JPH0457690A - Vibration absorber, driving device, load uniting element and multi-shaft mechanism - Google Patents

Abstract

PURPOSE:To simultaneously absorb the vibrations of a vibrating body in plural directions by providing the plural faces of the mounting base of a vibration absorber with cantilever-shaped high damping members. CONSTITUTION:One face of a mounting base 12 formed of a polyhedron, for instance a cubic body, is set on a vibrating body 11, and the one end of a cantilever-shaped high damping member( formed of rubber inserted between metallic plates 14a, 14b) is fixed to a mounting base crossing the single or plural directions of vibrations having a large amplitude at a right angle. A weight 16 is then provided at the other end or any position of the high damping member, and the natural frequency of its primary bending vibration is made to coincide with the resonance frequency or anti-resonance frequency of the vibration body 11.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention aims to improve the settling characteristics in order to realize high-speed and high-precision operation of a load driven by an actuator directly or through a reduction gear. The present invention relates to a vibration absorbing device that stabilizes the servo characteristics of an actuator without significantly increasing the actuator load, and a multi-axis mechanism equipped with the vibration absorbing device.

[Conventional technology]

Means for stabilizing unstable characteristics of the motor servo system due to mechanical resonance of a load driven by an actuator, for example, an electric motor directly or via a reduction gear, and unstable behavior when the operation is stopped include ■ electric motor A filter is inserted into the control circuit to reduce the gain at the mechanical resonance frequency of the driving force command. They can be roughly divided into two types: mechanically reducing the vibration amplitude by providing a device (with a filter) and reducing the gain at the resonant frequency.

The former example, as described in Japanese Unexamined Patent Publication No. 60-39397, provides a notch filter in the driving force command generation section that has an inverse gain characteristic to the gain characteristic near the resonant frequency, thereby increasing the transfer gain at the resonant frequency. The objective is to reduce the amount of noise and improve control gain.

Also, here is an example using the latter passive vibration absorption device.

As described in the Transactions of the Japan Society of Mechanical Engineers, Volume 55, No. 516 (August 1989), Pages 2014 to 2021, there is an example of a cantilever-shaped bone dynamic vibration absorber using rubber and weights. It was stated. Further, as described in Japanese Utility Model Application Publication No. 1-166087, a device is described in which a passive vibration absorbing device using a cylindrical rubber and a weight is attached to the tip of a robot arm.

On the other hand, an example using an active vibration absorbing device is the 154th Proceedings of the Japan Society of Mechanical Engineers & 890-26 (June 1989).
As described on pages 155 to 155, a piezoelectric element is installed on the side of the cantilever beam to suppress its vibration, and a voltage is applied to the piezoelectric element based on the equation of motion of the beam's bending vibration to suppress the beam's bending vibration. A method of active suppression was described.

[Problem to be solved by the invention]

Although the above-mentioned first conventional technology has an excellent gain reduction effect, since the notch filter transfer characteristic G' is given by the following formula in the frequency domain (S domain), it is slightly lower than the center angular frequency ω of the notch filter. There is a phase delay of about 90° at the angular frequency, ω.Although there is a 180° phase lead at the second angular frequency, ω. At the following angular frequencies, the phase delay of the open loop transfer characteristic of the servo system exceeds 180°, making the servo system unstable, narrowing the bandwidth, and causing problems such as not being able to increase the control gain. However, there is a problem in that it is not an effective countermeasure against mechanical resonance points particularly in the low frequency range where the open loop transfer gain is high.

In addition, the second conventional technology provides a cantilever-shaped passive vibration absorbing device made of rubber, a metal plate, and a weight at the tip of the arm, which has a large vibration amplitude, when rotating an arm-shaped load using a step motor. There are some types of motors that are designed to vibrate, but since a weight is provided at the tip of the arm, the load inertia of the motor increases, making high acceleration/deceleration operations difficult. In addition, since actual vibration is a superimposition of vibration modes such as not only in the turning direction but also in the upward, downward, and radial directions, damping in multiple directions may be necessary, but in this known example, There are some issues that are not mentioned.

Furthermore, in the third conventional technology, a passive vibration absorption device made of rubber and a weight is provided at the tip of the arm in order to reduce the residual vibration of the robot arm, but this increases the load inertia of the arm rotation drive motor. There is a problem that prevents the robot arm from moving at high speed.

In addition, the fourth conventional technology focuses on the bending vibration of a cantilever beam and calculates the required vibration damping force from the equation of motion, but it cannot be applied to a multi-axis mechanism with multiple arms like a normal robot. However, since the calculation formula becomes complicated and the amount of calculation becomes enormous, the sample period of the control system becomes longer, and there is a problem that the responsiveness thereof deteriorates.

The purpose of the present invention is to effectively suppress vibrations in each direction of the drive device/load combination without significantly increasing the load inertia of the drive device, to stabilize the servo system of the drive device, and to perform work at the tip of the load. The objective is to enable high-speed, high-precision positioning operations.

[Means to solve the problem]

In order to achieve the above object, the present invention has taken the following measures. That is, in order to perform effective vibration damping without significantly increasing the load inertia of the motor, for example, when damping an arm-shaped load, a polyhedral, for example, a cubic mounting base is provided at the end of the arm. A vibration absorbing device in the form of a cantilever beam with a weight at the tip is provided from the mounting base toward the base of the arm.

[Effect]

As configured in this manner, the vibration absorbing device and the drive device/load combination of the present invention have the following effects.

First, the arrangement of the vibration absorbing device that does not significantly increase the load inertia of the electric motor will be described using FIG. Here, an external cylinder rotor type direct drive electric motor 1 is used as a drive device.
Consider an arm 3 equipped with a weight 4 at its tip as a load. FIG. 34 shows a perspective view of a drive device/load combination in which a cantilever-shaped vibration absorbing device 6 with a weight at the tip is provided at a large vibration amplitude portion of the arm 3.

Weight mass rnw of the vibration absorber, mass rn'+ of the vibration absorber mounting part, distance Q1 from the center of gravity of the motor to the weight
, Distance Ql from the center of gravity of the motor to the mounting part of the vibration absorber
Then, the mass of the beam part of the vibration absorber is mw )m
Since it is smaller than i, the increment ΔJ of the motor's load inertia due to the installation of the vibration absorber is (2)
It is shown by the formula.

