JP2022055235A - Control device - Google Patents

Abstract

To suppress vibration due to resonance between a driving system and a load motor.SOLUTION: A control device 10 comprises: an electric inertia torque calculation unit 11 which calculates an electric inertia torque τJ on the basis of a rotary speed ωG of a load motor and an electric inertia value Ja of the load motor; an electric friction torque calculation unit 12 which calculates an electric friction torque τD on the basis of the rotary speed ωG of the load motor and an electric friction value Da of the load motor; a torque command generation unit 13 which generates a torque command τin on the basis of the electric inertia torque τJ and the electric friction torque τD; and a correction unit 14 which corrects the electric inertia value Ja and the electric friction value Da on the basis of a sampling period Ts.SELECTED DRAWING: Figure 11

Description

The present invention relates to a load motor control device that is connected to a drive system via a twisting element and reproduces the load of the drive system.

In a single verification of an engine that is an internal combustion engine, the performance of the engine is tested by connecting the engine as a drive system and a dynamometer equipped with measuring equipment such as a load motor and a torque meter that reproduces the load of the engine. (See, for example, Non-Patent Document 1).

Masayuki Kobayashi, Nobuyoshi Okui, "Examination of HILS that enables evaluation of hybrid heavy-duty vehicle models using actual engines", Proceedings of the Society of Automotive Engineers of Japan, vol.47, No.5, pp.1185-1190 (2016)

FIG. 1 shows a configuration example of the engine test device used for the above-mentioned unit verification of the engine.

As shown in FIG. 1, the engine test apparatus 1 includes an engine 2, a dynamometer 3, a shaft, and a coupling 4 connecting them. The engine test device 1 shown in FIG. 1 can be modeled as a bi-inertial resonance system by the twisting element of the coupling 4. FIG. 2 is a block diagram in which the engine test device 1 shown in FIG. 1 is modeled as a bi-inertial resonance system.

In FIG. 2, J _E is the moment of inertia of the engine 2. _DE is the coefficient of viscous friction of the engine 2. J _G is the moment of inertia of the load motor that reproduces the load of the engine 2 provided in the dynamometer 3. _DG is the coefficient of viscous friction of the load motor. K _s is the spring constant of the coupling 4. τ _E is the torque (engine input) of the engine 2. τ _s is the shaft torsion torque. The shaft torsion torque τ _s is measured by a torque meter (not shown). ω _E is the rotation speed of the engine 2. ω _G is the rotation speed of the load motor. s is a Laplace operator.

When the engine test device 1 is defined as the bi-inertial resonance system shown in FIG. 2, the bi-inertial resonance system is the spring constant K of the engine 2 as a drive system, the inertial moments J _E , J _G of the load motor, and the coupling 4. It has a resonance frequency corresponding to _s . When the vibration frequency of the torque generated by the engine 2 and the resonance frequency match, the torque vibration applied to the coupling 4 and the shaft is amplified, and a large load is applied. Therefore, it is required to perform resonance suppression control so that torque vibration is not amplified at the resonance frequency.

As a conventional method of resonance suppression control in a bi-inertial resonance system, there are methods such as control design by Manabe polynomial, resonance ratio control, and H∞ control. However, in the engine test apparatus 1 shown in FIG. 1, the torque generated by the engine 2 and the rotation speed of the engine 2 cannot be acquired, and only the torque command of the load motor can be operated. Therefore, the conventional method as described above cannot be used.

An object of the present invention made in view of the above problems is to provide a control device capable of suppressing vibration due to resonance between a drive system and a load motor.

In order to solve the above problems, the control device according to the present invention is a control device that is connected to a drive system via a twisting element and controls a torque command of a load motor that reproduces the load of the drive system at a predetermined sampling cycle. Therefore, an electric inertia torque calculation unit that calculates the electric inertia torque of the load motor based on the rotation speed of the load motor and the electric inertia value of the load motor, the rotation speed of the load motor, and the load motor. An electric friction torque calculation unit that calculates the electric friction torque of the load motor based on the electric friction value of the above, and a torque command generation unit that generates the torque command based on the electric inertia torque and the electric friction torque. A correction unit for correcting the electric inertia value and the electric friction value based on the sampling period.

