JPH03245777A - Driver - Google Patents

Abstract

PURPOSE:To perform high-speed, high-operation of a load body by providing a large-amplitude vibrating section at time of the load body being coupled and a transferred quantity detecting element with vibration absorbers. CONSTITUTION:A motor rotor 5 is coupled with the input axis of a harmonic drive type speed reducer 14, and its axis 15 is coupled with a load body. This construction reduces the motor speed to one nth at the output axis 15, and increases the generated torque of the motor by n times at the output axis 15, in the case of a speed reduction ratio of n, for example. Usually the torsional rigidity of the speed reducer 14 is smaller than the rigidity of a load body, and vibration caused by the torsional rigidity of the speed reducer exerts a baneful influence upon the motor control system. Then a high-damping member 8 and a deadweight 9 are provided on the output axis 15 as a vibration absorber, and a high-damping member 6 and a deadweight 7 are provided in the periphery of a transferred quantity detector 4 as a vibration absorber to damp vibration effectively.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention is directed to a servo control system for an electric motor to enable high-speed, high-precision operation of a multi-axis mechanism driven directly by an electric motor or via a reduction gear. The present invention relates to a drive device having a vibration absorption device that stabilizes characteristics without a significant increase in motor load, and a multi-axis mechanism equipped with the same.

[Conventional technology]

The means to stabilize the unstable characteristics of the motor servo system due to the mechanical resonance of the load body driven by the electric motor are: ■ inserting a filter into the motor control circuit to reduce the driving force command gain; ■ motor/load It can be roughly divided into two types: reducing the gain at the resonance point by reducing the vibration amplitude by installing a vibration absorber (dynamic vibration absorber) at the large amplitude position in the vibration mode of the resonance point of the body-combined body. .

As an example of the former, as described in Japanese Patent Application Laid-Open No. 60-39397, gain reduction at the resonance point is achieved by providing the torque command generation section with a filter having an inverse characteristic to the gain characteristic near the resonance point. This is intended to improve the control gain.

In addition, the latter example is from the Transactions of the Japan Society of Mechanical Engineers, Volume C, Volume 55, 5.
No. 16 (August 1989), pages 2014 to 2021, there are examples of passive vibration absorbers using rubber and weights, and as described in Japanese Patent Application Laid-Open No. 62-228730. In this active vibration absorption device, a piezoelectric element is installed between the outer periphery of a bearing that supports a rotating shaft and a fixed body or a weight, and a vibration damping effect is generated by energizing the piezoelectric element based on the required applied voltage calculation amount. Examples and JP-A-1-1
As described in Publication No. 90271, an example of a bioactive vibration absorption device is described in which a vibration isolator made of a piezoelectric material is provided between a vibrating body and a base, and both terminals of the piezoelectric material are short-circuited via a resistor. Ta.

[Problem to be solved by the invention]

Although the first conventional technology has an excellent gain reduction effect, the notch filter transfer characteristic GF is in the frequency domain (S domain).
Because it is given by the following formula.

As shown in FIG. 18, the central angular frequency ω of the notch filter. There is a phase delay of about 90° at a slightly lower frequency,
Above the center angular frequency of the notch filter, the phase advances by 180°, but below the center angular frequency of the notch filter, the phase delay of the open-loop characteristic of the servo system exceeds 180°, making the servo system unstable, and the band There is a possibility that the width becomes narrow and the control gain cannot be increased, and in particular, it is not an effective countermeasure against mechanical resonance points in the low frequency range where the open loop gain is high.

In the second prior art, when a cantilever beam is rotated by a step motor, a passive vibration absorber made of rubber and a weight is provided at the tip of the beam, which has a large vibration amplitude, to absorb vibration. In direct-drive multi-axis mechanisms such as direct-drive robots that require high-speed operation, installing a vibration absorber at the end of the arm increases the load inertia of the motor, making high acceleration/deceleration operations difficult. There is.

The third conventional technology is for damping vibrations of rotating shafts.

Piezoelectric elements are arranged radially around the outer circumference of the rotating shaft support bearing, and voltage is applied to each piezoelectric element by a piezoelectric element drive circuit based on a piezoelectric element applied voltage command calculated by a magnetic bearing control algorithm, resulting in a complicated configuration. be.

In addition, the fourth conventional technology maximizes the energy consumption in the resistor and minimizes the vibration energy of the vibrating body by always short-circuiting the piezoelectric body through a resistor and configuring an RLC series resonant circuit. However, there is a problem that the damping effect is small.

The purpose of the present invention is to perform effective vibration suppression with a simple configuration without significantly increasing the load inertia of the motor, stabilize the servo system of the motor, and improve the high-speed, high-precision positioning of a multi-axis mechanism. It's about doing.

[Means to solve the problem]

In order to achieve the above object, the present invention takes the following measures.

In order not to significantly increase the load inertia of the electric motor, the vibration absorbing device is set so that its center of gravity has a very small deviation from the axis of movement of the electric motor (direct motion) or the central axis of movement (rotation).

[Effect]

With this configuration, the drive device of the present invention achieves the above object through the following actions.

