JPH11332272A - Dual inertia resonance system torque control method - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress a shaft twisting vibration and to simultaneously improve command followability of a dual inertia resonance torque control by using a PID-P control for applying a feedback proportional compensation of a shaft torque or the like detected by a torque motor at an output of a PID controller. SOLUTION: A PID-P torque control system is constituted by feeding back a shaft torque detected by a torque meter for a torque command, calculating deviation, amplifying the deviation by a PID controller, amplifying the torque according to the gain of a feedback proportional compensator, obtaining the deviation of the output of the compensator from an output of the PID controller, and using the deviation as a motor torque command. The method has good command followability of the shaft torque even near a resonance frequency simultaneously by suppressing a shaft twisting vibration by obtaining gains of the PID-P control as a function of a machine constant of the dual inertial resonance system, further setting a proportional item of a transfer function of the feedback to zero, thereby minimizing the gains of the PID-P control.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a torque control method for a two-inertia resonance system in which a motor and a load are connected by a low-rigidity elastic shaft.

[0002]

2. Description of the Related Art In general, in a motor drive system in an industrial plant or an industrial robot, if a motor and a load are connected by a low-rigidity elastic shaft, a resonance system is formed, which causes a problem of torsional vibration. There is. The outline will be described with reference to FIGS. FIG. 3 shows a two-mass resonance system, wherein 1 is an electric motor, 2 is a load, and 3 is an elastic shaft.
When the mechanical system is coupled with the elastic shaft 3 as described above, the mechanical system has a torsional vibration mode, and becomes a two-inertial resonance system. FIG. 4 is a block diagram showing the two inertial resonance system of FIG.
become. Where Tm is the motor torque command, Tc is the shaft torque, TL is the disturbance torque on the load side, ωm is the motor speed,
ωL is the load speed, θc is the torsion angle of the shaft, Kc is the spring constant of the shaft, Dc is the viscosity coefficient of the shaft, Jm is the motor inertia, and JL is the load inertia. In FIG. 4, as a torque transmission characteristic of the open loop system, a transfer function G (s) from the motor torque command Tm to the shaft torque Tc is given by the following equation (1).

[0003]

(Equation 1)

Where s is the Laplace operator, ωo and ζo
Is the natural resonance frequency and damping coefficient of the two inertial resonance system,
They are expressed by the equations (2) and (3) shown in Equation 2, respectively.

[0005]

(Equation 2)

In the two-inertia resonance system torque control, in order to make the shaft torque Tc follow the torque command T * quickly and without vibration, the shaft torque Tc is calculated from the torque command T *.
The transfer function Φ (s) must have a desirable frequency response gain characteristic as shown by the solid line (a) in FIG. That is, in the frequency band from the frequency 0 to the vicinity of the resonance frequency ωo, the gain characteristic must always have a value close to the constant 0 dB. FIG. 5B shows a phase characteristic corresponding to the gain characteristic shown in FIG.

[0007]

In general, the viscosity coefficient D of a shaft
c is a very small value, and as can be seen from equation (3), since the damping coefficient ζo is small, the gain characteristic | G (jω) | of the open-loop frequency response is represented by a dotted line (b) in FIG. ), A high peak occurs in the gain characteristic near the resonance frequency ωo, and the axial torsional resonance at the frequency ωo is likely to occur. As a result, it is possible to obtain a desirable response characteristic as shown by the solid line (a) in FIG. Can not. Conventionally, PID (proportional-integral-
Differential) control has been used, but with the development of modern control theory in recent years, H と し た control, which is a theory relating to shaping of the frequency response of a control system, and control in combination with PID have been widely studied. Further, when the PID control is applied to the torque control of the two inertial resonance system, the torsional vibration of the shaft can be easily suppressed.
In order to increase the responsiveness of the response to follow the torque command of the shaft torque, it is necessary to set each gain value of the PID controller large. However, if the gain of the controller is set too large, it is difficult to apply to a real machine due to hardware limitations such as a limiter.

