is increased can be obtained. As a result, when the thrust of the testing machine is insufficient, the weight member 11 is attached to the fulcrum 21 to ensure the thrust of the testing machine, and when the thrust of the testing machine has a margin, the weight member is separated from the fulcrum 21. Mounting and vibration waveform distortion can be further improved.

  In the embodiments described above, the distortion of the excitation waveform is improved by increasing the mass of the movable part of the actuator. However, the mass of the component of the test object that is rigidly coupled to the movable part is the weight member. The same effect as when the mass of the moving part is increased can be obtained even if the weight of the part under test is increased by attaching a weight member to the component. It is done.

  For example, when the device under test is a spring, as shown in FIGS. 22 and 23 (a), the test may be performed by using the spring 23 and the weight member 11 as an integrated body. When the body is a damper, as shown in FIG. 23 (b) or (c), the weight member 11 is attached to a portion (here, the outer tube 24a) rigidly coupled to the movable portion of the actuator among the components of the damper 24. It is sufficient to perform the test with the attached.

  By adopting such a configuration, even if the vibration tester is used with the conventional structure, the same effect as when the mass of the movable part of the vibration tester is increased can be obtained. It is possible to reduce the influence on the vibration waveform due to the friction of the actuator without modification.

It is the figure which showed schematic structure of the vibration testing machine which concerns on this invention. It is the figure which showed the Example which changed the attachment position of the weight member. It is the block diagram which added the control system to the linear model of a vibration testing machine. It is the block diagram which showed the transfer characteristic from disturbance (friction) to acceleration. It is the figure which showed the characteristic change at the time of changing the controller C4. It is the figure which showed the characteristic change at the time of changing the controller C3. It is the figure which showed the characteristic change at the time of changing the controller C2. It is the figure which showed the characteristic change at the time of changing the controller C1. It is the figure which showed the characteristic change at the time of changing movable part mass. It is the figure which showed the acceleration waveform at the time of changing movable part mass. It is the figure which showed target value followability at the time of changing movable part mass. It is the figure which showed the improvement effect of target value followability by change of a control system. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed another embodiment of this invention. It is the figure which showed the example of attachment of the weight member to a to-be-tested body. It is the figure which showed schematic structure of the conventional vibration testing machine. It is the block diagram which showed the linear model. It is the figure which showed the acceleration waveform when the conventional vibration testing machine is sine-wave-excited. It is the block diagram which showed the transfer characteristic from control input to acceleration. It is the gain diagram. It is the figure which showed the acceleration waveform when the signal of a step is applied.

Explanation of symbols

3 Hydraulic cylinder 4 Hydraulic source 5 Servo valve 7 Cylinder rod 9 Mounting jig 10 Mounting jig 11 Weight member

Claims (1)

In a vibration / vibration testing machine that drives an actuator that linearly expands and contracts to vibrate a device under test with arbitrary vibration characteristics, the component of the device under test that is rigidly coupled to the movable part of the actuator Within the allowable range of thrust of the testing machine, attach a weight member that is heavier as the thrust of the testing machine is larger than the maximum thrust required for vibration,
It has a plurality of control systems designed according to the mass of the component of the device under test, and the control system is changed due to the deterioration of the responsiveness due to the mass of the component of the device under test being increased by the weight member A vibration / vibration tester characterized by comprising a control device that can be improved by doing so.