First, using FIGS. 20 to 26, the torque and speed transmission characteristics of the uniaxial swing arm drive device will be described with and without a passive vibration absorption device using a high damping member (rubber) and weight. Fig. 20 shows the appearance of the single-axis swing arm drive device, and Figs. 21 and 22 show the structure without the passive vibration absorption device.
The model of the single-axis swing arm drive device with the
Figure 3 shows the torque and speed transmission characteristics of the single-axis swing arm drive device with and without the passive vibration absorption device, Figure 24 shows a control circuit block diagram of the single-axis swing arm drive unit, and Figures 25 and 25 show the control circuit block diagram of the single-axis swing arm drive unit. FIG. 26 shows the A-B transfer characteristics of the control circuit shown in FIG. 24 without and with the passive vibration absorber. In FIG. 20, the swing drive device 5 is composed of a direct drive electric motor, and an arm 69 is directly connected to the rotating part of the electric motor, and a weight 7 is attached to the tip of the arm.
0 is connected. Here, the motor rotating part inertia Jr (kg body), the weight mass m (kg), and the equivalent torsional rigidity seen from the motor rotating shaft are θ (Nm/rad), and as shown in FIG. 21, the motor rotating part angular velocity ( 11r
(rad/s) T-arm tip angular velocity ωr' (r
ad/s) and the motor generated torque T (Nm, ), and if the damping effect is ignored because it is minute, the equation of motion shown in equation 0 will be obtained, and the torque/speed transmission characteristics shown in equation 0 will be obtained in the frequency domain (S domain). can get.

This characteristic is shown by the solid line in FIG. 23, and the anti-resonance angular frequency ω^x=J positive 7; Σ negative 11-. Since the damping effect is minute, the peak at the resonance point 9 anti-resonance point is sharp. Further, the characteristics of a drive device control circuit including a low-pass filter having a second-order lag characteristic shown in FIG. 24 are investigated. ω so that the filter cutoff frequency and resonance point gain can be reduced
^1. Let us consider a case between ωr1. The open loop characteristic (between A and B) excluding the proportional gain is shown in FIG. 27, and the phase exceeds 1180° near the resonance point. The gain surplus at that time is the maximum value that can be set for the proportional gain, and there is a problem that the proportional gain cannot be increased very much. Therefore, in FIG. 20, a case will be considered in which a passive vibration absorbing device shown by a broken line is installed on the direct drive motor 5. It is assumed that the passive vibration absorbing device is composed of, for example, an annular high-damping member (for example, rubber) and an annular weight provided on the outer periphery. A passive vibration absorption device composed of a high damping member and a weight is added to the dynamic model of FIG. 21 as shown in FIG. 22. In this case, the torsional stiffness of the passive vibration absorber is θ (Nm/rad).

Inertia Ja (kgmF), angular velocity ωr' (rad
/s), the equation of motion and torque-speed transmission characteristics of the swing drive device are shown in equations (1) and (2). (The attenuation term in the ■ Valor 3 equation has been ignored to simplify the equation.) Here, when kθ / mQ'' = θ and /Ja, the denominator of the ■ formula is S'' = - θ / m Q 2 - θ (J −+ J a+ mΩ2) / J r m Q ”, and Equation 0 is simplified as Equation 0.

mΩ2...■, and the torque/speed transmission characteristics are as shown in the broken line in Figure 23, anti-resonance point angular frequency ω^2 (=41ko1)]π7]-) resonance point angular frequency ωr2 (= kθ( Jr+JamR2)/JrmQ”
), which shows an improvement in the resonance point frequency and a reduction in the resonance magnification due to the high damping properties of the vibration absorber. Since this vibration absorbing device is installed coaxially with the direct drive device 5 as shown in FIG. 20, the increase in load inertia is small and kθ1(
If kθ is Ja(J,+m Q ”), the change in the resonance point angular frequency depending on whether there is a vibration absorber or not is small.

On the other hand, since the resonance magnification is reduced due to the high damping property of the vibration absorber, the A-B transfer characteristic of the control system shown in Figure 24 becomes as shown in Figure 26, which is compared to the case without the vibration absorber. The maximum value of the speed proportional gain can be increased, the torsional rigidity θ can be increased, the resonance frequency can be improved, and highly accurate positioning can be achieved with a short settling time when the arm operates at high speed.

In order to actively damp the vibration of a drive device with a simple configuration, we will describe the operating principle of an active vibration absorber constructed using a piezoelectric force material and a weight using Figs. 27 to 35. . Figure 27 shows the structure of the piezoelectric body, Figure 28 shows the model and equivalent circuit of the piezoelectric body, Figure 29 shows the frequency characteristics of the impedance of the piezoelectric body, and Figure 30 shows a vibration absorbing device including the piezoelectric body. 31 shows a model of the vibration absorbing device shown in FIG. 30, and FIG. 32 shows an equivalent circuit of the model shown in FIG. 31.

As shown in FIG. 27, the piezoelectric body has electrodes attached to both end faces in the polarization direction of a piezoelectric ceramic thin plate polarized in the direction of the arrow, and expands in the polarization direction by applying a voltage in the polarization direction. The relationship of equation 0 is established between the strain ε and the applied electric field E (V/m).