On the other hand, when the H∞ control theory is used, a controller having good command followability and robust stability can be designed. However, since the dimension of the H∞ controller generally increases, a high-speed, high-performance C
A PU is required, and it is difficult to apply to a real machine from the viewpoint of cost and software. SUMMARY OF THE INVENTION The present invention has been made in view of the above-described problems of the related art, and aims at suppressing shaft torsional vibration and improving command followability of shaft torque. In Claim 1, there is no shaft torsional vibration without using a controller using a functional CPU and using only a controller with a minimized gain.
In addition, the present invention provides a two-inertia resonance system torque control with good command followability of the shaft torque even in the vicinity of the natural resonance frequency ωo, and when the shaft torque Tc cannot be detected,
Using the estimated shaft torque Tce by the disturbance observer 2
It provides inertial resonance system torque control.

[0009]

According to the present invention, in order to achieve the above object, the torque command T
* And deviation from shaft torque Tc detected by torque meter
(PID controller Gpid (s) having T as input, and feedback proportional compensation section (Kf) for shaft torque.
A two-inertial-resonance-system torque control method using a means for controlling a shaft torque Tc using a deviation between an output of a D controller and an output of a feedback proportional compensator as a motor torque command Tm. In the case of a system that does not have this, the shaft torque Tc detected by the torque meter described in claim 1 cannot be used. Therefore, a disturbance observer attached to the motor side is provided, and the estimated value Tce of the shaft torque is set.
This is a two-inertial-resonant-system torque control method using the above as a compensation means.

[0010]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The present invention provides a gain characteristic of a frequency response which is a cause of vibration in order to achieve the above-mentioned object of suppressing the shaft torsional vibration and the purpose of causing the shaft torque to quickly follow the command torque. An object of the present invention is to solve the problem by providing a compensator that suppresses the peak and has good command followability of the shaft torque even near the natural resonance frequency. The details of these means will be described below with reference to the drawings. FIG. 1 is a block diagram for explaining claim 1 of the present invention. In FIG. 1, the gain characteristic of the frequency response of the closed loop system transfer function Φ (s) from the torque command T * to the shaft torque Tc has a high peak. In the case of a system having a torque meter in order to prevent the occurrence of the torque error, as shown in FIG. 1, the torque control unit 2 inputs the deviation (T) between the torque command T * and the shaft torque Tc detected by the torque meter. PID controller Gpi
d (s) and a feedback proportional compensator 3 (Kf) that receives the shaft torque Tc as input, obtains a deviation between the output of the PID controller and the output of the feedback proportional compensator 3, and outputs the deviation. Is a motor torque command Tm, thereby forming a two-inertial resonance system torque control system. P
The transfer function expression of the ID controller Gpid (s) can be expressed as the following equation (4).

[0011]

(Equation 3)

Here, Kp, Ki and Kd are each P
The proportional gain, the integral gain, and the differential gain of the ID controller, and s is a Laplace operator. The determination of each of the gains Kp, Ki, and Kd of the PID controller can be performed by, for example, a Manabe polynomial based on a coefficient diagram. For a detailed explanation of the coefficient projection and the Manabe polynomial, see Manabe's "Unified Interpretation of Classical Control, Optimal Control, and H∞ Control" (1991
Journal of the Society of Measurement and Control in October 30-10) and Mr. Manabe's "Design of a Two-Inertia Resonant System Controller Using the Coefficient Diagram" (January 1998, 118-D-1).
It is known. Here, the outline of the coefficient projection will be briefly described.

The coefficient diagram method is a kind of algebraic design method on a polynomial ring, and is characterized in that a characteristic polynomial and a controller are simultaneously designed using a coefficient diagram with the appropriateness of the form as a measure. The various mathematical relationships used in the CDM are listed below. It is assumed that a characteristic polynomial Δ (s) is given to an n-order closed-loop system as shown in Expression (5) shown in Expression 4.

[0014]

(Equation 4)

The stability index γi indicating the stability and response speed of the control system and the equivalent time constant τ are defined as shown in equations (6) and (7).