ΔJ=mwQw2+m1Q12...(
2) Assuming that the length of the beam of the vibration absorber is Q, in the drive device/load combination of the present invention, since the weight of the vibration absorber is provided in the direction of the center of gravity of the electric motor when viewed from the mounting part, Qw
=Q+Qb.

In the drive device/load combination of the second conventional example, since the weight of the vibration absorber and the mounting portion are provided at the same distance from the center of gravity of the electric motor, QIl=QI. Therefore, when the drive device/load combination of the present invention has the same mw as the conventional example,
m-Qb (It is possible to reduce the motor load inertia by 2111100 J.

Next, the vibration reduction effect of the vibration absorbing device will be explained using FIGS. 35 to 43. 35 shows a dynamic model of the drive device/load combination shown in FIG. 34 without the vibration absorbing device, FIG. 36 shows a dynamic model of the drive device/load combination shown in FIG. 34, and FIG. The figure shows a block diagram of the motor speed control section of the drive device/load combination shown in Fig. 34, Fig. 38 shows the torque speed transmission characteristic of the mechanism model shown in Fig. 35, and Fig. 39 shows the transfer characteristic of Fig. 38. 40 shows the transmission characteristics of the speed detection section, and FIG. 41 shows the transmission characteristics between A and 8 in FIG. 37 when no vibration absorption device is used. FIG. 42 shows the torque/speed transmission characteristics of the mechanical model when a vibration absorber is used, and FIG. 43 shows the transmission characteristics between A and 8 in FIG. 37 when a vibration absorber is used.

The drive device/load combination shown in Fig. 34 has no vibration absorption device.
If it exists, it can be expressed by the dynamic model shown in FIGS. 35 and 36. Further, the motor speed control circuit can be expressed by the block diagram shown in FIG. 37.

A point 1 where the deviation between the speed command and speed output is multiplied by the proportional gain
The specific frequency component of the torque command is blocked by a filter, and torque is applied to the mechanism to control the speed. Here, a method is shown in which speed detection is performed by performing analog speed conversion from position detection pulses without using an analog speed detector, which is a trend in recent years.

The larger the speed control proportional gain K is, the higher the response frequency of the control system can be, allowing high-speed positioning operation. Here, we will investigate the limits of each proportional gain from the transmission characteristics of the mechanism and filter speed detection section, with and without a vibration absorber.

The dynamic model (FIG. 35) without a vibration absorber will be explained.

Motor rotor inertia Jr (kgm2), weight mass m (kg), equivalent torsional rigidity k seen from the motor rotation axis
o (Nm/rad), motor rotor angular velocity ω, (r
ad/S), angular velocity ωr' (rad/s) seen from the motor rotation axis at the end of the arm, motor generated torque T (
If the viscosity effect is ignored because it is minute, the equation of motion shown in equation (3) is established, and the torque and speed (open loop) transmission characteristics shown in equation (4) are obtained in the frequency domain (S domain).

kθ S2+□ When the viscosity term is taken into account, the characteristic is expressed by equation (5).

Here, the damping ratios ξl and ξ2 are small.

This characteristic is the characteristic shown in Fig. 38, Anti-resonance angular frequency: ω^1=kB/mQ”.

Since the viscous effect is minute, the peak at the resonance point 2 anti-resonance point is sharp. Here, in order to increase the S/N ratio in the response frequency band of the command value, a low-pass filter having the characteristics shown in equation (6) (having the characteristics shown in FIG. 39) is used.

In addition, when the speed detection section converts the position detection pulse of the position detector such as the encoder into analog with the F/V converter, (7)
The transfer characteristic shown in the equation (see FIG. 40) is shown, and a significant phase delay occurs. (For example, see Figure 4 of the Journal of the Robotics Society of Japan, Vol. 7, No. 3, pp. 254-259.) G d (s) = s e-8L = (7) where e
-3L is a dead time element that approximates the phase delay characteristic of the speed detection section.

From the above, the transfer characteristic between A and 8 of the drive device control circuit in FIG. 37 becomes as shown in FIG. The gain margin of the transfer characteristic obtained is the maximum value that can be set for the proportional gain.

Next, we will describe the characteristics of the drive/load combination equipped with the vibration absorber. Figure 36 shows the dynamic model of Shikotsu. The vibration absorbing device was equivalently modeled using additional inertia (inertia Ja, rotational angular velocity ω,') and a torsion spring (spring constant θ). The operating equation and torque/speed transmission characteristics are shown in equations (8) and (9).

...(8) Here, k B / m (12 = kθ, / H - = (1
0), the denominator of equation (9) is S"=k6/mQ", kθ(Jr+Ja+
mR2)/ J r m Q ”, (9)
The equation is simplified as equation (11), and when the viscosity term is added, it is expressed as equation (12), which becomes 5...(
11) Here, the damping ratio ξ3. ξ number) ξl and ξ2.

The torque/speed transmission characteristics are as shown in Figure 42.

Anti-resonant angular frequency ω^z (=, r "7ji'mQ"), resonant angular frequency ω, z (= to θ(Jr+a+vIIQ")/
JraaQ2) and without vibration absorber (3rd
8), the direction of the resonant angular frequency is slightly higher, J:, (J
a < J r + m, Q 2), and a reduction in the resonance magnification is observed. Due to the reduction in the resonance magnification, the transfer characteristics between A and B in FIG. 37 become as shown in FIG. 43, and the gain margin increases significantly compared to the case shown in FIG. 41.
The proportional gain can be significantly increased. From the above, it has become clear that the high damping properties of the vibration absorber have the effect of stabilizing the control system and increasing the response frequency. In addition, since the vibration absorbing device of the present invention has a configuration in which a weight is attached to a cantilever beam, the natural frequency adjustment shown by equation (8) can be easily performed by changing the mass and mounting position of the weight. .

Next, the structure and operating principle of a vibration absorbing device that enables damping in multiple directions will be explained using FIG. 44.

