JP3360935B2 - Machine resonance detection device and vibration suppression control device in motor control system - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an electric motor control device provided with a mechanical vibration detecting device in an electric motor control system for detecting a mechanical resonance signal of a mechanism driven by an electric motor.

[0002]

2. Description of the Related Art In a mechanism control system in which a force transmission mechanism such as a ball screw is driven by an electric motor, the mechanical resonance phenomenon of the transmission mechanism is
There is a problem of determining the upper limit of the loop gain of the control system and deteriorating the control performance. For example, in a direct-acting mechanism including a ball screw having an electric motor attached and a movable base, in order to avoid mechanical resonance such as torsional vibration of the ball screw, the loop gain of the control device of the electric motor is set to a standard setting value (setting of the electric motor alone).
However, there is a problem that the motion characteristics of the movable table deteriorate. It is known that when the speed of the movable table can be directly detected, the mechanical resonance signal can be detected from the speed and the angular velocity of the electric motor, and this signal can attenuate mechanical vibration and increase the loop gain of the control system. (“Design of a motor control system for driving a low-rigidity load machine using a reference model”, 1987 Proceedings of National Conference of Industrial Applications Division of the Institute of Electrical Engineers of Japan, p452). In many cases, since the speed detector cannot be attached to the movable table or the like, the speed cannot be detected, the mechanical resonance signal cannot be detected, and the mechanical vibration cannot be controlled.

Therefore, as a first conventional technique, a motion equation of a mechanism and an electric motor is established in consideration of a resonance phenomenon, a state observer is constructed with respect to a state equation derived from this motion equation, and the observer estimates. A method of outputting the mechanical vibration signal from the state quantity is known (for example,
"One-speed control method of separately excited DC machine having two-inertia resonance system", IEEJ Transactions B, pp111-118 (Showa 61)).
Further, as a second conventional technique, there is a method of detecting the machine resonance signal by a simulator device including a current control system simulator (first-order lag filter) and an inertia simulator (integration) (“speed control with simulator follow-up control”). , 1985 Proceedings of the Annual Meeting of the Institute of Electrical Engineers of Japan, 60
4, p. 714). Specifically, as shown in FIG. 16, the mechanical resonance signal component is included in the difference signal obtained by subtracting the angular velocity signal of the electric motor from the output signal of the simulator 801 to which the current command signal of the electric motor control device is input. Also,
As a third conventional technique, there is a method of attenuating mechanical vibration by using a minimum-dimensional disturbance torque observer (“Vibration suppression control of multi-inertia system based on disturbance torque observer”, Proc. , 666, p6-7
3). The disturbance torque observer estimates the torque disturbance of the electric motor and feeds it back to the torque control device (generates the driving torque through the current regulator) (gain K).
1) to suppress mechanical vibration.

[0004]

However, in the first prior art, in order to design the state observer, a model of the electric motor and the mechanical resonance element (secondary system) is essentially required as a controlled object. . In order to estimate the mechanical vibration, it is necessary to calculate the model at a high speed (about 30 times the resonance frequency), but when the vibration frequency is high, there is a problem that the calculation cannot catch up. For example, at a resonance of about 1000 Hz that occurs in an encoder that observes the angle of a motor, the calculation cycle is about 30 μs (30 kHz), which is difficult to realize. Since the model of one mechanical resonance element (secondary system) can describe only one frequency mechanical resonance phenomenon,
When a resonance phenomenon of a frequency or higher is involved, the order of the state observer becomes very large and the amount of calculation increases. In such a case, realization becomes more difficult. In an actual machine, a plurality of resonance phenomena often occur, such as a torsional resonance phenomenon of a ball screw or a motor shaft and a resonance phenomenon of the encoder, or a resonance phenomenon of the torsional resonance and a coupling of a transmission system. In addition, for the operation of the observer, it is necessary to set the parameters of the mechanical resonance system in addition to the electric motor.
The physical parameters of the mechanical elements change significantly if the type of mechanism changes, and it is also difficult to measure all the physical parameters for each mechanism. Therefore, it is practically impossible to set all the parameters required for the observer in the field. In particular, as the order of the observer increases, the number of parameters to be set by the observer increases, and further, it becomes extremely difficult to adjust the observer.
At the production site, gain adjustment and parameter setting should be as simple as possible, and the problem of the prior art is serious in this respect.