[0016]

(Equation 5)

In the coefficient map method, the standard form recommended by Manabe is represented by the following equation (8), which is called a Manabe polynomial.

[0018]

(Equation 6)

Hereinafter, a method for determining the gain of the controller of the two-inertial-resonant-system torque control using the Manabe polynomial based on the above-described coefficient diagram will be specifically described. Since the value of the viscosity coefficient Dc of the shaft is very small, ignoring and setting Dc = 0 (ζo = 0), the closed loop system from the torque command T * to the shaft torque Tc with respect to the torque control system shown in FIG. Transfer function Φ
(S) is given by Expression (9) shown in Expression 7.

[0020]

(Equation 7)

Here, the characteristic polynomial Δ (s) of the closed loop system
Is given by Expression (10) shown in Expression 8.

[0022]

(Equation 8)

First, a torque control system of a two inertial resonance system based on only the PID control is designed. That is, Kf = 0 is set. At this time, the condition for the Manabe polynomial is shown in Equation 9 (1
1)

[0024]

(Equation 9)

Where τ is an equivalent time constant, γ1, γ
2 is a stability index. As a solution of equation (11), PID
Each of the gains Kp, Ki and Kd of the controller is determined by the equation (12) shown in Expression 10 as a function of the mechanical constant of the two inertial resonance system and the equivalent time constant τ.

[0026]

(Equation 10)

Normally, since the proportional gain Kp ≧ 0, (1
The equivalent time constant τ is expressed by the following equation 11 by the equation (2) (1
3) It can be expressed as an equation.

[0028]

[Equation 11]

As can be seen from equation (12), the equivalent time constant τ
Is too small, the gains Kp,
Ki and Kd increase. Therefore, P with a large gain
The equivalent time constant τ is set to 5 / (√
2ωo). In particular, τ = 5 / (√2
ωo), each gain Kp, K of the PID controller
i and Kd are minimized, and can be obtained only from the mechanical constants of the two inertial resonance system as shown in the following equation (14).

[0030]

(Equation 12)

Since the proportional gain Kp becomes 0, the PID controller can be simplified as an ID controller. In the two-inertial-resonant-system torque control using only the ID controller designed as described above, as shown by the broken line (c) in FIG. 5A, the gain characteristic of the frequency response from the torque command T * to the shaft torque Tc is high. Since there is no peak, the axial torsional vibration can be suppressed, but the gain characteristic is greatly reduced from 0 dB at the frequency of ω = ωo / √5, which is a large difference from the desired frequency gain characteristic of the solid line (a) in FIG. , The command followability of the shaft torque is not good. The cause of the decrease in the gain characteristic at the frequency of ω = ωo / √5 can be clarified from the frequency response Φ (jω) of the transfer function of the closed loop system represented by the expression (15) shown in Expression 13.

[0032]

(Equation 13)

Where ω is the frequency, Rm (ω) and Im
(Ω) is the real and imaginary parts of the numerator polynomial M (ω), respectively, and Rn (ω) and In (ω) are the denominator polynomial N
The real part and the imaginary part of (ω). τ = 5 / (√2ωo)
(That is, Kp = Kf = 0), (1
4) Substituting equation (9) into equation (9), and Laplace operator s
If the operator is replaced by jω, the frequency response of the numerator and the denominator polynomial of the equation (15) can be expressed by the following equations (16) and (17), respectively.

[0034]

[Equation 14]

From equations (15) to (17), the numerator and denominator have the same real part (Rm (ω) = Rn (ω)),
Further, since the imaginary part of the numerator is Im (ω) = O, the gain characteristic of the frequency response of the closed loop system is | Φ (jω) | ≦ 1
In particular, Φ (jωo) = 1 (= od) at the resonance frequency ωo as indicated by a broken line (c) shown in FIG.
B), it can be seen that the torsional vibration can be suppressed only by the ID control. However, at the frequency of ω = ωo / √5, Φ (jωo / √5) = 0, so that the command followability of the shaft torque becomes worse around this frequency. Further, even in the case of the PID control in which τ is a value close to 5 / (√2ωo), the command followability of the shaft torque is deteriorated.