In Fig. 44, a mounting base 12 is attached to a vibrating body that vibrates in two directions indicated by arrows 7a and 7b, and a weight 1 is attached to the tip on two sides of the mounting base.
6 is provided, and the parallel force and rotational moment generated at the base of the high damping member by the vibration of the weight 16 are transmitted to the vibrating body via the mounting base, and generates a vibration damping effect. Here, the high damping member is constituted by, for example, a rubber 15 inserted between metal plates 14a and 14b, and the spring constant is mainly determined by the metal plates 14a and 14b, and the viscosity constant C is mainly determined by the rubber 15.

Next, the configuration and driving method of a vibration absorbing device that realizes a high-performance vibration absorbing device using a piezoelectric material without performing calculations will be explained using FIGS. 45 to 54. FIG. 45 shows a perspective view of a vibration absorbing device using a piezoelectric body, FIG. 46 shows the structure of the piezoelectric body, and FIG. 47 shows an equivalent circuit when force is applied at both ends of the piezoelectric body in FIG. 46, Fig. 48 shows the frequency characteristics of the impedance of the piezoelectric material shown in Fig. 46, Fig. 49 shows the mechanical model of the vibration absorbing device shown in Fig. 45, and Fig.
The figure shows an equivalent circuit of the vibration absorbing device of Fig. 49, and
- The figure shows the drive circuit of the vibration absorber of Fig. 45, and
Fig. 2 shows the temporal change in the piezoelectric current when the drive circuit shown in Fig. 51 is driven, and Fig. 53 shows the mechanism transmission characteristics of the drive device/load combination shown in Fig. 34 when the vibration absorber shown in Fig. 45 is installed. 54 shows the A-B transmission characteristic when the drive device/load combination shown in FIG. 34 is driven by the drive circuit shown in FIG. 37 when the vibration absorbing device shown in FIG. 45 is installed.

In the vibration absorbing device shown in FIG. 45, a beam 18 having a weight 16 at its tip is coupled to a mounting base 12 attached to a vibrating body via a piezoelectric body 17. here,
A hollow hole is provided in the piezoelectric body 17, and the mounting base 12. Although an example is shown in which the piezoelectric body 17 is fastened to the beam 18 and the pole 1, the piezoelectric body 17 may be bonded. Vibrating body 11 is indicated by arrow 7
The purpose of this vibration absorbing device is to reduce this vibration. First, the structure and characteristics of the piezoelectric body will be explained using FIGS. 46 to 48.

As shown in FIG. 46, the piezoelectric body 17 has a piezoelectric ceramic thin plate 35 polarized (polarization direction arrow 36) in the deformation direction (arrow 37) and a counter electrode 34a.

34b is provided, and by applying a voltage between the electrodes, it is deformed in the polarization direction (inverse piezoelectric effect), and by applying a force in the polarization direction, an electromotive force is generated between the electrodes (piezoelectric effect). The relationship shown in equation 10) holds between the strain ε and the applied voltage E (V/m).

ε=dsaE...(lO)
daa (m/V) is the piezoelectric constant in the thickness direction of the piezoelectric ceramic. Electric field E (=V/l, V: applied voltage.

The strain can be increased by increasing t (distance between electrodes), but there is a limit in terms of dielectric strength and manufacturing technology in reducing . Piezoelectric bodies are often used that are made by laminating multiple layers of piezoelectric ceramic thin plates with different values. In this case, if the number of layers per layer is n, the amount of strain is (
U) The same applied voltage can be obtained with a large displacement δ proportional to the calendar number.

■ δ= d as −” nt = nd gaV
- (11) is based on the piezoelectric equation that expresses the characteristics of the electric-mechanical coupling system of piezoelectric bodies, ``Basics of Solid Vibration Axis'', Ohmsha (1982), p.
When the equivalent circuit model is obtained as described on pages 138-142, it is as shown in FIG. 47.

Shown as an electromechanically coupled system. Here, φ (N/V
) is a constant indicating electromechanical coupling characteristics and is called a force constant, and the frequency-dependent equivalent spring constant k and equivalent concentrated mass m of the piezoelectric body are shown by equation (12).

...(12) Here h=nt, η””ω/u]Z oD= 5
A r p : density, Ca8D: longitudinal elastic constant, ω: angular frequency, A: electrode area I F1+ F z + external force acting on the end surface.

Here, the applied voltage at the electromechanical coupling part when voltage is applied■
9 Generated force F, current i2 vibration speed ul at both ends.

62, equation (13) holds true.

F=φV, ui+uz=i/φ (13) The frequency characteristics of the electrical impedance between the electrical energy input terminals A and B are as shown in Figure 48, and at the resonant frequency, the electrical impedance is minimum and the maximum current is The current flows and becomes electrically pure resistance.

It vibrates at a high vibration speed according to equation (13).

Next, Figures 49 to 49 show a model and driving method when the piezoelectric body described above is attached to the vibration absorber shown in Figure 45, and the vibration absorber is applied to the drive device/load combination shown in Figure 34. This will be explained using FIG. 54. The drive device shown in FIG. 34 is a rotating type, but when handled on the piezoelectric body axis,
The dynamic model is shown as shown in FIG. Since the purpose of this vibration absorbing device is to reduce the vibration of a specific frequency of the drive device/load combination, the drive device, piezoelectric body, and beam + weight were each expressed as a -degree-of-freedom model. This model is expressed as an equivalent circuit as shown in FIG. The excitation force applied to the drive device is shown as a force F applied to the open terminal of the mechanical system, and a (vibrating) current j flows into the electrical system, charging the equivalent capacitor element of the piezoelectric body, and the equation ( 1
The electromotive force Vo shown in 4) is generated. From now on, if an anti-phase current that cancels the oscillating current flowing into the piezoelectric body is caused to flow from the piezoelectric body side, vibration can be reduced.

When driven at the two electrical and mechanical resonance points where the natural frequencies of the vibrating body and vibration absorbing device match, as shown in (8), the equivalent capacitor Ci between ED and the parallel section in Fig. 50 Electrical impedance Ze is expressed by equation (15), and (
Since the relationship in equation 16) holds true, it becomes a pure resistance.

jωCa φ2 jωC4φ2 Here, Z m'' is the mechanical impedance of the mechanical system in FIG. 50 and is expressed by equation (17).