In addition, the second conventional technique has a problem that the parameter variation directly affects the difference signal.
For example, when the moment of inertia of the mechanism fluctuates, the integral gain of the inertia simulator is inversely proportional to the moment of inertia. Therefore, in the difference signal output, removal of the low-frequency velocity signal component becomes insufficient, and only the mechanical resonance signal is extracted. Cannot be performed (the quality of the mechanical resonance signal deteriorates). Further, since the simulator is a model in an ideal state in which no disturbance or the like exists, disturbance signal components such as the frictional force of the transmission mechanism and the reaction force from the load are directly included in the difference signal output. Therefore, there is a problem that only the mechanical resonance signal cannot be extracted and the quality of the mechanical resonance signal deteriorates. In the third conventional technique, the estimated signal of the torque disturbance of the electric motor essentially includes the mechanical resonance signal and the disturbance signal due to the frictional force of the transmitted climate. Originally, the disturbance signal is mainly composed of low-frequency components, and the mechanical resonance signal is mainly composed of high-frequency components, so that the optimum compensator designs are usually different. In the third conventional technique, the two cannot be essentially separated, so that there is a big problem that it is difficult to design an optimal compensator for disturbance suppression and mechanical resonance suppression. This means that feedback gain adjustment is not easy,
There also arises a problem that it is not easy to simultaneously pursue the widening of the speed control system, which is the main control system, and the suppression of mechanical vibration.

The problem to be solved by the present invention is to estimate mechanical vibration at high speed, simplify gain adjustment and parameter setting on site, and obtain mechanical resonance signal with high accuracy. It is an object of the present invention to provide a vibration suppression control device capable of simultaneously achieving a wide band of a detection device and a main control system and vibration suppression of mechanical vibration.

[0007]

In order to solve the above-mentioned problems, the present invention provides a torque of an electric motor when a torque command is given.
Vibration detection device applied to a motor control device for controlling motors
And γ is the reciprocal of the moment of inertia of the motor and load
And Do is the viscous friction coefficient of the electric motor and the load.
When divided by the moment of inertia of the motor and the load
, The input signal is multiplied by the above γ and the output of the integrator
When the difference between the signals multiplied by Do is input to the integrator,
An equivalent rigid body model whose output is the output of the integrator
And a first compensating means for proportional amplification, and an integral operation
A second compensating means and an output which is an output of the second compensating means.
A first adding the constant disturbance signal and the output of the first compensating means
Adder, the torque command and the output of the first adder
Second addition to add as the input of the equivalent rigid body model
And the speed of the electric motor obtained by the speed detection means
The output of the equivalent rigid body model from the velocity signal corresponding to
And subtract the resulting mechanical resonance signal with the first compensating means.
And a subtracter for inputting to the second compensating means .

Further, according to the present invention , when a torque command is given.
Machines applied to electric motor control devices that control the torque of electric motors
A mechanical vibration detector, in which γ is the inertial motor of the motor and the load.
And the viscosity of the electric motor and the load.
Divide the friction coefficient by the moment of inertia of the motor and the load.
The input signal multiplied by the above γ
The difference of the signals obtained by multiplying the output of the divider by Do is
It is used as an input and the output of the integrator is used as its output.
An equivalent rigid body model and a first compensation means for performing proportional amplification,
Second compensating means for performing integral calculation, and integral calculation or first-order delay
Third compensating means for performing this calculation, and integral calculation or primary
Fourth compensating means for performing delay calculation, and the second compensating means
The estimated disturbance signal which is the output of the first and the output of the first compensating means
A first adder for adding the torque command and the first adder
The output of the adder of is added to the input of the equivalent rigid body model
Before the second adder to perform the speed detection means
The equivalent rigid body model is calculated from the speed signal corresponding to the speed of the motor.
Dell's output is subtracted and the resulting mechanical resonance signal is
Subtractor for inputting to the first compensation means and the second compensation means
And inputting the mechanical resonance signal to the third compensating means,
The output of the third compensating means is input to the fourth compensating means.
To the output of the fourth compensating means and the motor controller.
A third adder that receives the torque command signal of
The output signal of the third adder is input to the torque command input terminal of the electric motor control device.

In this vibration damping control device, when the electric motor is a control system for controlling the angular velocity, the output of the third compensating means and the speed command signal to the electric motor control device are set to the first value.
3 and the output signal of the third adder is input to the speed input terminal of the motor control device. In the present invention, the equivalent rigid body model means a rigid system in which the input / output relationship is the rest of the model of the mechanical system including the electric motor including the angular velocity detector, from which mechanical vibration elements causing mechanical vibration are removed.

[0010]

First, the principle of the present invention will be described. In FIG. 1, a block 101 surrounded by a broken line represents a mechanism (including an electric motor) accompanied by a mechanical resonance phenomenon, and a block 102 surrounded by a one-dot chain line represents a mechanical vibration detecting device according to the present invention. As an example of the mechanism, consider a two-inertia torsional resonance system shown in FIG. The equation of motion of the mechanical system in the figure is

[Equation 1] Becomes Where θ _m : Motor shaft angle [rad] θ _L : Load shaft angle [rad] δ θ: Equivalent linear spring torsion angle [rad] J _m : Motor shaft side inertia moment [Kgm ² ] J _L : Load Moment of inertia on shaft side (converted to motor shaft) [Kgm ² ] K: Spring constant of equivalent linear spring [Nm / rad] D _m : Viscous friction coefficient of motor shaft [Nm / (rad / s)] D _L : Load shaft Viscous friction coefficient [Nm / (rad / s)] _DT : Equivalent viscous friction coefficient of equivalent linear spring system [Nm / (rad / s)] u ₁ : Torque generated by motor shaft [Nm]. Where the state variable is

[Equation 2] Observation output (angular velocity of electric motor)

[Equation 3] To get However,

[Equation 4] Is.