In order to correct the reduction of the gain characteristic near the frequency, the shaft torque amplified by the gain Kf is fed back to the output of the PID controller as shown in FIG.
That is, PID-P control is configured by adding proportional compensation as Kf ≠ 0. When the respective gains of the PID-P control are obtained by the Manabe polynomial, Ki and Kd are the same as in the expression (12), but Kp and Kf are obtained by the expression (18) shown in Expression 15.

[0037]

(Equation 15)

When the same equivalent time constant τ = 5 / (√2ωo) as that of the ID control is set to minimize the gain of the controller, from the above equation (18), Kp + is added to the PID-P control.
It is sufficient to set Kf = 0. At this time, in the frequency response Φ (jω) of the closed loop system by the PID-P control, the denominator polynomial N (ω) is the same as the equation (17), but Kp = −Kf ≠.
Because of 0, the numerator polynomial M (ω) differs from equation (16),
This is expressed by the following equation (19).

[0039]

(Equation 16)

Under the condition of Kp + Kf = 0, the PID-P control has the same closed-loop characteristic polynomial as the ID control. However, by appropriately setting the value of the gain Kp (= −Kf), ω = The problem of reduction of the gain characteristic at the frequency of ωo / √5 can be solved. Where Φ
The value of Kp may be determined so that (jωo / √5) = 1. As can be seen from equations (19) and (17), Rm
Since (ω) = Rn (ω), Φ (jωo / √5) =
In order to set it to 1, Im (ωo / √5) = Inωo / √
The value of Kp may be determined so as to satisfy 5). From this, Kp and Kf are determined as functions of the mechanical constants of the two-inertial-resonant system as shown in equation (20) shown in equation (17).

[0041]

[Equation 17]

When Ki and Kd determined by the above equation (14) and Kp and Kf (= −Kp) determined by the above equation (20) are applied to the torque control of the two inertial resonance system shown in FIG. The response characteristic Φ (jω) is as shown by the solid line (C) in FIG. The frequency that intersects 0 dB in the gain characteristic curve is obtained from | Φ (jω) | = 1 as shown in Expression (21) shown in the following Expression 18.

[0043]

(Equation 18)

Here, the frequency ω3 having the largest value is called the crossover frequency ωcg. The frequencies ωa and ωb at which the value of the gain characteristic curve | Φ (jω) | is minimized and maximized in the frequency band of 0 to ωcg are shown in the following Expression 19 from d | Φ (jω) | / dω = 0 (22) It can be calculated like an equation.

[0045]

[Equation 19]

The minimum value and the maximum value corresponding to the frequency in the above equation (22) are obtained as in equation (23) shown in equation (20).

[0047]

(Equation 20)

From the above equation (23), the gain characteristic curve in the frequency band of 0 to ωcg is | Φ (jω) | is 1 (0 dB).
Since the value is close to the desired frequency response gain characteristic, the problem of reducing the gain characteristic by only the ID control while suppressing the shaft torsional vibration is solved, and the command followability of the shaft torque is improved.

The above-described two-inertia resonance-system torque control is according to claim 1 of the present invention in which a torque meter is provided so that the shaft torque Tc can be detected. However, a system without a torque meter is actually provided. In many cases, the shaft torque Tc cannot be detected. Therefore, hereinafter, a disturbance observer attached to the motor side is provided instead of the shaft torque Tc detected by the torque meter, and an estimated value Tce of the shaft torque is calculated and used as a compensation means. The method will be described.