Here, if A-8 is short-circuited via a resistor R' that is sufficiently larger than the resistor R as shown in FIG. Equation (18)) flows in the opposite direction to when A and B are open.

Here, the short circuit is made through the resistor R'(>R) because the discharge current is mostly discharged from the terminal E in the direction of the terminal C rather than in the direction of the terminal A in FIG. 50. Therefore, (
By discharging at a period twice the time constant Td of C a R in equation 18), it is possible to discharge an anti-phase current I that is 95% of the oscillating current i, and then 3 T a
By opening terminals A and B at a period of l max
can be charged to 95%. Therefore, a high voltage FET 39 is connected between both terminals A and B of the piezoelectric body through a resistor R'.
As shown in FIG. 52, the piezoelectric current is attenuated and the vibration speed is decreased by connecting the pulse generator 38 to a 6T-period rectangular pulse with, for example, 3T: high level and 3Ti: low level. (i/φ) is attenuated and vibrations are suppressed. In FIG. 52, it is possible to flow a damping current having an envelope having an opposite phase to the oscillating current i by the above-described charging/discharging driving method. When such a piezoelectric type vibration absorbing device is applied to the drive device/load combination shown in Fig. 34, the mechanism transmission characteristic G becomes as shown in Figs. 42 to 53 using a vibration absorbing device using rubber. The A/B transfer characteristic shown in FIG. 37 becomes the result shown in FIG. 54, the gain margin becomes large, and the proportional gain K shown in FIG. 37 can be increased. Therefore, by using a driving method of a vibration absorbing device using a piezoelectric body, high-performance vibration reduction is possible without performing calculations, and high-speed positioning of the driving device/load combination becomes possible.

〔Example〕

The first embodiment of the present invention will be explained below with reference to FIGS. 2 to 15 and FIGS. 44, 45, 51, and 52. This embodiment has particularly low attenuation characteristics. Drive device with
The configuration of the vibration absorbing device of the present invention, which is useful for reducing vibration of a load combination, will be described.

Fig. 2 is a perspective view of the vibration absorbing device of the present invention, and Fig. 3 is a perspective view of the vibration absorbing device of the present invention.
Figure 4 is an explanatory diagram of the high damping member removed body of Figure 3, Figure 5 is an explanatory diagram of the mounting part of Figure 3, and Figures 6 to 8 are diagrams of Figures 3 to 5. The relationship diagram between the bending primary natural frequency of the vibration absorbing device and parameters such as dimensions of the vibration absorbing device shown in the figure, Figures 9 and 10 are diagrams of the damping characteristics during impact excitation of the device shown in Figures 4 and 3, Fig. 11 is an explanatory diagram of a vibration absorbing device constructed using a bimorph type piezoelectric element, Fig. 12 is a simplified mechanical model diagram of the vibration absorbing device of Fig. 11, and Fig.
An equivalent circuit model diagram showing the configuration of the vibration absorbing device shown in the figure as an electromechanical coupling system, and FIG. 14 is an electrical system terminal A of the equivalent circuit model of FIG. 13.

Mechanical system model diagram when electrical system equivalent mechanical system is replaced when B is open, Figure 15 is a perspective view of the three-way vibration absorber, Figure 44 is a perspective view of the two-way vibration absorber, and Figure 45 is the thickness vibration. FIG. 51 is a perspective view of a vibration absorbing device using a piezoelectric element, FIG. 51 is a drive circuit diagram of the piezoelectric vibration absorbing device, and FIG. 52 is a diagram illustrating the driving principle of the piezoelectric vibration absorbing device.

First, the basic configuration and characteristic examples of a cantilever beam type vibration absorbing device will be described using FIGS. 2 to 10.

First, the basic configuration of the present vibration absorbing device will be explained using FIG. A mounting base 12 made of a polyhedron attached to the vibrating body 11
Metal plates 14a and 1.4b are connected to each other via a shim metal plate 13. Metal plate 14a.

A high damping member 15 made of a rubber plate or the like is provided between the parts 14b and 14b.
are inserted, and weights 16a and 1.6b are bolted to both sides of the metal plates 14a and 14-b. Compared to the cantilever type vibration absorber shown in the second conventional example, this vibration absorber has a metal plate attached to a mounting base made of a polyhedron. (2) Since the weights 16a and 16b are screw-fastened to each other, there is an advantage that there is no need to use an adhesive to install a high damping member such as rubber.

Further, the weight 16 is provided with a counterbore hole and a counterbore hole to facilitate connection of the weight. This vibration absorber uses the fact that the bending stiffness in the direction of arrow 7 is significantly lower than the stiffness in other directions to adjust the primary natural frequency of bending in the direction of arrow 7 to the vibration body 1.
By matching the resonant frequency of 1, the vibration of the vibrating body 11 can be reduced. Next, the characteristics of this vibration absorbing device will be described. In Figure 3, the metal plate thickness is 2t+
, high damping member, shim metal plate thickness t2. The metal plate protrusion distance Qo, the metal plate width b, and the weight (heavy electric w) attachment distance Q. For comparison, the case without high damping member and shim plate (
The natural vibration characteristics and damping characteristics were investigated for the case of only the mounting base (Fig. 4) and the case of the mounting base (Fig. 5). Phosphor bronze plate was used as the metal plate and shim plate, and butyl rubber (hardness 5) was used as the high damping member.
5. The relationship between the natural frequency and each shape parameter when N5rIIn) is used is shown in FIGS. 6 to 8. In the figure, the case with the high damping member (Fig. 3) is shown by a solid line, the case without the high damping member (Fig. 4) is shown by a broken line, and only the mounting base (Fig. 5) is shown by a dashed line. Figure 6 shows the relationship between the natural frequency and the thickness of the metal plate when a 0.55 kg weight is provided at the center of the overhang. When the metal plate thickness is small, the stiffness of the high damping member is not significantly smaller than the rigidity of the metal plate, so when installing the high damping member, an improvement in the natural frequency due to the increase in rigidity can be seen, but when the metal plate thickness is 5I [ At 111 or more, the rigidity of the high damping member is significantly smaller than the rigidity of the metal plate, so there is almost no difference in the natural frequency depending on the presence or absence of the high damping member. In addition, the mounting base (
Figure 5) Natural frequency (depending on mounting bolt shear rigidity, etc.)
When is low, as shown in Figure 1 (2t1≧8++!11), even if the thickness of the metal plate is increased, the natural frequency of vibration absorption Vi[ does not increase. By increasing the rigidity of the mounting base, it is possible to further increase the natural frequency of the vibration absorbing device when the metal plate thickness is 8 m or more. FIG. 7 shows the relationship between the natural frequency of the present vibration absorbing device and the mounting position of the weight, and it shows that the natural frequency tends to increase by arranging the weight near the mounting base. FIG. 8 shows the relationship between the natural frequency of the present vibration absorbing device and the weight of the weight, and it shows that the natural frequency tends to decrease as the weight of the weight increases. From this, the present vibration absorbing device can adjust the natural frequency by changing the weight of the weight and the mounting position of the weight even when the thickness of the metal plate and the thickness of the high damping member are constant.