The transfer function from the torque generated by the motor to the angular velocity (observation output) of the motor is

[Equation 5] Is. However, the characteristic equation P (s) is

[Equation 6] Is. The expression (6) is a combination of the parameters of the electric motor and the parameters of the mechanical resonance system, and it was very difficult to factor the expression into the product of the linear expression and the quadratic expression. For this reason, the physical meaning of the equation is unclear, and the state observer must be configured for the entire control system.

Therefore, in the present invention, in the equation (6),

[Equation 7] far. Here, α is the ratio of the load inertia moment to the motor α = J _L / J _m (8). When Expression (6) is expressed using (7a) to (7c), P (s) = s ³ ((ω ₀ / Q ₀ ) + D ₀ ) s ² + ω ₀ ² s + ω ₀ ² D ₀ (9) Become. Where new

[Equation 8] think of. However, in the equation (10b), ε = (1 / Q ₀ ) (D ₀ / ω ₀ ) (11) is set. From equations (9) and (10b), when ε = 0,

[Equation 9] Therefore,

[Equation 10] Is an approximation formula of P (s), and ε represents an approximation error.
As can be seen from the equation (11), the larger the Q ₀ , the smaller the approximation error ε. For example, Q ₀ > 10, ω ₀ > 10D
_{When 0} , ε <0.01.

By assuming the equations (7a) to (7c), the characteristic equation P (s) is expressed by the equations (10a) and (10b).

[Equation 11] so,

[Equation 12] Can be approximated by Here, in equation (5)

[Equation 13] (12) and (13a) and (13b) are used in equation (5), the transfer function approximation equation

[Equation 14] Can be derived. By using the equations (14a) and (14b), the two-inertia resonance system has an equivalent rigid system G ₁ (s) shown in the block 101 of the broken line in FIG.

[Equation 15] And mechanical resonance system G ₂ (s)

[Equation 16] Can be separated into If the velocity of the equivalent rigid system can be detected, only the mechanical vibration component can be detected by the difference signal obtained by subtracting the detected velocity of the equivalent rigid system from the observation output.

The velocity of the equivalent rigid system is a virtual state quantity and cannot be detected . Using the state observer , the velocity of the equivalent rigid system can be observed , but simply applying the known design theory of the state observer would force the observer to include the mechanical resonance system G ₂ (s). The object of the present invention cannot be achieved. Therefore, in the present invention, attention is paid to the fact that the mechanical system can be separated into an equivalent rigid system and a resonant system, and the equivalent rigid system physically has a low-pass characteristic and hardly responds to a resonant component of high frequency. Considering that, the angular velocity of the electric motor (the output of the resonance system in FIG. 1) is approximated to the observed output, and an equivalent rigid body observer (a state observer modeled on the equivalent rigid system) is constructed.

That is, in FIG. 1, the velocity of the equivalent rigid system

[Equation 17] Differential equation with respect to

[Equation 18] Holds. Equation (16b) is a disturbance model.
In equations (16a) to (16c), the state variables are

[Formula 19] , The equation of state and output equation of the equivalent rigid system

[Equation 20] To get However,

[Equation 21] Is.

Here, c _{1 in the} equations (19a) and (19c)
And determinant consisting of A ₁

[Equation 22] Is regular, the control system of equations (18a) and (18b) is observable, and the observer is concerned with equations (18a) and (18b).

[Equation 23] Is configurable. In equation (21),

[Equation 24] Is the observer gain, and (19a), (19c)
From formula and formula (22a)

[Equation 25] Is.

From equations (21) and (22b),

[Equation 26] To get Here, letting k ₁ = ω _f / Q _f (24a) k ₂ = ω _f ² / γ (24b), equations (23a) and (23b) are

[Equation 27] Becomes That is, u is a variable proportional to the torque of the motor.
Enter in (t), enter the angular velocity of the motor in y (t),
When formulas (25a) and (25b) are calculated,

[Equation 28] Becomes a mechanical resonance signal,

[Equation 29] Is the estimated disturbance signal.