1==dsδE...■
daa (m/V) is the piezoelectric constant in the thickness direction of the piezoelectric ceramic. Electric field E (=V/l: V: applied voltage, t
Although the strain can be increased by increasing the distance between the electrodes (distance between A piezoelectric body consisting of multiple laminated layers of different piezoelectric ceramic thin plates is often used. In this case, when the number of layers is n, the amount of strain is expressed by the equation 0, and a large displacement δ proportional to the calendar number can be obtained with the same applied voltage.

■ δ=daa-y nt=ndxsV
-■ Piezoelectric materials have an inverse piezoelectric effect that deforms when voltage is applied, and a piezoelectric effect that generates voltage when pressure is applied. Physical fucaustics based on piezoelectric equations Volume 1 Bart A.
Academic Press.

(1964) pp 234-237 (Phys.
icalAcoustics, I-Part A, A
Academic Press (1964) pp234
When an equivalent circuit is obtained as described in 237), it is shown as an electromechanical coupled system as shown in FIG. Here, φ (N/V) is a constant that indicates the electromechanical coupling characteristics and is called a force constant, and the equivalent resiliency constant of the piezoelectric body is 9. The equivalent concentrated mass m is 0.
It is shown by the formula.

Here, h=nt, ηD=ω/2Σd.

Z o = lA, p: density, Cs%: longitudinal elastic constant, ω: angular frequency, A: electrode area.

Here, the applied voltage V at the electromechanical coupling part when voltage is applied
9 When the generated force F, the current i, and the vibration speed uz+uz, the [phase] formula is established.

F=φV, u1+uz=i/φ −ta When the ends of the piezoelectric body are free, the equivalent circuit is short-circuited, and when external forces pi and Fz are applied to both end faces, the equivalent circuit is shown as shown in FIG. 28. The frequency characteristics of the impedance between input side terminals A and B are shown in Figure 29, where the maximum current flows at the resonance point (minimum impedance), it becomes electrically pure resistance, and the vibration speed increases due to the [phase] equation. It vibrates.

Next, the principle of operation when the piezoelectric body described above is used for vibration damping of a rotary drive device will be explained. In FIG. 30, while the drive device rotor 5a vibrates in the direction of the arrow, a piezoelectric body 40 provided on a protrusion 5b provided integrally with the first rotor is bolted to an annular weight 9. Although a plurality of piezoelectric bodies 40 can be provided in the circumferential direction, the principle of operation in the case of one piezoelectric body will be described. When the mechanical model of the vibration absorbing device in the case of one piezoelectric body is considered on the axis of the piezoelectric body and expressed as shown in FIG. 31, the equivalent circuit model seen from the piezoelectric body is shown as shown in FIG. 32. Here, the purpose is to apply an external force F to the rotary drive device.
The purpose is to make the vibration speed u1 very small when the From the equivalent circuit in Figure 32, piezoelectric terminals A-B
If an external force F acts on the piezoelectric body when the gap is opened, a current j flows through the piezoelectric body, charging the equivalent capacitor inside the piezoelectric body, and an electromotive force shown by equation C is generated between A and B.

When driven at an electrical/mechanical resonance point, the electrical impedance Ze between the EDs is expressed by the equation 0, and is a pure resistance.

MI K*/φ21/C* ...0 Here, ZII is mechanical impedance (force/vibration speed)
, and is explained by replacing it with the Ippaku Yude spring-mass system isometrically. At the resonance point, Equation 0 holds true.

Here, as shown in FIG. 33, the resistance R' is sufficiently larger than R.
(Connection between switch FG in the figure to short-circuit between A and B via)
When discharged, the current flows in the opposite direction to when it is open (connection between switch FH in the figure) with the magnitude shown in equation 0.

”　l　mkl　e　C:’　　　　　　　-,.

Here, the short circuit through the resistor R' where R<R' is
This is because the discharge current is mostly discharged from terminal E in the direction of terminal C rather than in the direction of terminal A in FIG.

Therefore, by discharging at a cycle that is, for example, three times the time constant T4 of CaR, it is possible to discharge an anti-phase current that is 95% of the oscillating current. Therefore, as shown in Figure 34, 3
T6 is the charging time, 3T1 is the discharging time, and the timer 42. A drive circuit (third
By repeating charging and discharging using the piezoelectric material (Fig. 3), feedback control is performed so that the current i flowing through the piezoelectric body becomes zero, and the vibration speed u (=i/φ) of the mechanical system approaches zero with time. Vibration is suppressed. Furthermore, by charging and discharging a plurality of piezoelectric bodies with different phases, the damping current and the absolute value of the oscillating current can be made to substantially match each other in FIG. 34, thereby enabling highly efficient damping. When such an active vibration damping device is applied to the rotary drive device shown in Fig. 30, the transfer characteristic between A and B in Fig. 24 is improved as shown in Fig. 35, and the phase lead/lag due to mechanical resonance is eliminated. The proportional gain can be increased, and the load body can be operated at high speed and with high precision.