This has the advantage of making it easier to deal with variations in the natural frequency due to machine differences in the vibrating body. Next, Figure 4,
Figures 9 and 10 show the time change in the weight acceleration when the weight is struck and excited (white arrow in the figure) in Figure 3. from now,
When there is no high damping member, the damping ratio ξ: 0.065, but when there is a high damping member, the damping ratio ε = 0.221
This results in significantly higher attenuation.

In order to further enhance the damping effect, FIG. 11 shows an example in which a bimorph piezoelectric element is used as the high damping member. The bimorph piezoelectric element has white arrows (19) on both sides of the elastic body 18.
a, 19b) piezoelectric elements 17a, 17b polarized in directions
are pasted together, and a hole is made at the joint with the mounting base 12, and the mounting base 12 is fastened with bolts.

A weight 16 is installed in the bimorph type piezoelectric element through an insulating layer (not shown) in the same manner as shown in FIG. The polarity and electrode terminal configuration are as follows.

FIG. 12 shows a mechanical model of a vibration absorbing device constructed using the bimorph piezoelectric element shown in FIG. 11. Since the piezoelectric element is shown as a model with mass at both ends of the spring constant, it is shown as a model in FIG. 12 in which a weight and a mounting base equivalent mass are added to that mass. The equivalent circuit model of a transversely vibrating bimorph piezoelectric element is shown in Figure 13, and the voltage generated between the electrode terminals due to the vibration acting force F (= F e''') of the vibrating body is shown in Figure 51. By opening and shorting the circuit, a vibration damping current with a phase opposite to the vibration current i as shown in Fig. 52 is applied to the piezoelectric element, thereby producing a significant vibration damping effect.The equivalent circuit model shown in Fig. 13 When the electrical system is rewritten into an equivalent mechanical system model, it becomes as shown in Fig. 14.The oscillating current i flowing through the piezoelectric element due to the oscillating force F is expressed by equation (19).

i = (ra1-62)φ...(1
9) F Zs...(20) Here, Table 1 Zs'=: 4mM.

By configuring the vibration absorbing device as described above, active vibration absorption becomes possible, so that vibration can be rapidly reduced as shown in FIG. 52. In the circuit shown in Fig. 51, the period for intermittently opening and shorting both terminals of the piezoelectric element is a time constant τ determined from the capacitance Ca and resistance R of the piezoelectric element.
(=CdR), the damping current amplitude can be 95% or more of the vibration current, and a sufficient damping effect can be obtained. Table 1 shows examples of characteristics of Cm and R values.

From this, 3 = 4.29μs, 233kHz
This shows that sufficient charging and discharging is possible by turning on and off the electrodes at a frequency of It is possible to generate a vibration damping current 1, which enables rapid vibration damping of the vibrating body. In this way, the case where a bimorph type piezoelectric element is used as a high damping member is shown, but it can also be realized by providing a thickness vibrating type piezoelectric element 17 between the metal plate 18 and the mounting base 12 as shown in FIG. . In addition, by installing cantilever beams on multiple sides of the mounting base 12 made of a polyhedron, a two-way vibration absorption device (the 44th
), a three-way vibration absorbing device (FIG. 15) can be realized. Normally, the vibrating body 11 often vibrates with large amplitude in a vibration shank mode in multiple directions at the resonance frequency, so by installing a multi-directional vibration absorbing device, it becomes possible to effectively damp the vibration of the vibrating body.

Next, FIG. 1 shows a second embodiment of the present invention.

This will be explained using FIG. 16 to FIG. 29 and FIG. 37. This embodiment describes the structure of a drive device/load combination body equipped with the vibration absorbing device described in the first embodiment of the present invention, and the servo characteristics of the drive device. FIG. 1 is a perspective view of a uniaxial arm device equipped with the vibration absorbing device of the present invention;
17 is a vertical cross-sectional view of a direct drive rotary electric motor, and FIGS. 18 to 20 each show the control circuit shown in FIG. 37, and the motor speed of the uniaxial arm device shown in FIG. 16. C-B transfer characteristic diagram when applied to control, A
-B transfer characteristic, A-C transfer characteristic diagram, Figure 21 is 1
6 is a vibration mode diagram at the primary resonance frequency fxr of the uniaxial arm device shown in FIG.
4 and 25 show the servo system A-F3 when the natural frequency of the vibration absorbing device matches the uniaxial arm device resonance frequency f1r and anti-resonance frequency fl^ in the uniaxial arm device shown in FIG. 22 or 23. Figures 26 to 29 are configuration diagrams showing the installation of multiple vibration absorbers to the single-axis arm device shown in Figure 16, and Figure 37 is a drive device speed control block diagram. .