In the case where the shaft torque Tc cannot be detected as in a system without a torque meter, as shown in FIG. 2, the torque control of the shaft torque Tc detected by the torque meter is performed in the two inertial resonance system torque control of the present invention. Instead, the shaft torque Tce estimated by a disturbance observer attached to the motor can be applied. The disturbance observer is configured by calculating an estimated value Tce of a shaft torque as a disturbance of the electric motor by using information of the motor torque command Tm and the electric motor speed ωm, as shown in the disturbance observer unit 4 in FIG.
As shown in FIG. 2, a torque command T *
(A PID controller Gpid (s) having ΔT as an input and a deviation from the shaft torque Tce estimated by the disturbance observer 4 attached to the motor side, and the estimated shaft torque T
Feedback proportional compensator 3 (Kf) with ce as input
And calculating the deviation between the output of the PID controller and the output of the feedback proportional compensator, and using the deviation as the two-inertial-resonant-system motor torque command Tm to control the shaft torque. A resonance system torque control system is configured.

In order to quickly and accurately estimate the shaft torque,
The filter time constant Tf of the disturbance observer must be sufficiently fast. When Tf ≒ 0, the estimated shaft torque Tce is almost the same as the actual shaft torque Tc. Therefore, the determination of each gain of the PID-P controller of the torque control system in FIG.
Kp (= −Kf) is calculated as (2)
It can be determined as a function of the mechanical constant of the two inertial resonance system by the equation (0).

FIG. 6 shows a PI to which the estimated shaft torque Tce is applied.
3 shows a frequency response characteristic of a closed loop system by the DP control. Set the filter time constant Tf of the disturbance observer to Tf = 1.
ms and Tf = 10 ms are shown in FIG.
As shown by the solid line (d) and the dashed line (f) shown in FIG. 5, when a disturbance observer (Tf = 10 ms) having a slow time constant is applied,
Since the estimated shaft torque Tce cannot quickly and accurately reproduce the actual shaft torque Tc, a high peak occurs in the gain characteristic (f) of the frequency response. However, the dotted line (e) in FIG.
It is an open-loop system frequency response characteristic G (jω) of an inertial resonance system.

In summary, according to the torque control method of the two inertial resonance system of the present invention, when the shaft torque can be detected as in a system having a torque meter, the torque control system is controlled by the PID as shown in FIG. -P control. First, P is calculated by the equation (14) using a Manabe polynomial based on a coefficient diagram.
The integral gain Ki and the differential gain Kd of the ID-P control are obtained as a function of the mechanical constant of the two inertial resonance system, and the frequency ω = ωo
In order to correct the reduction of the closed-loop frequency response characteristic around / √5, the proportional gain Kp of the PID-P control and the feedback proportional compensation gain Kf of the shaft torque Tc are set to (2
Means for obtaining as a function of the mechanical constant of the two inertial resonance system by the equation (0) is used. When a shaft torque cannot be detected as in a system without a torque meter, a torque control system is configured as shown in FIG. 2, the shaft torque is estimated by a disturbance observer having a fast time constant, and PID-P control is performed. Each of the gains takes a means which is obtained as a function of the mechanical constant of the two-inertial-resonant-system by the equations (14) and (20).

Hereinafter, specific embodiments of the present invention will be further described with reference to numerical examples. The mechanical constants of the two-inertia resonance system as numerical examples are the respective gains Kp, Ki, and Kd of the PID-P control when the motor inertia, load inertia, and shaft spring constant are set to the values of Expression (24) shown in Equation 21 below. And an example of determining Kf will be described.

[0055]

(Equation 21)

The two inertial resonance system having the above mechanical constant is
From equations (2) and (3), ωo = 77.3 [rad / s
ec] and a damping coefficient of ζo = 0.038 (when the viscosity coefficient Dc of the shaft is set to 0.3 [Nmsec / rad]). Since the damping coefficient ζo is small, the frequency response characteristic (torque transmission characteristic from Tm to Tc) of the open-loop system has a peak at the resonance frequency ωo as shown by the dotted line (b) in FIG. Occurs. The time response of the shaft torque Tc to the step-like motor torque command Tm (4Nm) and the disturbance torque TL (4Nm) is such that the torsional vibration of the shaft is generated as shown by the dotted lines (h) and (g) shown in FIGS. I do.