〔Example〕

A first embodiment of the present invention will be described below with reference to FIGS. 1 to 7. In this example, the power generated by the electric motor is
A vibration absorbing device consisting of a high damping member (e.g. rubber) and a weight is installed in advance in the drive device itself, which transmits power directly or indirectly to the load through a power transmitting member such as a speed reducer. This aims to reduce the vibration of the quantity detection section and construct a stable motor control system. Fig. 1 shows the structure of an external cylinder rotor type direct drive motor, Fig. 2 shows the structure of an internal cylinder rotor type direct drive motor, and Fig. 3 shows the structure of a reduction gear type rotary drive device. Figure 4 shows a shaft that moves in a straight line and makes two rotations, and its drive device.
FIG. 5 shows a vibration absorbing device for reducing torsional and radial vibrations of the rotary drive device, FIG. 8 shows a vibration absorbing device for reducing torsional and axial vibrations of the rotary drive device, and FIG. 5 shows a variable weight structure of the weight of the vibration absorbing device shown in FIG. As mentioned in the previous chapter, a vibration absorbing device consisting of a passive high damping member and a weight is particularly effective by matching the resonant frequency when directly connected to a load and its natural frequency with a drive device with low damping characteristics such as a direct drive motor. , the resonance magnification can be reduced, the control gain of the motor control system can be increased, and the drive device can operate at high speed and with high precision.

First, the configuration of the drive device of this embodiment will be explained using FIG. 1 and FIGS. 5 to 7. As shown in FIG. 1, an external cylinder rotor type direct drive electric motor is a rotor 5 that is supported on the stator 1 via a bearing by energizing a coil 2 that is installed on the stator 1. The operating principle is that a generated force based on Fleming's left-hand rule acts in the circumferential direction between the permanent magnet 3 and the rotor 5 rotated relative to the stator 1. The motor has rotor 5
A movement amount detector 4 is provided to detect the relative movement amount of the stator 1 to the stator 1, and the movement amount detection is detected by the movement amount command given to the motor control circuit and the movement amount detector 4 and fed back to the control circuit. A motor current command is generated based on the value deviation, amplified and applied to the coil. When a rigid load body is coupled to the rotor 5 of the electric motor, the load of the electric motor becomes a secondary load consisting of the motor rotor and the load body.
It vibrates based on the model shown in Figure 1, and has a low natural frequency when the load body has low rigidity or heavy load, and the movement amount detector 4 detects the vibrational movement amount, so the response frequency of the motor control system is (bandwidth) or more cannot be compensated for by the above-mentioned feedback control, and both the load body and the motor operate in an oscillatory manner. Taking a directly driven robot arm as an example, the above-mentioned problem occurs because the natural frequency is about 100 Hz, but the speed control system bandwidth is about 30 Hz. As a countermeasure against this problem, it is effective to provide a vibration absorbing device around the large-amplitude vibrating part of the motor and the movement detector. The former reduces vibrations in each direction of the electric motor, and the latter stabilizes the detected amount of movement by reducing vibrations in the rotational direction of the movement amount detector. This has the effect of preventing vibrations from being accelerated. When an electric motor rotates, in addition to rotational vibration, radial and axial vibrations occur due to bearing rigidity, load body rigidity, etc., which impede high-speed positioning of the load body. Therefore, it is effective to use a vibration absorbing device for reducing vibration in each direction shown in FIGS. 5 and 6. In FIG. 5, a high damping member (for example, rubber) 8 and a weight 9 are provided around the outer periphery of the motor rotor 5. The annular high damping member 8 is brought into close contact with the motor rotor 5 by bolting a two-part weight, and as shown in FIG. By tightening, the weight can be easily changed. The vibration absorbing device shown in Fig. 5 is effective in reducing vibrations in the circumferential direction, radial direction, and axial direction of the motor rotor by matching the natural frequencies in each direction with the resonant frequencies in each direction of the motor rotor and load body combination. There is. In addition, as shown in FIG. 6, a vibration absorbing device in which a protrusion is provided on the motor rotor and the annular high damping member 8 and the annular weight 9 are fastened with bolts is used in the circumferential direction and axial direction of the motor rotor. Effective in reducing vibration in the radial direction. However, since passive vibration absorbers with such a configuration usually have a natural frequency in each direction, it is difficult to suppress vibrations at resonance points that vibrate in multiple directions. Table 1 shows the experimental results of the natural vibration characteristics in the circumferential direction when a weight is attached as shown in Fig. 6 to butyl rubber (hardness 55) with an inner diameter of 90 mm and a thickness of 15 mm.

Table 1 The natural frequency changes as the weight of the weight changes, and the damping ratio is significantly improved considering that the damping ratio of ordinary mechanical structural materials is about 0.05.

In the direct drive device shown in Fig. 1, a vibration absorber (high damping member 8°, weight 9) configured as shown in Fig. 5 is installed on the top of the motor rotor, where the vibration amplitude is expected to be maximum when the load is connected, and a movement amount detector 4 is installed. A vibration absorbing device (high damping member 6, weight 7) having the configuration shown in FIG. 6 was provided at the bottom. Since the center of gravity of all vibration absorbing devices coincides with the motor rotation axis, the increase in motor load inertia is small.

Next, a structural example of an internal cylinder rotor type direct drive motor will be explained using FIG. 2. By energizing the coil 2 provided on the outer cylinder stator 1, the electric motor is connected to the permanent magnet 3 provided on the inner cylinder rotor 5, which is supported by the outer cylinder stator 1 via a bearing. A generated force based on Fleming's left-hand rule acts in the circumferential direction, causing the rotor 5 to rotate relative to the stator 1. The amount of movement of the rotor 5 is the large gear 10. The increased speed is transmitted to the movement amount detector 13 via the small gear 11 and the flexible coupling 12. The small gear 11 has a two-split spring connection structure in the axial direction, and is designed to minimize backlash when meshing with the large gear 10 due to the tension between the two split gears.