First, with reference to FIG. 16, the structure of the drive device/load combination will be described, with reference to the effect of mounting the vibration absorber in the description of this embodiment. A stator 1b of a direct drive rotary electric motor 1 used in this embodiment as an example of a drive device is coupled to the base 2, and an arm 3 having a load 4 at its tip is coupled to the rotor 1a. An impact member mounting member 5 to which an impact member 9 is connected is integrated with the lower part of the rotor 1a by welding or the like. A stopper 10 is coupled to the base 2, and an overrun sensor (not shown) detects the overrun when the arm performs an overrun operation within a limited movement range of approximately 300 degrees. When the collision member 9
The stopper 10 comes into contact with the stopper 10 to forcibly stop the operation. 1st
The structure of the direct drive rotary electric motor 1 shown in FIG. 6 will be explained using FIG. 17. The motor stator 1b and rotor 1a are supported by a pair of bearings, and the motor stator 1b has a coil wound around a pole 21 made of a magnetic material, and a permanent magnet 20 is wound around the inner circumference of the motor rotor 1a. A magnetic circuit is formed between the motor stator and rotor by energizing the coil, and according to Fleming's left hand rule, the motor rotor 1a rotates relative to the motor stator 1b (
A force is generated in the direction arrow 8). 18 to 21 show the servo characteristics when the motor 1 of the single-axis arm device shown in FIG. 16 is driven by the motor control circuit shown in FIG. 37.
This will be explained using figures. In FIG. 37, it is assumed that the speed detection section characteristic G- is shown in FIG. In the motor control circuit shown in Fig. 37, the resonance characteristics of the mechanism can be known by knowing the mechanism transfer characteristic G. However, in this shaft arm device, an analog speed detector is not used and the position detection pulse is Since the configuration detects speed by converting voltage (F/V conversion), the transfer characteristic between C and B
We decided to investigate the mechanistic properties by detecting G4. Figure 18 shows its characteristics, flr,
Two resonant frequencies far are within the measurement frequency range. The frequency Jc at which the phase is 180° or less is j
■ku fc〈f2. Therefore, the secondary resonance peak is blocked by a notch filter, and in order to increase the overall S/N ratio, a filter (characteristic diagram Figure 20) that also has a low-pass filter effect with a band frequency of T -1 is used. When used as a filter characteristic block OF in the circuit shown in the figure, the AB transfer characteristic GFGG- becomes as shown in FIG. 19, and the gain margin at the frequency fC where the phase is -180° is shown as shown in the figure.

This gain margin indicates the maximum value that can be set for the proportional gain in the speed control circuit shown in FIG. 37, and by increasing this value, it becomes possible to realize high-speed follow-up operation to the command of the uniaxial arm device. However, in FIG. 19, there is a problem that this gain margin is narrowed by the resonance peak at frequency flr and does not have a very large value. Therefore, we will consider mechanically reducing the primary resonance of the shaft arm device using a vibration absorber. FIG. 21 shows the vibration mode in the rotational direction of the uniaxial arm device at the primary resonance frequency flr. At the primary resonance frequency, the electric motor causes rotational direction vibration accompanied by the fall of the arm support member. Furthermore, since the position detector exhibiting the characteristics shown in FIGS. 18 and 19 is provided at the bottom of the electric motor as shown in FIG. 17, a vibration absorbing device is attached to the collision member mounting member at the bottom of the electric motor. I will consider that. Figures 22 and 23
The figure shows a motor rotation direction vibration absorbing device provided on a collision member mounting member so that the longitudinal direction of the vibration absorbing device coincides with the radial direction, upward direction, and downward direction. Figure 24 shows the transmission between A and 8 in Figure 37 when the natural frequency of the vibration absorbing device in Figure 22 or Figure 23 is matched with the primary resonance frequency f lr and the primary anti-resonance frequency of the uniaxial arm device. It shows the characteristics. The broken line in the figure shows the characteristics without the vibration absorber. From now on, the vibration absorbing device has the effect of reducing the resonance magnification and smoothing the phase change by matching its natural frequency with the resonant frequency or anti-resonant frequency, especially when the natural frequency is matched with the resonant frequency (24th (Fig. 19) makes it possible to significantly increase the gain margin compared to the case without a vibration absorber (Fig. 19). By the way, a 700III11 arm (6 kg load at the tip) using a D-D motor
When tested using the vibration absorbing device shown in Figure 2, the second
6clB (double) in the configuration shown in Figure 2, and 6clB (double) in the configuration shown in Figure 23.
A gain margin increase effect of 4 dB (1,4; times) was observed. Furthermore, by using the piezoelectric vibration absorbers shown in Figs. 11 and 45, it is possible to significantly reduce the resonance magnification at the resonant frequency, and furthermore, increase the gain margin (increase in proportional control gain). This makes it possible to achieve high-speed positioning operations.

In addition, if the anti-resonance frequency is below the speed control system bandwidth (matching the low-pass filter cut-off frequency T-1),
This causes a drop in the gain in the closed-loop characteristics of the speed control system (a significant delay in the phase in the closed-loop characteristics), which causes the problem of narrowing the bandwidth of the position control system. , it becomes possible to expand the position control system bandwidth. This problem will be explained using an example in the description of the third embodiment.

Next, a second example of damping vibration using multiple vibration absorbers will be explained.
This will be explained using FIGS. 6 to 29.

FIG. 26 shows a device in which two motor rotational direction vibration absorbers are provided on the collision member mounting member of the uniaxial arm device. By matching the natural frequency of each vibration absorbing device with the anti-resonant frequency f1^ and the resonant frequency frr of the uniaxial arm device, the transfer characteristic between A'B in Fig. 37 becomes as shown in Fig. 27, and the anti-resonant It is possible to reduce the resonance magnification of both the resonance point and the resonance point. Furthermore, when there are multiple resonance points, the vibration reduction effect is increased by using a vibration absorber that matches each resonance point. Further, by providing a plurality of vibration absorbers that match a single resonant frequency, it is possible to reduce the resonance magnification at the resonant frequency and increase the gain margin of the control system. In addition, by arranging the vibration absorbing device so that its weight is closer to the center of gravity of the motor than the mounting base, as shown in Fig. 28 and Fig. 1, effective vibration damping can be achieved without significantly increasing the load inertia of the motor. It becomes possible. In FIG. 29, a two-way vibration absorber @ (two rotational directions in the vertical direction) is provided on the top of the electric motor so as to reduce vibrations in multiple directions. With such a configuration, it is possible to suppress resonance that causes large-amplitude vibrations in multiple directions.