When an ID controller having an equivalent time constant τ of τ = 5 / (√2ωo) is applied to the above-described two-inertial-resonant-system torque control, each gain of the ID controller is obtained from the above equation (14). Ki = 33.23 and Kd = 0.028. This I
When only the D controller is applied to the two inertial resonance system torque control,
The closed loop frequency response characteristic from the torque command T * to the shaft torque Tc is as shown by a broken line (C) in FIG. Although there is no peak in the gain characteristic, the gain characteristic is significantly reduced from 0 dB at the frequency of ω = ωo / √5, and the command followability of the shaft torque is not good. The time response of the shaft torque Tc to the step-like torque command T * and the disturbance torque TL is shown in FIG. 7 (the shaft torque Tc detected by the torque meter).
c) and FIG. 8 (when using the shaft torque Tce estimated by a disturbance observer with a time constant Tf = 1 ms), the torsional vibration of the shaft does not occur as shown by broken lines (i) and (nu). Slow response at rising.

The PID-P control of the present invention is applied to the torque control system in order to correct the reduction of the gain characteristic near the frequency of ω = ωo / √5 in the frequency response characteristic of the closed loop system only by the ID control described above. Then, the shaft torque Tc detected by the torque meter (or the shaft torque Tce estimated by the disturbance observer) is amplified by the gain Kf, and the difference between the output of the feedback proportional compensator Kf and the output of the PID controller is calculated by two inertial resonances. Motor torque command Tm
And Calculating according to the above equation (20), the values of the proportional gain Kp and the feedback proportional compensation gain Kf are Kp = 1.22 and Kf = −Kp = −1.22, respectively.

When the PID-P control of the present invention is applied to the two-inertia resonance system torque control, the closed loop system frequency response characteristic from the torque command T * to the shaft torque Tc is shown in FIG. 5 (shaft torque detected by a torque meter). Tc) and FIG. 6 (when the shaft torque Tce estimated by a disturbance observer with a time constant Tf = 1 ms) is shown by solid lines (c) and (d), and the gain characteristic has a maximum value | Φ
(Jωb) | = 1.3 (= 2.29 dB), but the minimum value is | Φ (jωa) | = 0/91 (=
−0.82 dB), the problem of reducing the gain in the frequency band of 0 to ωcg is substantially solved while suppressing the torsional vibration of the shaft, and it is understood that the command followability of the shaft torque is better than the torque control by the ID control alone. . Further, a step-like torque command T * of the shaft torque Tc and a disturbance torque T
The time response to L is shown in FIG. 7 (when the shaft torque Tc detected by the torque meter is used) and FIG. 8 (time constant Tf =
Shaft torque Tc estimated by disturbance observer of 1 ms
As shown by solid lines (g) and (g) shown in (e), there is no torsional vibration of the shaft and the response is fast convergence, and it can be seen that the two inertial resonance system torque control with good command followability of the shaft torque can be performed.

[0060]

As described above, according to the present invention, 2
An inertial resonance torque control system is provided by using a PID- which adds feedback proportional compensation of shaft torque to the output of a conventional PID controller.
P control, each gain of PID-P control is obtained as a function of the mechanical constant of the two inertial resonance system, and (Kp + K
By setting f) to zero, the control gains (Kp, Ki, Kd,
By minimizing Kf), it is possible to provide torque control of a two inertial resonance system without shaft torsional vibration and with good command followability of shaft torque.

[Brief description of the drawings]

FIG. 1 is a block diagram for explaining claim 1 of the present invention (when a shaft torque Tc detected by a torque meter is used).

FIG. 2 is a block diagram for explaining claim 2 of the present invention (the shaft torque Tce estimated by a disturbance observer);
If you use).

FIG. 3 is a diagram showing a two inertial resonance system.

FIG. 4 is a block diagram of a two inertial resonance system.

FIG. 5 is a diagram showing a frequency response characteristic for explaining the effect of claim 1 of the present invention (when using a shaft torque Tc detected by a torque meter).

FIG. 6 is a diagram showing a frequency response characteristic for explaining the effect of the second aspect of the present invention (when the shaft torque Tce estimated by a disturbance observer is used).