The reason why the small gear 11 and the movement amount detector 13 are connected by the flexible coupling 12 is to absorb the deviation of their axes. When using the movement amount detection method having such a structure, the movement amount detector 13 detects not only vibrations of the motor rotor 5 but also vibrations caused by the rigidity of the gear 10.degree. 11, the flexible coupling 12, etc. Therefore, a vibration absorbing device (high damping member 8, weight 9) is provided in the vicinity of the permanent magnet 3 in the motor rotor 5, and the small gear 11. A vibration absorbing device (high damping member 6, weight 74 is not clearly shown in the figure) is provided between the flexible couplings 12, and vibration reduction is achieved by matching each natural frequency with the motor resonance frequency.

Next, the structure of the speed reducer interposed rotary drive device of this embodiment will be explained using FIG. 3. The electric motor has the same configuration as the internal cylinder rotor type electric motor shown in FIG. The output shaft 15 is connected to a load body. With this structure, when the reduction ratio is n, the motor rotation speed is 1/n at the output shaft 15.
The torque generated by the electric motor is multiplied by n at the power shaft 15, and the electric motor can be made smaller and lighter. Since the torsional rigidity of the reducer is usually lower than the rigidity of the load body, vibrations caused by the torsional rigidity of the reducer adversely affect the motor control system. Therefore, in this device, a vibration absorbing device (high damping member 8, weight 9) is installed on the force shaft 15, and a vibration absorbing device (high damping member 6, weight 7) is installed around the movement amount detector 4 to achieve high damping. planned.

Next, the linear motion/rotary drive device of this embodiment will be explained using FIG. 4. This device has a structure that is often used as a wrist device for horizontal articulated robots.

The upper and lower shaft motors 19 transmit their rotational power to pulleys 25, belts 26, . A ball screw nut portion 29 that transmits information to the ball screw shaft 28 via the pulley 27 and is threadedly engaged with the ball screw shaft 28.
The bracket 30 coupled with the spline shaft 31 is restrained upwardly and downwardly.

They are connected freely in the rotational direction, and move vertically by the lead amount per rotation of the ball screw shaft 28. The spline shaft 31 is spline-fitted to a spline bearing 32, and the spline bearing 32 is connected to the spline bearing mounting member 3.
3. It is supported by the fixed part 34 via a bearing.

A spline shaft rotation pulley 24 is provided on the spline shaft mounting member 33 , and the speed reducer 17 . Pulley 22. It is rotationally driven via a belt 23. Therefore, the vertical and rotational movements of the spline shafts are driven by the respective shaft drive motors without interference. A travel amount detector 21.18 is directly connected to each shaft motor, and a mechanical brake 20 is provided on the output side of the vertical shaft motor 19 to prevent the vertical shaft from slipping due to gravity during a power outage. In such a drive device, a vibration absorbing bag W (high damping member 35, weight 36) is provided in the ball screw nut part in the vertical direction, and a vibration absorbing device (high damping member 37, weight 38) is provided in the rotating shaft pulley 24. By providing this, vibrations caused by belts, speed reducers, ball screws, etc. can be reduced.

By driving the driving device provided in the large-amplitude vibration part and the movement amount detection part of the above-mentioned passive vibration absorbing device so that its natural frequency matches the resonant frequency of the load body combination,
The resonance magnification can be reduced, the control gain can be improved, and the load body can operate at high speed and with high precision.

Next, a second embodiment of the present invention will be described with reference to FIGS. 39, 40, 8 to 17, and 36 to 38. This embodiment uses an active vibration absorbing device using a piezoelectric body to damp the vibration of the drive device. The purpose of this vibration absorbing device is to easily compensate for the resonance of the drive device/load body combination, improve the control gain of the motor control circuit beyond the configuration described in the first embodiment, and achieve higher speeds and higher speeds. The aim is to achieve precision operation.

Fig. 39 shows a vibration absorbing device and rotary drive device that reduce circumferential vibration, Fig. 40 shows the operating principle of Fig. 39, and Fig. 8 and 9 show a type of vibration suppressing device in the axial and radial directions. An example of the structure of a piezoelectric body is shown, and FIGS. 10 and 11 show a vibration absorber and a rotary drive device that reduce vibrations in the radial and axial directions. FIG. 12 shows a vibration absorber control circuit of this embodiment. Fig. 13 shows the current pattern during operation of the vibration absorber shown in Fig. 12, and Fig. 14 shows the frequency characteristics of the vibration acceleration power spectrum of the drive device when the vibration absorber of this embodiment is not used. 15 shows the frequency characteristics of the vibration acceleration power spectrum of the drive device when the vibration absorbing device of this embodiment is used, FIG. 16 shows an example of the transfer characteristic of the notch filter, and FIG. 36 to 38 show the structure of the vibration absorbing device of this embodiment. first,
An example in which piezoelectric rubber is used for the high damping member 8 portion of FIGS. 5 and 6 of the first embodiment is shown using FIGS. 8 and 9. Regarding piezoelectric rubber, see "Functional Materials J, April 1983 issue pp.
As described in 29-34, piezoelectric ceramic powder and raw rubber are mixed, heated, cross-wired, and vulcanized to make it rubber-like, and then electrodes are attached and piezoelectricity is imparted by applying high voltage and performing polarization treatment. be done. The piezoelectric constant of piezoelectric rubber is about 1/3 of that of piezoelectric ceramic sintered body, but the shape can be changed,
It has the advantage that various shapes can be manufactured. Here, the operation will be described using as an example the case where the polarization direction and the electrode facing direction are made to match, which is the most standard method of using piezoelectric materials. In this case, when a force acts when the electrodes are opened and the distance between the electrodes changes, a potential difference is generated between the electrodes due to the piezoelectric effect, so the piezoelectric body can be used as a vibration sensor. Conversely, when a voltage is applied between the electrodes, an inverse piezoelectric effect occurs. This changes the distance between the electrodes and generates force when restrained, so it can be used as a vibration damping actuator. By providing the annular segmented members shown in FIGS. 8 and 9 at the locations of the high damping members 8 shown in FIGS. 6 and 5, the effects described below can be obtained. Piezoelectric materials usually exhibit large piezoelectric effects in the following two directions.