Next, a third embodiment of the present invention will be explained using FIGS. 30 to 33. In this sixth embodiment, the vibration absorbing device described in the first embodiment is driven by a plurality of drive devices. As an example of application to a multi-axis mechanism, we will discuss a horizontal hook-jointed direct drive robot. FIG. 30 is a perspective view of a horizontal hook joint type direct drive robot.

Fig. 31 is an A-B transfer characteristic diagram in Fig. 37 of the -axis direct drive motor without this vibration absorption device, and Fig. 32 is the A-B transfer characteristic diagram of the single axis direct drive motor with this vibration absorption device.
FIG. 33 shows a transfer characteristic diagram between D and 8 of the single-shaft direct drive motor when this vibration absorbing device is installed. First, the configuration of this robot will be explained using FIG. 30.

This robot has a configuration in which the first arm 24 and the second arm 26 are directly driven by direct drive motors 23 and 25, respectively, and the wrist shaft (spline shaft) 28 transfers the rotational power of the upper and lower shaft drive motors 30 to the belt. (not shown) to the ball screw shaft 27, and is driven upward and downward together with the bracket 33 connected to the ball screw nut screwed thereto, and further decelerated by the wrist rotation shaft drive electric motor 31. The machine 32 is configured to be rotationally driven via a belt (not shown). The spline shaft 28 is rotatably supported by the bracket 33, and a tool 29 is attached to its lower end. By positioning the tool 29 using the lower shaft drive motors 23 and 25° 30 and positioning the tool 29 using the wrist rotation shaft drive motor 31, it is possible to realize the desired load transfer operation. Become. This transfer operation requires high-speed operation, and in particular requires high-speed operation of the two-axis direct drive motors 23, 25, which are involved in positioning operations in the horizontal plane. To achieve this, it is essential to have a high proportional control gain in the motor drive circuit shown in Figure 37.
It is essential to increase the gain margin in the B-band transfer characteristic. First, FIG. 31 shows the transmission characteristics between A and 8 of the single-axis direct drive motor in the robot shown in FIG. 30 without a vibration absorbing device. From now on, it is the secondary resonance that determines the gain margin, but the speed control system bandwidth setting value (
There is first-order anti-resonance and resonance below the low-pass filter cut-off frequency T-1 (almost the same as the low-pass filter cut-off frequency T-1), and when looking at the closed loop speed transfer characteristic between D and B in Figure 37, there is a significant valley and a significant valley as shown by the broken line in Figure 33. A peak is seen, and the speed control system bandwidth is at the valley frequency in
It is constrained by s. Therefore, we decided to install a vibration absorption device based on the following principles. In other words, (1) installing a vibration absorber with a natural frequency of resonance frequency and anti-resonance frequency that is below the speed control system bandwidth; (2) achieving sufficient speed proportionality by using a notch filter above the speed control system bandwidth; Installing a vibration absorber with a natural frequency of the resonant frequency where no control gain can be obtained If this principle is followed, the frequency fl^ + f1rg in Figure 31
f2. It is necessary to provide a vibration absorption device compatible with r+f3r. Although not shown, each order of resonance indicates the following vibration mode.

(1) Primary: Bending of the first arm f IAI f 5r
(2) Falling of the secondary knee-shaft motor fzr (3) Turning of the upper part of the tertiary knee-shaft motor fsr Therefore, the vibration absorbing device 6a corresponding to the primary mode is provided.

A vibration absorber 6b corresponding to the second-order mode was provided on the upper part of the two-axis motor, and a vibration absorber 6C corresponding to the third-order mode was provided on the second part of the single-axis motor. As a result, the transfer characteristics between A and 8 of the direct drive motor for the negative axis are as shown in Figure 32, and the gain margin is significantly increased compared to the case without a vibration absorber (Figure 31), and as shown in Figure 33. It was found that the bandwidth fB2 in the speed closed-loop transfer characteristic was significantly improved compared to fnl, and it was found that high-speed operation could be achieved by significantly improving the proportional control gain of the speed control system (FIG. 37).

〔Effect of the invention〕

Since the present invention is configured as described above, it produces the effects described below.

(1) By arranging the weight drive device closer to the center of gravity than the mounting base of the vibration absorption device, effective vibration absorption can be performed without significantly increasing the load inertia of the drive device.

(2) By configuring the vibration absorbing device by providing cantilever-like high-damping members on multiple sides of the mounting base, vibration absorption of the vibrating body in multiple directions can be performed simultaneously.

(3) By configuring the high damping member of the vibration absorbing device using a piezoelectric element, and by intermittently opening and closing the terminals through a large resistance at a cycle that is more than twice the time constant, Can perform vibration absorption.

(4) When using this vibration absorption device to absorb vibrations of the drive device/load combination, resonance/anti-resonance frequencies below the drive device speed control bandwidth, and resonance frequencies above the bandwidth that cannot be blocked by a notch filter. By installing a vibration absorber with a natural frequency that matches the vibration mode at the large amplitude position of the vibration mode, it is possible to improve the speed proportional control gain of the drive device/load combination.
It becomes possible to expand the speed control system bandwidth and enable high-speed operation.