Further, in the control device according to the present invention, the electric inertia value and the electric friction value are J _a and D _a , respectively, and the electric inertia value and the electric friction value that change depending on the delay time due to the sampling cycle are J a ₂ and D, respectively. _{Let a2} be, the angular frequency of the input torque of the drive system be ω, the sampling cycle be T _s , and the delay time due to the sampling cycle T _s be ΔT _s .
It is preferable that the correction unit corrects the electric inertia value J _a and the electric friction value D _a so as to satisfy the following equations.

According to the control device according to the present invention, vibration due to resonance between the drive system and the load motor can be suppressed.

It is a figure which shows the structural example of an engine test apparatus. It is a block diagram which shows the structure of the bi-inertial resonance system which modeled the engine test apparatus shown in FIG. 1. It is a block diagram which added the electric inertia and the electric friction of a load motor to the bi-inertia resonance system shown in FIG. It is a figure which described the relationship between the electric inertia torque and the electric friction torque by a vector when the electric inertia and the electric friction are separated. It is a figure which described the relationship between the electric inertia torque and the electric friction torque by the element of a vector when the electric inertia and the electric friction are separated. It is a figure which shows the relationship between a phase lag and a torque command of a load motor. It is a figure which shows the relationship between the electric inertia torque _{τ'J, the phase component τ JJ without} the phase lag, and the phase lag component τ _JD . It is a figure which shows the relationship between the electric friction torque _{τ'D, the phase component τ DD without} the phase lag, and the phase lag component τ _DJ . It is a figure which shows the relationship between a phase lag and each torque in a positive electric inertia. It is a figure which shows the relationship between a phase lag and each torque in a negative electric inertia. It is a block diagram which shows the structure of the bi-inertial resonance system in which continuous system control is performed. It is a block diagram which shows the structure of the bi-inertial resonance system which performs discrete system control. It is a Bode diagram which shows the relationship from the engine input to the shaft torsion torque in the bi-inertial system shown in FIGS. 2, 9A and 9B. It is a figure which shows the structural example of the control apparatus which concerns on one Embodiment of this invention.

Hereinafter, embodiments for carrying out the present invention will be described with reference to the drawings.

The control device according to the embodiment of the present invention controls the torque command of the load motor for reproducing the load of the engine 2 which is the drive system provided in the dynamometer 3 in the engine test device 1 shown in FIG. It suppresses vibration due to resonance between the engine 2 and the load motor. First, resonance suppression control in the bi-inertial resonance system shown in FIG. 2, which models the engine test device 1, will be described.

The transfer function from the engine input τ _E of the bi-inertial resonance system shown in FIG. 2 to the shaft torsional torque τ _s is shown by the following equations (1) and (2). The theoretical value ω _n0 of the resonance angular frequency is expressed by the following equation (3).

The theoretical value ω _n0 of the resonance angle frequency shown in the equation (3) indicates that the input / output characteristics become ∞ when a signal having an angular frequency equal to the resonance angle frequency is input in the two-inertia resonance system shown in FIG. Shows. However, when the moment of inertia J _E and the coefficient of viscous friction _DE of the engine 2 cannot be ignored, the damping coefficient ξ _n and the time constant α in the equation (2) exist. By comparing the coefficients of the denominator polynomials of equations (1) and (2), the following equations (4), (5) and (6) are obtained.

When the attenuation coefficient ξ _n and the time constant α are calculated from the equations (4), (5) and (6), they are shown by the following equations (7) and (8).

Further, from the above equation, the substantial resonance angular frequency ω _n considering the viscous friction coefficients _DE and _DG is expressed by the following equation (9).

From the above, it can be seen that the resonance angular frequency ω _n and the damping coefficient ξ _n of the bi-inertial resonance system are determined by the moments of inertia J _E , J _G and the viscous friction coefficients _DE , _DG of the engine 2 and the load motor.

As a result of diligent studies, the inventors of the present application have found that the resonance angular frequency ω _n can be controlled by controlling the electric inertia and the electric friction of the load motor in the bi-inertia resonance system shown in FIG. In the following, the control of the resonance angular frequency ω _n by the electric inertia and the electric friction of the load motor will be described.