(1) Longitudinal strain where the electrode facing direction and piezoelectric material polarization direction match: piezoelectric constant dss (2) Shear direction strain where the electrode facing direction and piezoelectric material polarization direction are orthogonal: piezoelectric constant dxb Here, the polarization direction is determined using the configuration shown in Figure 8. If the direction of polarization is set to the axial direction, axial vibration can be detected, and if the direction of polarization is set to the circumferential direction or the radial direction, vibrations in the circumferential direction or radial direction can be detected, respectively. Further, in the configuration shown in FIG. 9, if the polarization direction is set to the radial direction, radial vibration can be detected, and if the polarization direction is set to the circumferential direction and the axial direction, the vibration will be detected in the circumferential direction.

Axial vibration can be detected. Further, although a single layer structure is shown in FIGS. 8 and 9, the sensitivity as a sensor or actuator is increased by thinning the plate in the direction of the electrodes and creating a laminated structure as shown in FIG. Alternatively, the shapes shown in FIGS. 8 and 9 may be constructed using a piezoelectric ceramic sintered body instead of piezoelectric rubber. A vibration absorbing device configured as shown in FIGS. 6 and 5 using such a piezoelectric material is piezoelectric as described above when its natural frequency matches the resonant frequency of the drive device rotor 5. The body is a capacitor C
It can be expressed isometrically as a parallel combination element of m and resistance R.

FIG. 12 shows its configuration and drive circuit. In this case, when the electrodes are opened, the current flows in the direction of the solid line in the figure, and the capacitor C-
is charged with electric charge, and a resistance R' (resistance value is R
'>R), the current flows in the direction of the broken line in the figure, and the charging current is discharged. Therefore, by setting this charging/discharging to about twice the time constant of the CR circuit, 95%
charging and discharging can be achieved. For example, Hitachi Metal Film Pressure Element [laminated prismatic shape.

RZT ceramic sintered body) Table 2 shows an example of the characteristics.

From this, time constant Ti = CaR = 0, 26 μs
This is significantly shorter than the vibration period of 0.018 (for example, 100 Hz) at the resonant frequency of the drive device rotor. When this is done, Tn/Ts for the piezoelectric body 1 in Fig. 13 becomes 6410, and a damping current with a sufficiently fine resolution acts to cancel the vibration current, actively suppressing the vibration of the drive device rotor in each direction. Can be removed. This charging/discharging cycle is managed by the switch 4 by the timer 42.
4 on and off to drive the vibration absorbing element 45.

This makes it possible to suppress vibrations without supplying energy from the drive circuit.

Furthermore, in order to eliminate damping unevenness due to damping force that occurs during charging and discharging cycles, if multiple vibration absorbers are provided for the same vibration mode as shown in Figure 12, multiple phase (
By charging and discharging each vibration-absorbing element at the timing shown in the figure (a three-phase example), the sum of the vibration damping currents of each phase shown in Figure 13 can be made to almost match the negative phase current of the vibration current. This makes it possible to suppress vibrations in a shorter time. Figure 14 shows the drive device in any direction (axial direction,
The frequency characteristics of the vibration acceleration power spectrum in either the radial direction or the circumferential direction are shown.

The value of the resonant frequency (actually there is more than one) is large. FIG. 15 shows the frequency characteristics of the power spectrum of the vibration acceleration of the drive device when n-phase phase shift driving is performed using the vibration absorption device of this embodiment. From this, the vibration acceleration power spectrum at the resonant frequency Cf, ) of the drive device is the first
This is significantly reduced compared to the case shown in Figure 4. but,
Charge/discharge frequency Cf, ) and its n during multi-phase (n-phase) driving
Oscillations of twice the charging/discharging frequency (f s) are observed. However, if the charge/discharge cycle is 6Tt using the piezoelectric material shown in Table 1, fs = 641KHz, which poses no practical problem.However, if the distance between the electrodes becomes thick, fs becomes low, so fn of 5 If the frequency is more than double that, it will have an adverse effect on the drive control system. Therefore, by inserting a notch filter (see formula 0) into the drive device driving force command generating section, which sets the center frequency to the fs and n13 frequencies and reduces the gain, as shown in Figure 16, this vibration can be reduced as shown in Figure 17. As shown, the drive control system is stabilized.