Since the equations (25a) and (25b) can be expressed by the block diagram in the block 102 indicated by the alternate long and short dash line in FIG. 1, the mechanical vibration detecting device 102 is constructed by the block diagram. That is, according to the present invention, as shown in FIG. 1, the estimated equivalent rigid body velocity is essentially

[Equation 30] Equivalent rigid body model means for outputting and first proportional calculation
And a second compensating means for performing integral calculation. In addition, the formula (7c) and the formula (13a) are added to the formula (16c).
Using the formula,

[Equation 31] To get In the equation (26), usually D _m D _T << J _m K, D _L
Since D _T << J _m K and D _m D _T << J _m K, the formula (26) is obtained by using the formula (8).

[Equation 32] Can be approximated by Also, in the formula (7a), D _m D _T << (1 + α)
J _m K, D _L D _T << (1 + α) J _m K, D _m D _T << (1 + α)
Since it is J _m K, D ₀ is

[Expression 33] Can be approximated by 1 or (25), a mechanical vibration detecting apparatus of the present invention, the state observer band omega _f
And Q _f , and γ as a parameter of the mechanism including the electric motor
And the parameters D ₀ must be set. ω _f and Q _f are parameters that can be freely set as needed. From equations (27) and (28), parameters γ and D ₀
Can be easily calculated or measured from the moment of inertia of the motor and load (transmission mechanism) and the viscous friction coefficient. Alternatively, it is easy to calculate the gamma by measuring the resonance frequency and the anti-resonance frequency of the mechanical resonance system. That is, in the mechanical vibration detecting device of the present invention, the parameters can be easily adjusted.

[0019]

EXAMPLES Specific examples of the present invention will be described below. A first embodiment of the present invention will be described with reference to FIGS. 3 and 4. Figure 3
Is an equivalent conversion of the block diagram of the mechanical vibration detecting device (indicated by the one-dot chain line) in FIG. 1. In the figure, an integral element 301 and gains 302 and 303 are equivalent rigid body models, 304 is an inverting amplifier, 305 is a first compensating means (proportional element), 3
Reference numeral 06 is a second compensating means (integrating element). Gain 30
2 means the sum of the moments of inertia of the electric motor and the load machine. The gain 303 is a viscous friction between the electric motor and the load, which can usually be ignored.

FIG. 4 is a block diagram of FIG. 3 realized by an operational amplifier. In the figure, the operational amplification circuits 401 to 404 constitute a mechanical vibration detection device. When the integral calculation circuit 401 is an equivalent rigid body model and D ₀ is ignored when the viscous friction in FIG. 3 is small, the feedback resistor R ₃ is removed. A signal u proportional to the torque of the electric motor
(t) (for example, a torque monitor signal or a torque command signal of a torque control device for an electric motor)
When the angular velocity of the electric motor is input to the operational amplifier circuit 402 as an observation output signal, the operational amplifier circuit 402 can output a mechanical resonance signal and the integral operation circuit 404 can output an estimated disturbance signal. The present embodiment has an advantage that the mechanical vibration signal and the estimated disturbance signal can be output independently.

When the estimated disturbance signal is not required, the proportional operation circuit 403 and the integral operation circuit 404 shown in FIG. 4 can be integrated. This is shown in FIG. 5 as a second embodiment. 5 is the same as the embodiment of FIG. 4 except for the proportional-plus-integral arithmetic circuit 503, and the operation as the mechanical vibration detection device is equivalent to that of FIG. In this embodiment, there is an advantage that the mechanical vibration detection device can be configured with three operational amplifier circuits.

FIG. 6 shows the experimental results of estimation of the mechanical resonance signal obtained by applying the second embodiment (FIG. 5) to the linear motion mechanism consisting of the electric motor (AC servo motor), the ball screw and the table. FIG. 6A is a result of measuring the frequency characteristic from the torque command input of the power amplifier for driving the electric motor to the electric motor angular velocity (detection output) with a servo analyzer. -6d from low frequency to about 100Hz
It is B / OCT, and the characteristic of the equivalent rigid system (see 101 in FIG. 1) can be confirmed. Mechanical resonance characteristics (see 101 in FIG. 1) at an antiresonance frequency of 145 Hz and a resonance frequency of 395 Hz can be confirmed at 100 Hz to 1 KHz. FIG. 6B shows a machine resonance signal (machine resonance detection device 10 based on the torque command).
2 is the result of measuring the frequency characteristics up to (output 2) with a servo analyzer. Except for the resonance characteristics of 395 Hz, it is almost 2
Since it is reduced to 0 dB or less, the mechanical resonance detecting device 1
It can be confirmed that 02 can detect the mechanical resonance signal. Gain adjustment and the like are performed as follows. (1) The band of the mechanical vibration detector (ω _{f in} FIG. 3) is reduced. In FIG. 4 or 5, for example, R4 is reduced. In FIG. 14, VR2 is made small (or zero). (2) The electric motor is driven by a torque control device or a speed control device of the electric motor. However, the control device is adjusted to a gain that does not cause mechanical vibration. (3) Observe mechanical vibration (resonance) signal output. In FIG. 4 or FIG. 5, it is the output of the operational amplifier 402. (4) The parameter (γ in FIG. 3) that means the moment of inertia of the electric motor and the load machine is adjusted so that the signal output becomes substantially zero. In FIG. 4 or 5, R2 may be changed. In FIG. 14, VR1 is changed. (5) The band (ω _{f in} FIG. 3) of the mechanical vibration detector is used as the design value. In FIG. 4 or 5, R4 is used, and in FIG. 14, VR2 is used.
To the design value.