[Brief explanation of drawings]

FIG. 1 is a perspective view of a drive device/load combination body equipped with a vibration absorbing device according to an embodiment of the present invention, FIG. 2 is a perspective view of a vibration absorbing device of the present invention, and FIG. 3 is a perspective view of the vibration absorbing device of the present invention. Explanatory diagram, Figure 4 is an explanatory diagram of the vibration absorber shown in Figure 3 without the high damping member, Figure 5 is an explanatory diagram of the mounting base, and Figures 6 to 8 are the natural frequencies of the vibration absorber. Figures 9 and 10 are diagrams of the damping characteristics of the devices shown in Figures 4 and 3. Figure 11 is an explanatory diagram of a vibration absorption device using a bimorph piezoelectric element. Figure 12 is a simplified mechanical explanatory diagram of the vibration absorbing device in FIG. 11, FIG. 13 is an equivalent circuit model diagram of the vibration absorbing device in FIG. 11, and FIG. 14 is an equivalent mechanical system model diagram of the equivalent circuit model in FIG. 13. 15 is a perspective view of the three-way vibration absorbing device of the present invention, FIG. 16 is a perspective view showing an example of the structure of a drive device/load combination, and FIG. 17 is a longitudinal sectional view of the drive device of FIG. 16. Figures 18 to 20 show C when the circuit shown in Figure 37 is used to drive the drive unit/load combination shown in Figure 16.
-8, A-8, and A-C transfer characteristic diagrams. Figure 21 is a vibration mode diagram at the resonant frequency flr of the drive device/load combination of Figure 16. Figures 22 and 23 are A side view of a single vibration absorber installed in the drive/load combination shown in the figure.
Figure 24. FIG. 25 shows the drive circuit A- in FIG. 37 when the natural frequencies of the vibration absorbers match the primary resonant frequency and the primary anti-resonant frequency in FIG. 22 or 23, respectively.
B transmission characteristic diagram, Figure 26 is a side view when two vibration absorbers whose primary resonance frequency and primary anti-resonance frequency match the drive device/load combination shown in Figure 16 are installed, and Figure 27 is a side view of the drive device/load combination shown in Figure 16. Figure 37 is a transfer characteristic diagram between A and B, which is the driver drive circuit for the drive unit/load combination shown in Figure 26, and Figure 28 is a diagram showing the transmission characteristic between A and B, which is the drive circuit for the drive unit/load combination shown in Figure 16, and the weight is driven from the mount in the drive/load combination shown in Figure 16. A side view when multiple vibration absorbers are installed near the center of gravity of the device, Figure 29 is a side view when a two-way vibration absorber is installed on the drive device/load combination shown in Figure 16, and Figure 30 is a side view of the horizontal pot. FIG. 31 is a perspective view of the articulated direct drive robot; FIG. 37 is a drive circuit of the uniaxial direct drive motor without vibration absorption equipment; Fig. 32 shows the drive circuit of the single-axis direct drive motor of the robot shown in Fig. 30;
3 is a diagram showing the transfer characteristics between D and B in FIG. 27, which is the drive circuit for the single-axis direct drive motor of the robot shown in FIG. 30, FIG. 34 is a perspective view of the simple axis arm device, and FIG. Dynamic model diagram without vibration absorber, Figure 36 is Figure 34
Figure 37 is the motor control circuit block diagram in Figure 34, Figure 38 is the frequency characteristic diagram of the mechanical transfer function G of the uniaxial arm device without the vibration absorber in Figure 37, Figure 39 is the mechanical model diagram in Figure 37. is the filter transfer function OF in FIG.
FIG. 40 is a frequency characteristic diagram of the speed detection section transfer function Gd in FIG. 37. No. ′? ! : Figure 41 is the frequency characteristic diagram of the transfer function between A and B in Figure 37, which is the electric motor drive circuit of the axis arm device when there is no vibration absorber, and Figure 42 is the mechanical transfer function of the single axis arm device when the vibration absorber is installed. Figure 43 is a frequency characteristic diagram of the A-B transfer function, Figure 43 is a motor drive circuit for a single-axis arm device with a vibration absorber, and Figure 44 is a perspective view of the two-way vibration absorber. Figure 45 is a perspective view of a vibration absorbing device using a thickness vibration type piezoelectric element, Figure 46 is a perspective view of a thickness vibration type piezoelectric element, and Figure 47 is a dynamic model diagram and equivalent circuit of a thickness vibration type piezoelectric element. Model diagram, Figure 48 is a frequency characteristic diagram of impedance of a thickness vibrating piezoelectric element,
Fig. 49 is a dynamic model diagram when the piezoelectric vibration absorber is attached to the drive device, Fig. 50 is an equivalent circuit model diagram of Fig. 49, Fig. 51 is a piezoelectric vibration absorber drive circuit diagram, and Fig. 52 is a diagram of the piezoelectric vibration absorber drive circuit. Diagram explaining the operating principle of the piezoelectric vibration absorber in Fig. 45, No. 53
The figure shows the frequency characteristic diagram of the mechanism transfer function of the uniaxial arm device in Fig. 34 when the piezoelectric vibration absorber shown in Fig. 45 is installed, and the frequency characteristic diagram shown in Fig. 54.
The diagram shows the motor drive circuit in the uniaxial arm device shown in Fig. 34 when the piezoelectric vibration absorber shown in Fig. 45 is installed.
FIG. DESCRIPTION OF SYMBOLS 1... Direct drive electric motor, 2... Base, 3... Arm, 4... Load, 5... Collision member mounting member, 6... Vibration absorber, 7... Vibration absorber damping direction, 8... Electric motor rotation direction, 9... Collision member, 10... Stopper,
11 Figure 3 Y4 Figure 〆2 #14 Eye Figure 22 Aσ inter-Buddha Masu Ogi 23 Prisoner A・CM Buddha characteristic (Ki, lll Hani number (Ha Cr2-Shirihagiyashi → Life (#1 history) Zhou〃Good Rebellion (Hz) Characteristics of A・σ Listening Ac11 & Il Healthiness Figure 30 #34 Trouble Figure 35 Figure 3 Otsu Figure Dormitory 40 Figure 41 Figure f, 4sEJ 11so Mouth f45 Figure 54-Figure

Claims (1)

[Claims] 1. In order to suppress vibrations of a vibrating body, in a vibration absorbing device coupled to a large-amplitude vibrating part of the vibrating body, one side of a mounting base made of a polyhedron, for example, a cube, is installed on the vibrating body. , one end of the cantilever-shaped high-damping member is fixed to the mounting surface perpendicular to one or more large-amplitude vibration directions, and a weight is provided at the other end of the high-damping member or at an arbitrary position, and the primary bending vibration of the high-damping member is fixed. A vibration absorbing device characterized by matching a natural frequency with a resonant frequency or an anti-resonant frequency of a vibrating body.