In the above description of this embodiment, the structure of the vibration absorbing device using a piezoelectric body shown in FIGS. 8 and 9 has been described, but examples of other vibration absorbing devices are shown in FIGS. This will be explained using figures. In FIG. 10, a plurality of laminated expansion piezoelectric bodies as shown in FIG. 27 are provided in the radial direction from the drive device rotor 5a,
It is coupled with the weight 9, and by setting the polarization directions of the piezoelectric body in the radial direction, circumferential direction, and axial direction, vibration damping in each direction is possible.

In addition, the piezoelectric body has a hole in the center and a bolt-fastened structure, which allows it to have a preloaded structure.

The piezoelectric body has a structure with a small allowable load and no tensile force acting on it, which has the advantage of extending its life.

FIG. 11 shows a plurality of second
A laminated expansion piezoelectric body as shown in Fig. 7 is provided and combined with a weight 9. By setting the polarization directions of the piezoelectric body in the axial direction, the radial direction, and the circumferential direction, vibration damping force in each direction can be made 1. .

FIG. 30 shows a device in which a plurality of laminated electric bodies are provided in the circumferential direction from the drive device rotor 5b and combined with a weight 9, and the polarization directions of the piezoelectric bodies are set in the circumferential direction, axial direction, and radial direction. It becomes possible to suppress vibrations in each direction.

The vibration absorbing device described above in this embodiment has the disadvantage that piezoelectric bodies with different polarization directions must be used to damp vibration in each direction, and usually resonance occurs in multiple directions simultaneously at one resonance point. In this case, there is a problem in that a vibration absorbing device configured for each direction must be used. An embodiment that solves this problem will be described with reference to FIGS. 39, 40, and 36 to 38. The vibration absorbing device shown in Fig. 39 connects the drive device rotor 5 and the weight 9 by arranging two piezoelectric bodies polarized in the thickness direction diagonally, so that the resonance point can be set in the axial direction and the achieved vibration in the circumferential direction. Therefore, it is effective for damping the mode in which the drive device rotor 5 vibrates in both directions. In this case, the operation is as shown in FIG.
Regarding deformation in the circumferential direction, the piezoelectric body 40a extends in the polarization direction and therefore operates in the direction of contraction, and the piezoelectric body 40b contracts in the polarization direction and therefore operates in the direction of extension. Also, for deformation in the axial direction (not shown), the piezoelectric 40a is used.

Since both 40b deform in the same direction and generate acting force in the same direction, the pair of diagonally arranged piezoelectric bodies has a vibration damping function in two directions. Moreover, by adopting the configuration of the vibration absorbing device shown in FIG. 36 and FIG. 37, it is possible to suppress vibration in the radial direction, the @ direction, and the radial and two circumferential directions, respectively. In addition, as shown in Fig. 38, four piezoelectric bodies (40a to 40
If arranged as in b), there is an advantage that vibrations in three directions can be suppressed simultaneously. The driving direction is the same as the above method.

As described above, in the vibration absorbing device described in this embodiment, the piezoelectric body functions as a vibration sensor and a vibration damping actuator by commanding only the charging/discharging cycle of the piezoelectric body from the drive circuit, and has a simple structure that allows it to function in one direction or two directions. Or vibrations in three directions can be suppressed simultaneously.

Therefore, when installed in a drive device, it is more effective to selectively install the vibration absorber described in this example depending on whether the vibration mode at the prominent resonance frequency when coupled with the load body is one-way, two-way, or three-way. vibration control becomes possible. In addition, by installing it in the drive device shown in FIGS. 1, 2, 39, and 40, more effective vibration damping can be achieved compared to the passive vibration absorbing device described in the first embodiment. becomes possible.

In addition, in this explanation, an example of a vibration absorbing device using a rotary type drive device was mainly given, but it is also possible to apply it to a rectangular type drive device by developing the vibration absorbing device so that the circumferential direction in the example is a linear direction. It is.

As mentioned above, regarding the example of the multi-axis mechanism using the vibration absorbing device and the drive device equipped with the vibration absorbing device described in the first and second embodiments,
This will be explained using FIGS. 18 and 19. FIG. 18 shows the external appearance of the horizontal articulated direct drive robot, and FIG. 19 shows a longitudinal sectional view of the robot shown in FIG. 18. Horizontal articulated robots are widely used in industrial applications as assembly robots, and are required to operate at high speed and with high precision. -9 The system that uses a direct drive motor for two-axis drive has higher torsional rigidity compared to the system that uses a drive device with a reduction gear, so the natural frequency is increased, realizing faster and more accurate operation. There is a possibility that it can be done. On the other hand, vibrations with a high natural frequency and high resonance magnification caused by the low damping of the drive motor directly hinder the improvement of the control gain of the motor control system, making it impossible to directly achieve the high servo torsional rigidity that can be set in the drive motor. Therefore, it is effective to apply the vibration damping device and drive device of the present invention. An example is described below.