A third embodiment of the mechanical resonance suppression control system of the present invention will be described with reference to FIGS. 7 and 8. FIG. 7 is a block diagram of the vibration damping control device of the invention applied to the torque control device. In the figure, a mechanical system 101 including an electric motor and a mechanical vibration detecting device 102 are the same as those in the first embodiment or the second embodiment, and therefore their explanations are omitted. The mechanical resonance signal estimated by the mechanical vibration detecting device 102 is input to the third compensating means 720 constituted by an integral element or a first-order lag element.
The output of the third compensating means 720 is input to the fourth compensating means composed of an integrating element or a first-order lag element, and the output is input to the summing amplifying means 710. The addition input unit 710 and the damping feedback gain (Kf) unit 712 are inverted / inverted.
It comprises a non-inverting amplification means 711 and an addition means (indicated by an addition point in the figure). The adding means adds the mechanical resonance signal that has passed through the third and fourth compensating means and the inverting / non-inverting amplifying means and the torque command signal output from a higher-order control device (for example, compensator for speed control). And outputs it as a torque command signal of the torque control device for the electric motor. The torque command signal of the torque control device for the electric motor or the torque monitor signal for the torque control device for the electric motor is input to the mechanical vibration detection device 102 as a signal u (t) proportional to the torque of the electric motor.

FIG. 8 shows an example in which a block diagram of the third and fourth compensating means and the adding and amplifying means is realized by an operational amplifier. In the figure, the third and fourth compensating means are composed of integrators. When the DC offset voltage of the operational amplifier or the like has a bad influence, a resistor may be inserted in parallel with the capacitor C1 or C2 to form an incomplete integrator, which is a first-order lag system. The inverting / non-inverting amplifying means 711 is an opening / closing means K.
If is closed, inverting amplification is performed, and if K is opened, non-inverting amplification is performed. The damping feedback gain is adjusted using VR1, VR2, and VR3. VR2 and VR3 are used as auxiliary so that the mechanical resonance signal has an appropriate magnitude.
VR1 is mainly used as the damping feedback. FIG. 9 shows the results of a vibration damping test of mechanical resonance obtained by applying the third embodiment (FIG. 7) to a linear motion mechanism composed of an electric motor (AC servomotor), a ball screw and a table. The experimental apparatus (not shown) is provided with a speed control compensator (not shown) that inputs a speed command signal and an electric motor angular velocity signal and outputs a torque command signal. The speed control compensator is a proportional control of the speed control gain K _v .

FIG. 9 shows the measurement results of the frequency characteristic from the speed command to the electric motor angular velocity signal, and FIG. 9A shows the case where the damping control is not performed (VR1 → 0/10, K _f →
0) and (b) are for vibration damping control (VR1 → 10 / in FIG. 8).
10, K _f → 100%). Gain adjustment and the like are performed as follows. (1) VR1 → 0/10 in FIG. 8 is set, and the speed control gain is increased while moving the table of the linear motion mechanism until the speed control system becomes oscillated (FIG. 9 (a)). (2) In FIG. 8, the monitor M1 simultaneously observes the vibration waveform included in the torque command, and the monitor M2 simultaneously observes the machine resonance signal output by the machine resonance detecting device. (3) VR2 (if necessary, V2 so that the amplitudes of the vibration waveforms of the monitor M1 and the monitor M2 are almost the same).
R3 also). (4) The band ω _f of the equivalent rigid body observer included in the mechanical resonance detection device and the opening / closing means K of the addition amplification means 710 is adjusted so that the phases of the respective vibration waveforms of the monitor M1 and the monitor M2 are inverted phases. For example, R4 in FIG. 5 may be changed to change ω _f . (5) When VR1 is raised, mechanical resonance disappears. (6) Increase the speed gain K _V until mechanical resonance occurs. (7) Raise VR1 to eliminate mechanical resonance (use VR2 and VR3 if necessary). (8) The above (6) and (7) are repeated to increase the velocity loop gain (FIG. 9 (b)). It can be confirmed that the frequency band of the speed control system can be expanded while suppressing mechanical resonance (395 Hz) by the above adjustment (FIG. 9).