FIG. 3 is a block diagram in which the electric inertia and the electric friction of the load motor are added to the bi-inertial resonance system shown in FIG. 2. In FIG. 3, _Ja is the electric inertia value of the load motor. _Da is the electric friction value of the load motor. τ _J is the electric inertia torque as the electric inertia. τ _D is the electric friction torque as the electric friction. τ _in is the torque command of the load motor. In the following, it is assumed that the rotation speed ω _G of the load motor does not have a speed offset component.

As shown in FIG. 3, in the bi-inertia resonance system to which the electric inertia and the electric friction of the load motor are added, the electric inertia torque τ _J of the load motor is based on the rotational speed ω _G of the load motor and the electric inertia value J _a . Is calculated. Further, the electric friction torque τ _D of the load motor is calculated based on the rotation speed ω _G of the load motor and the electric friction value D _a . Then, the electric inertia torque τ _J and the electric friction torque τ _D are added to generate the torque command τ _in of the load motor.

In the bi-inertial resonance system shown in FIG. 3, the transfer function from the engine input τ _E to the shaft torsional torque τ _s is expressed by the following equation (10).

When the viscous friction coefficient _DE , DG and the electric friction value D _a are ignored in the equation (10), the resonance angular frequency _ω'no is obtained from the denominator polynomial of the equation (10 ₎ , and the following equation (11) is used. Shown.

Equations (10) and (11) are equivalent to rewriting the moment of _inertia J _G of the equations (1) and (3) to J _G + _Ja and the viscosity coefficient _DG to _DG + Da, respectively. Therefore, it can be seen that the resonance angular frequency ω _no can be controlled using the electric inertia and the electric friction, that is, the electric inertia value J _a and the electric friction value D _a .

The attenuation coefficient _ξ'n in the bi-inertia resonance system to which electric inertia and electric friction are added is shown by the following equation (12), and the substantial resonance angular frequency _ω'n considering the viscous friction coefficients DE and _{DG is} It is represented by the following equation (13).

When the torque command τ _in is controlled in a predetermined sampling period T _s , the electric inertia (electric inertia torque τ _J ) and the electric friction (electric friction torque τ _D ) are sampled in the sampling period T _s . In this case, the electric inertia and the electric friction are discretized. In the following, the effects of discretization of electrical inertia and electrical friction will be described.

From FIG. 3, the torque command τ _in , the electric inertia torque τ _J , and the electric friction torque τ _D of the load motor are represented by the following equations (14), (15), and (16), respectively.

Assuming that the engine input τ _E is a sinusoidal torque of amplitude A τ _E having a constant angular frequency ω as shown in the following equation (17), the rotation speed ω _G of the load motor is the same as the engine input τ _E. It is assumed that the amplitude Aω _G has an angular frequency and the phase lag Ψ is a sinusoidal vibration.

As shown in the equation (15), since the electric inertia torque _τJ includes a differential element (s: Laplace operator), the equations (15) and (16) are expressed in the following equations (19) and (20). Can be rewritten.

From equations (19) and (20), it can be seen that the electric inertia torque τ _J is an operation amount having a lead phase of 90 degrees with respect to the electric friction torque τ _D. The relationship between the electric inertia torque τ _J and the electric friction torque τ _D is shown in FIGS. 4A and 4B. FIG. 4A is a diagram illustrating the relationship between the electric inertia torque τ _J and the electric friction torque τ _D as a vector. FIG. 4B is a diagram describing the relationship between the electric inertia torque τ _J and the electric friction torque τ by the elements of the vectors shown in the equations (19) and (20). In FIGS. 4A and 4B, the angle formed by the electric friction torque τ _D and the torque command τ _in is defined as θ.

When the electric inertia and the electric friction are discretized by the sampling period _Ts , a sample delay (phase delay) occurs due to the sampling period Ts. Assuming that the delay time due to the sampling period T _s is ΔT _s and the engine input τ _E has an angular frequency component as shown in the equation (17), the phase delay φ is expressed by the following equation (21).

In the following, the electric inertia torque, the electric friction torque, and the torque command when the phase delay φ occurs are _τ'J , _τ'D , and _τ'in , respectively. FIG. 5 is obtained by adding the electric inertia torque _τ'J , the electric friction torque _τ'D , and the torque command _τ'in to the relationship between the electric inertia torque τ _J and the electric friction torque τ _D shown in FIG. 4A.