FIG. 18 shows a perspective view of a horizontal articulated direct drive robot. -9 As the two-shaft direct drive motor 49,52, an external cylinder rotor type motor shown in FIG. 1 is used. A two-axis direct drive motor 52 provided at the tip of the second arm 51 rotates the second arm 54 . Second arm 54
There is a spline shaft 31 having a drive mechanism shown in FIG. 4 and having a tool 39 attached to its tip, which moves in the vertical and rotational directions. Each drive device is provided with a vibration absorber at the position described in FIGS. 1 and 4.

The arrangement of the vibration absorbing device is fully expressed in the vertical sectional view of FIG. 18 shown in FIG. 19.

As shown in Figure 19, for safe operation, the robot is equipped with a specific position detector for overrun and home position inspection, as well as a detected object, and a braking command is issued after overrun detection. A stopper is provided that forcibly stops the operation when the operation goes too far.

As described above, by applying the vibration absorbing device and drive device described in the first and second embodiments of the present invention to the present robot, sufficient vibration suppression and high-speed, high-precision operation are possible.

〔Effect of the invention〕

According to the present invention, by providing a vibration absorbing device in the large-amplitude vibrating part and the movement amount detection part when the load body is coupled to the drive device, it is possible to reduce the vibration of the load body and the vibration of the movement amount detection value. It is possible to improve the control gain of the device control system, enabling high-speed, high-precision operation of the drive device.

[Brief explanation of drawings]

1 to 4 are longitudinal cross-sectional views of a driving device according to an embodiment of the present invention, FIG. 5 and FIG. 6 are perspective views and cross-sectional views of a vibration absorbing device according to an embodiment of the present invention, and FIG. is an assembled view of the variable weight structure of the weight of the vibration absorbing device shown in FIG. 5, FIGS. 8 to 11 are perspective views of the vibration absorbing device of the second embodiment of the present invention, and FIG. A drive block diagram of the vibration absorbing device of the second embodiment, FIG. 13 is a current characteristic diagram when driving the drive circuit of FIG. 12, and FIG. 14 is a vibration acceleration power of the drive device when the vibration absorbing device of the present invention is not installed. A frequency characteristic diagram of the spectrum, FIG. 15 is a frequency characteristic diagram of the vibration acceleration power spectrum of the drive device when the vibration absorber according to the second embodiment of the present invention is installed, and FIG. 16 is a frequency characteristic diagram of the notch filter transfer function. Fig. 17 is a frequency characteristic diagram of the vibration acceleration power spectrum of the drive device when the vibration absorbing device of the second embodiment of the present invention is used in combination with the notch filter of Fig. 16, and Fig. 18 is a frequency characteristic diagram of the vibration absorbing device of the present invention. A perspective view of a horizontal articulated direct drive robot using an attached drive device, FIG. 19 is a vertical sectional view of the robot in FIG. 18, FIG. 20 is a perspective view of a drive device that drives an arm-shaped load, and FIG. 21 , Fig. 22 is a mechanical explanatory diagram of Fig. 20 without and with a passive vibration absorber installed, respectively, and Fig. 23 is a frequency characteristic of the torque/velocity transfer function of the dynamic model shown in Figs. 21 and 22. 24 is a block diagram of the drive circuit shown in FIG. 22, and FIGS. 25 and 26 are a block diagram of the drive circuit shown in FIG. 22, respectively. A frequency characteristic diagram of the transfer function, Fig. 27 is a perspective view of the multilayer piezoelectric material, Fig. 28 is an equivalent mechanical model diagram and an equivalent circuit diagram of Fig. 27, and Fig. 29 is an impedance of the multilayer piezoelectric material of Fig. 27. 30 is a perspective view of the vibration absorbing device of the second embodiment of the present invention, FIG. 31 is a mechanical explanatory diagram of FIG. 30, and FIG. 32 is an equivalent view from the piezoelectric body of FIG. 30. Circuit diagram, Figure 33 is the third
0 is a drive circuit diagram of the vibration absorbing device shown in FIG. 34 is a current characteristic diagram when driving the drive circuit of FIG. 32, FIG. 35 is a frequency characteristic diagram of the torque/speed transfer function when the drive absorption device of FIG. 30 is installed in the drive device of FIG. 20, FIG. 36 is a sectional view of the vibration absorbing device of the present invention, FIG. 37 is a plan view of FIG. 36, and FIG. 38 is a perspective view of the piezoelectric body of the present invention.
FIG. 39 is a perspective view of a vibration absorbing device according to a second embodiment of the present invention, and FIG. 40 is a diagram showing the operating principle of the vibration absorbing device of FIG. 39. DESCRIPTION OF SYMBOLS 1... Drive device stator, 2... Coil, 3... Permanent magnet, 4... Travel amount detector, 5... Drive device rotor, 6,8°35.37... High damping member, 7.9,3
6,38゜70...Weight, 10.11...Gear,
12...Flexible coupling, 13,18,2
1,64°65...Movement amount detector, 14. ! 7...
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Claims (1)

[Claims] 1. An electric motor that generates rotation or linear relative movement between a stationary body and a moving body by supplying energy, a movement amount detection section that detects the amount of the relative movement, and a movement amount detection section that detects the amount of the relative movement, and transmits the generated power from the electric motor to a load body. A drive device comprising a transmission mechanism including a deceleration mechanism for transmitting transmission, characterized in that a vibration absorption device is provided in the movement amount detection section of the drive device and/or the large amplitude vibration section of the drive device/load body combination. drive device.