FIG. 7 is a diagram showing a time response according to claim 1 of the present invention (when a shaft torque Tc detected by a torque meter is used).

FIG. 8 is a diagram showing a time response according to claim 2 of the present invention (when using a shaft torque Tce estimated by a disturbance observer).

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 2 inertial resonance system which has an elastic axis 2 Torque controller which applied PID controller 3 Feedback proportional compensation part of axis torque 4 Disturbance observer part Jm Motor inertia Jmn Nominal value of motor inertia JL Load inertia Kc Spring constant of axis Dc axis Coefficient of viscosity T * Torque command Tm Motor torque command Tc Shaft torque Tce Estimated value of shaft torque TL Disturbance torque on load side ΔT Deviation between torque command and shaft torque (or estimated value of shaft torque) ωm Motor speed ωL Load speed θc axis torsion angle Gpid (s) PID controller Kp Proportional gain of PID controller Ki Integral gain of PID controller Kd Differential gain of PID controller Kf Feedback proportional compensation gain of axis torque Tf Time constant of filter of disturbance observer ωo 2 Natural resonance frequency of inertial resonance system ζo Packaging coefficient ωcg gain crossover frequency τ equivalent time constant γi stability index

────────────────────────────────────────────────── ───

[Procedure amendment]

[Submission date] July 10, 1998

[Procedure amendment 1]

[Document name to be amended] Statement

[Correction target item name] 0009

[Correction method] Change

[Correction contents]

[0009]

According to the present invention, in order to achieve the above object, the torque command T
* And deviation from shaft torque Tc detected by torque meter
(PID controller Gpid (s) having ΔT as input, and feedback proportional compensation section (Kf) for shaft torque.
A two-inertial-resonant-system torque control method using a means for controlling a shaft torque Tc with a deviation between an output of an ID controller and an output of a feedback proportional compensator as a motor torque command Tm,
Further, in the case of the system in which the torque meter is not provided in claim 2, since the shaft torque Tc detected by the torque meter described in claim 1 cannot be used, a disturbance observer attached to the electric motor is provided, and Estimated value T
This is a two-inertial-resonance-system torque control method that calculates ce and uses it as compensation means.

[Procedure amendment 2]

[Document name to be amended] Statement

[Correction target item name] 0010

[Correction method] Change

[Correction contents]

[0010]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The present invention provides a gain characteristic of a frequency response which is a cause of vibration in order to achieve the above-mentioned object of suppressing the shaft torsional vibration and the purpose of causing the shaft torque to quickly follow the command torque. An object of the present invention is to solve the problem by providing a compensator that suppresses the peak and has good command followability of the shaft torque even near the natural resonance frequency. The details of these means will be described below with reference to the drawings. FIG. 1 is a block diagram for explaining claim 1 of the present invention. In FIG. 1, the gain characteristic of the frequency response of the closed loop system transfer function Φ (s) from the torque command T * to the shaft torque Tc has a high peak. In the case of a system equipped with a torque meter in order to prevent the occurrence of torque, as shown in FIG. 1, the torque control unit 2 inputs the deviation (ΔT) between the torque command T * and the shaft torque Tc detected by the torque meter. PID controller Gp
id (s), and a feedback proportional compensator 3 (Kf) that receives the shaft torque Tc as input.
A deviation between the output of the controller and the output of the feedback proportional compensator 3 is obtained, and the output is used as a motor torque command Tm to constitute a two-inertial resonance torque control system.
The transfer function expression of the PID controller Gpid (s) is given by
Can be expressed as shown in equation (4).

[Procedure amendment 3]

[Document name to be amended] Statement

[Correction target item name] 0035

[Correction method] Change

[Correction contents]

From equations (15) to (17), the numerator and denominator have the same real part (Rm (ω) = Rn (ω)),
Further, since the imaginary part of the numerator is Im (ω) = O, the gain characteristic of the frequency response of the closed loop system is | Φ (jω) | ≦ 1
In particular, Φ (jωo) = 1 (= 0d) at the resonance frequency ωo as indicated by a broken line (c) shown in FIG.
B), it can be seen that the torsional vibration can be suppressed only by the ID control. However, at the frequency of ω = ωo / √5, Φ (jωo / √5) = 0, so that the command followability of the shaft torque becomes worse around this frequency. Further, even in the case of the PID control in which τ is a value close to 5 / (√2ωo), the command followability of the shaft torque is deteriorated.