Next, referring to FIGS. 10 and 11, a fourth embodiment of the mechanical resonance suppression control system of the present invention will be described. FIG. 10 is a block diagram of the vibration suppression control device of the invention applied to the speed control device. In the figure, the mechanical system 101 including the electric motor and the mechanical vibration detecting device 102 are the same as those in the first embodiment or the second embodiment, and therefore their explanations are omitted. A third compensating means 7 which configures the mechanical resonance signal estimated by the mechanical vibration detecting device 102 with an integral element or a first-order lag element.
Enter in 20. The output of the third compensation means 720 is input to the summing amplification means 710. The summing amplification means 710 is a damping feedback gain ( _Kf ) means 712, an inverting / non-inverting amplification means 711, and an addition means (indicated by addition points in the figure).
Consists of. The adding means adds the mechanical resonance signal that has passed through the third compensating means 720 and the inverting / non-inverting amplifying means 711 and the speed command signal output from a higher-level control device (for example, a position control compensator). , As a speed command signal of the speed control device for the electric motor. A signal u proportional to the torque command signal of the motor speed control device or the torque monitor signal of the motor speed control device is proportional to the torque of the motor.
It is input to the mechanical vibration detection device 102 as (t).

FIG. 11 shows an example in which the block diagram of the third compensation means 720 and the summing amplification means 710 is realized by an operational amplifier. In the figure, the third compensation means 720 is composed of an integrator. When the DC offset voltage of the operational amplifier or the like has a bad influence, a resistor may be inserted in parallel with the capacitor C1 to form an incomplete integrator to form a first-order lag system.
The inverting / non-inverting amplification means 711 performs inverting amplification when the opening / closing means K is closed, and non-inverting amplification when K is opened. V
The damping feedback gain is adjusted using R1 and VR2. VR2 is used as an auxiliary so that the mechanical resonance signal has an appropriate magnitude, and VR1 is mainly used as damping feedback.

A fourth embodiment (see FIG. 1) is used in a linear motion mechanism composed of an electric motor (AC servomotor), a ball screw and a table.
Fig. 12 shows the results of the vibration control experiment of mechanical resonance obtained by applying 0).
Shown in. In the experimental device (not shown), the speed control compensator of the speed control device was a proportional control of the speed control gain K _v . FIG. 12 shows the measurement result of the frequency characteristic from the speed command to the electric motor angular velocity signal. (A) shows the case where the vibration damping control is not performed (VR1 → 0/10, K _f → 0 in FIG. 11),
(B) shows the case of performing the vibration suppression control (VR1 → 10 /
10, K _f → 100%). (C) shows the case of performing damping control (VR1 → 10/10, K _f → 100 in FIG. 11).
%), _Which raises the velocity loop gain K _v .

The gain adjustment and the like are performed as follows. (1) Set VR1 → 0/10 in FIG. 11 and increase the speed control gain until the speed control system becomes oscillating while moving the table of the linear motion mechanism (FIG. 12 (a)). (2) In FIG. 11, the vibration waveform included in the detected speed of the electric motor is simultaneously observed by the speed monitor of the electric motor speed control device, and the mechanical resonance signal output by the mechanical resonance detection device is simultaneously observed by the monitor M2. (3) VR2 is adjusted so that the amplitudes of the vibration waveforms of the concentration monitor and monitor M2 are approximately the same. (4) The band ωf of the equivalent rigid body observer included in the mechanical resonance detecting device and the opening / closing means K of the addition amplifying means 710 is adjusted so that the phases of the vibration waveforms of the speed monitor and the monitor M2 are inverted phases. For example, R4 in FIG. 5 may be changed to change ωf. (5) Mechanical resonance disappears as VR1 is raised (Fig. 1
2 (b)). (6) Increase the speed gain K _V until mechanical resonance occurs. (7) Raise VR1 to eliminate mechanical resonance (use VR2 if necessary). (8) The above (6) and (7) are repeated to increase the velocity loop gain (FIG. 12 (c)). By the above adjustment, it can be confirmed that the frequency band of the speed control system can be expanded while suppressing the mechanical resonance ( 660 Hz ). From the experimental results of FIG. 12, it can be confirmed that the frequency band of the speed control system can be expanded from 25 Hz to 90 Hz.

FIG. 13 shows a fifth embodiment of the present invention. In the mechanical vibration suppressing device 601 of this embodiment, the first connector 602, the second connector 603, and the adding means 60 are included.
4. It has a phase compensation means 605. The first connector 602 is connected to the connector 607 of the electric motor control device 606, and transmits the torque monitor signal of the electric motor control device 606 to the mechanical vibration detection means 608 and the second connector 603. The mechanical vibration detection means 608 outputs the mechanical resonance signal to the phase compensation means 605. The mechanical vibration detecting means 6
Since 08 has been described in detail in FIG. 1 and its description, it is omitted here. The output signal of the phase compensating means 605 is appropriately gain-adjusted by the damping gain adjusting means 609,
It is input to the adding means 604 as a mechanical resonance braking signal.
The second connector 603 is connected to the command side connector 610, and outputs the torque monitor signal and the speed monitor signal from the command side connector 610. The speed command signal of the command side connector 610 is input to the adding means 604 through the second connector 603. The output signal of the adding means 604, which is the sum of the mechanical resonance braking signal and the speed command signal, passes through the first connector 602 and becomes the speed command of the electric motor control device 606. FIG. 14 shows an example of the mechanical vibration detecting means. In the figure, VR1 is a variable resistor for adjusting the moment of inertia of the electric motor and the load, and VR2 is a variable resistor for adjusting the phase of the vibration waveform. The principle of detecting the mechanical resonance signal is as described above.

FIG. 15 shows an embodiment of the phase compensating means 605, the damping gain adjusting means 609 and the adding means 604. In this example, the adding means 604 is an inverting addition amplifying circuit, the phase compensating means 609 is an integrating circuit for delaying the 90 ° phase, and the damping gain adjusting means is realized by a variable resistor.
As described above, in the fifth embodiment, the mechanical vibration detecting means, the phase compensating means, the damping gain adjusting means and the adding means are realized by the operational amplifier, but it is also possible to realize by using a microprocessor or the like. is there. In this case, the mechanical vibration detecting means, the phase compensating means, the damping gain adjusting means and the adding means are realized by software in the microprocessor, and in the case of an analog speed signal, the A / D converter and the D / A are used. A converter may be used to interface with the microprocessor. In the present embodiment, the case of the speed command type motor control device has been described, but the same applies to the torque command input type motor control device only by changing the speed command of FIG. 13 into a torque command signal. It can be implemented.

[0032]

As described above, according to the present invention, a mechanical resonance system can be theoretically clearly separated into an equivalent rigid system and a resonant system. From this fact, the equivalent rigid system physically has a low-pass characteristic. Therefore, by approximating the angular velocity of the motor (the output of the resonance system in Fig. 1) to the observed output, based on the idea that it hardly responds to the resonance component of high frequency, a state observer that models the equivalent rigid system A certain equivalent rigid body observer can be constructed for the first time, so that mechanical vibration components and disturbance components can be detected independently from the angular velocity signal of the motor and the signal proportional to the torque of the motor, and the deterioration of the mechanical resonance signal due to mixing of disturbance components is prevented. it can.

Further, in the present invention, the model of the mechanical vibration system is not used for the equivalent rigid body observer. For this reason, the order of the mechanical vibration detector is always the order of the equivalent rigid body model + 1 even if the resonance is reduced by more than 2 frequencies, and the problems of increasing the amount of calculation and the number of parameters to be set can be solved. There is an effect that it is not necessary to set a mechanical resonance parameter that is difficult to measure. In addition, the equivalent rigid body observer
Since the feedback loop including the first compensating means (proportional) and the second compensating means (integral) is provided, there is also an effect that the detected mechanical vibration signal component is less likely to be affected by the parameter variation of the observer.

Further, as described in the second embodiment, the mechanical resonance detection mounting of the present invention enables detection of high-frequency mechanical vibration (395 Hz) (FIG. 6), and high-speed and high-accuracy mechanical vibration can be detected. Can be detected. Furthermore, as described in the third and fourth embodiments , the vibration damping device suppresses high-frequency mechanical vibrations (395 Hz and 660 Hz) and the speed control system has a wide band (about 3 times and 3.6 times). There is an effect that it can be done at the same time by simple adjustment. In particular, it is considered that the mechanical vibration suppression at a high frequency of 660 Hz in a mechanical system using a ball screw is almost impossible by the conventional technique, and the effect of the present invention is great.

[Brief description of drawings]

FIG. 1 is a block diagram illustrating the principle of the present invention.

FIG. 2 is an explanatory diagram illustrating an example of a resonance mechanism.

FIG. 3 is a block diagram of the first embodiment of the present invention.

FIG. 4 is a circuit diagram of a first embodiment of the present invention.

FIG. 5 is a circuit diagram of a second embodiment of the present invention.

FIG. 6 is a graph showing an experimental result of estimating a mechanical resonance signal.

FIG. 7 is a block diagram of a third embodiment of the present invention.

FIG. 8 is a circuit diagram of a third embodiment of the present invention.

FIG. 9 is a graph showing the experimental results regarding the third example of the present invention.

FIG. 10 is a block diagram of a fourth embodiment of the present invention.

FIG. 11 is a circuit diagram of a fourth embodiment of the present invention.

FIG. 12 is a graph showing experimental results regarding the fourth example of the present invention.

FIG. 13 is a block diagram of a fifth embodiment of the present invention.

FIG. 14 is a circuit diagram showing an example of mechanical vibration detecting means in a fifth embodiment of the present invention.

FIG. 15 is a circuit diagram showing an example of an adding means and a phase compensating means in a fifth embodiment of the present invention.