[Procedure amendment 4]

[Document name to be amended] Statement

[Correction target item name] 0040

[Correction method] Change

[Correction contents]

Under the condition of Kp + Kf = 0, the PID-P control has the same closed-loop characteristic polynomial as the ID control. However, by appropriately setting the value of the gain Kp (= −Kf), ω = The problem of reduction of the gain characteristic at the frequency of ωo / √5 can be solved. Where Φ
The value of Kp may be determined so that (jωo / √5) = 1. As can be seen from equations (19) and (17), Rm
Since (ω) = Rn (ω), Φ (jωo / √5) =
In order to set it to 1, Im (ωo / √5) = In (ωo / √
The value of Kp may be determined so as to satisfy 5). From this, Kp and Kf are determined as functions of the mechanical constants of the two-inertial-resonant system as shown in equation (20) shown in equation (17).

[Procedure amendment 5]

[Document name to be amended] Statement

[Correction target item name] 0059

[Correction method] Change

[Correction contents]

When the PID-P control of the present invention is applied to the two-inertia resonance system torque control, the closed loop system frequency response characteristic from the torque command T * to the shaft torque Tc is shown in FIG. 5 (shaft torque detected by a torque meter). Tc) and FIG. 6 (when the shaft torque Tce estimated by a disturbance observer with a time constant Tf = 1 ms) is shown by solid lines (c) and (d), and the gain characteristic has a maximum value | Φ
(Jωb) | = 1.3 (= 2.29 dB), but the minimum value is | Φ (jωa) | = 0.91 (=
−0.82 dB), the problem of reducing the gain in the frequency band of 0 to ωcg is substantially solved while suppressing the torsional vibration of the shaft, and it is understood that the command followability of the shaft torque is better than the torque control by the ID control alone. . Further, a step-like torque command T * of the shaft torque Tc and a disturbance torque T
The time response to L is shown in FIG. 7 (when the shaft torque Tc detected by the torque meter is used) and FIG. 8 (time constant Tf =
Shaft torque Tc estimated by disturbance observer of 1 ms
As shown by solid lines (g) and (g) shown in (e), there is no torsional vibration of the shaft and the response is fast convergence, and it can be seen that the two inertial resonance system torque control with good command followability of the shaft torque can be performed.

[Procedure amendment 6]

[Document name to be amended] Drawing

[Correction target item name] Figure 2

[Correction method] Change

[Correction contents]

FIG. 2

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Claims (2)

[Claims] In a two-inertial resonance torque control for transmitting a drive torque from an electric motor to a load via an elastic shaft, a torque command T * and a shaft torque T detected by a torque meter are provided.
PID controller Gpid which receives deviation ΔT from c as input
(S) and a feedback proportional compensator (Kf) that receives the shaft torque Tc as an input, obtains a deviation between the output of the PID controller and the output of the feedback proportional compensator, and calculates the deviation as A two-inertia resonance system constitutes a torque control system with a motor torque command Tm, and the respective gains (Kp, Ki, Kd) of the PID controller and the gain (Kf) of the feedback proportional compensator are changed by the two-inertial resonance system. As a function of the mechanical constant, the shaft torque T
A two inertial resonance system characterized by minimizing the controller gain (Kp, Ki, Kd, Kf) by setting the proportional term (Kp + Kf) of the feedback transfer function from c to the motor torque command Tm to zero. Torque control method. 2. The motor torque command Tm and the motor speed ωm
The torque control method according to claim 1, wherein a disturbance observer is provided as an input, and an estimated shaft torque Tce is calculated by the disturbance observer, and is used instead of the shaft torque Tc.