FIG. 16 is a block diagram of a second conventional example.

[Explanation of symbols]

101 mechanical system including electric motor, 102 mechanical vibration detector, 301 integral element, 302, 303 equivalent rigid body model, 304 inverting amplifier, 305 first compensating means (proportional element), 306 second compensating means (integral element), 40
DESCRIPTION OF SYMBOLS 1 integral arithmetic circuit, 402 arithmetic amplifier circuit, 403 proportional arithmetic circuit, 404 integral arithmetic circuit, 503 proportional integral arithmetic circuit, 601 mechanical vibration suppression device, 602 first connector, 603 second connector, 604 adding means, 605 phase Compensation means, 606 electric motor control device, 6
07 connector, 608 mechanical vibration detection means, 609
Damping gain adjusting means, 610 Command side connector

─────────────────────────────────────────────────── ─── Continuation of the front page (58) Fields surveyed (Int.Cl. ⁷ , DB name) H02P 5/00

Claims (3)

(57) [Claims] 1. A torque of an electric motor when a torque command is given.
Vibration detection device applied to a motor control device for controlling motors
There is, the γ is the reciprocal of the moment of inertia of the motor load, the Do
The viscous friction coefficient of the electric motor and the load is
When the value divided by the load's moment of inertia is input,
Signal multiplied by γ and the output of the integrator multiplied by Do
The difference between the two signals is input to the integrator and the product
An equivalent rigid body model having the output of the divider as its output, a first compensating means for performing proportional amplification, a second compensating means for performing integral calculation, and an estimated disturbance signal which is an output of the second compensating means. The above
A first adder for adding the outputs of the compensation means 1 and the torque command and the output of the first adder
Corresponding to the speed of the electric motor obtained by the second adder used as the input of the equivalent rigid body model and the speed detecting means.
Subtract the output of the equivalent rigid body model from the velocity signal
The obtained mechanical resonance signal to the first compensating means.
And a subtractor for inputting to the second compensating means.
To mechanical resonance detection device. 2. When the torque command is given, the torque of the electric motor is changed.
With the mechanical vibration detection device applied to the motor control device to control
Then , let γ be the reciprocal of the moment of inertia of the motor and load, and Do be
The viscous friction coefficient of the electric motor and the load is
When the value divided by the load's moment of inertia is input,
Signal multiplied by γ and the output of the integrator multiplied by Do
The difference between the two signals is input to the integrator and the product
An equivalent rigid body model having the output of the divider as its output, a first compensating means for performing proportional amplification, a second compensating means for performing integral calculation, and a third compensating means for performing integral calculation or first-order delay calculation. And fourth compensating means for performing integral calculation or first-order delay calculation
And the estimated disturbance signal that is the output of the second compensating means,
A first adder for adding the outputs of the compensation means 1 and the torque command and the output of the first adder
Corresponding to the speed of the electric motor obtained by the second adder used as the input of the equivalent rigid body model and the speed detecting means.
Subtract the output of the equivalent rigid body model from the velocity signal
The obtained mechanical resonance signal to the first compensating means.
A subtracter for inputting to the second compensating means, and the mechanical resonance signal for inputting to the third compensating means,
The output of the third compensating means is input to the fourth compensating means,
The output of the fourth compensation means and the output to the motor control device
A vibration damping control device characterized in that a third adder which receives a Luk command signal and an output signal of the third adder are inputted to a torque command input terminal of the electric motor control device. 3. A mechanical vibration detecting device applied to an electric motor control device for controlling the torque of an electric motor when a torque command is given, wherein γ is the reciprocal of the moment of inertia of the electric motor and the load, and Do is the electric motor and the load. When the viscous friction coefficient is divided by the moment of inertia of the electric motor and the load, the difference between the signal obtained by multiplying the input signal by γ and the signal obtained by multiplying the output of the integrator by Do is input to the integrator. In addition, an equivalent rigid body model having the output of the integrator as its output, a first compensating means for performing proportional amplification, a second compensating means for performing an integral calculation, and a third compensating operation or a first-order delay calculation Compensating means, fourth compensating means for performing integration calculation or first-order lag calculation, and first addition for adding the estimated disturbance signal which is the output of the second compensating means and the output of the first compensating means. With a vessel Said equivalent rigid body from the second adder and the speed signal corresponding to the speed of the motor obtained by the speed detecting means for the input of said equivalent rigid body model by adding the output of the torque command first adder A subtracter for subtracting the output of the model and inputting the obtained mechanical resonance signal to the first compensation means and the second compensation means; and a mechanical resonance signal input to the third compensation means for the third compensation means. A fourth adder which receives the output of the compensating means and the torque command signal to the electric motor control device, and a configuration which inputs the output signal of the fourth adder to the speed input terminal of the electric motor control device. A vibration damping control device in an electric motor control system characterized by the above.