As shown in FIG. 5, the torque command τ'in when the electric inertia and the electric friction are discretized is delayed _in phase by φ with respect to the torque command τ _in . As shown in FIG. 5, the electric inertia torque and the electric friction torque of the torque command _τ'in on the continuous time without discretizing the electric inertia and the electric friction are τ _J2 and τ _D2 .

As shown in equations (19) and (20) and FIG. 4B, the magnitudes of the electric inertia torque τ _J and the electric friction torque τ _D are determined by the electric inertia value J _a and the electric friction value D _a . Let J _a2 be the electric inertia value in the case of the electric inertia torque τ _J2 , and let D a 2 be the electric friction value in the case of the electric friction torque τ _{D 2} . That is, the torque command _τ'in of the load motor when the electric inertia and the electric friction are separated by the sampling period T _s is the torque command based on the electric inertia value J _a2 and the electric friction value D a ₂ when the dispersal is not performed. Is equal to. In the following, τ _J 2 is defined as the true electric inertia torque, and τ _{D 2} is defined as the true electric friction torque. Further, in the following, J _a2 is defined as a true electric inertia value, and D a ₂ is defined as a true electric friction value.

The calculation of the true electric inertia value D _a2 and the true electric friction value J _a2 will be described. First, the electric inertia torque _τ'J and the electric friction torque _τ'D in which the phase delay φ is generated are decomposed into the electric inertia torque without the phase delay and the electric friction torque, respectively. For the electric inertia torque _τ'J and the electric friction torque _τ'D in consideration of the phase lag φ, the phase components without the phase lag are τ _JJ and τ _DD , respectively. Further, the phase lag components of 90 degrees with respect to the phase components τ _JJ and τ _DD without phase lag are τ _JD and τ _DJ , respectively. Then, the electric inertia torque _τ'J and the electric friction torque _τ'D in consideration of the phase delay φ are in the relationship shown in FIGS. 6A and 6B, respectively. FIG. 6A is a diagram showing the relationship between the electric inertia torque _{τ'J and the phase component τ JJ without} phase lag and the phase lag component τ _JD . FIG. 6B is a diagram showing the relationship between the electric friction torque _{τ'D and the phase component τ DD and} the phase lag component τ _DJ without phase lag.

When each vector of FIGS. 6A and 6B is expressed by an equation, the following equations (22) to (27) are obtained.

Note that the electric inertia torque _{τ'J and the electric friction torque τ'D considering the phase delay φ have the same vector magnitude as the electric inertia torque τ J and the electric friction torque τ D without the} phase delay, respectively. do.

FIG. 7 is a diagram showing the relationship between the phase lag φ and each torque shown in the equations (22) to (27) when the electric inertia is positive. When each vector in FIG. 7 is rewritten only by its magnitude, the equations (26) and (27) become the following equations (28) and (29).

Further, the true electric inertia torque τ _J2 and the electric friction torque τ _D2 can also be expressed by the following equations (30) and (31). Further, the electric inertia torque _τ'J , the electric friction torque _τ'D , and the torque command τ'in _in consideration of the phase delay φ have the relations shown in the equations (32) to (34).

Therefore, the following equations (35) and (36) can be obtained by summarizing the electric inertia value J _a2 without the phase delay φ and the electric friction value D a ₂ by the size of the vector.

From the above, dissociating the electric inertia and the electric friction in the sampling period T _s causes the electric inertia value J _a and the electric friction value D _a to become the true electric inertia value J a ₂ and the true electric friction value D a ₂ , respectively. It turns out that it is equivalent to changing. Further, the true electric inertia value J _a2 and the true electric friction value D _a2 have the angular frequency ω of the engine input τ _E , the electric inertia value J _a , the electric friction value D _a , and the delay time ΔT _s due to the sampling period T _s . It can be seen that it can be calculated based on. Note that the direction of the phase delay φ and the direction of θ are opposite.

It is also considered that the equation (11) changes the moment of inertia J _G of the load motor by changing the electric inertia value J _a by introducing the electric inertia. It is also possible to increase the resonance angular frequency ω _n by changing the electric inertia value J _a . Therefore, consider setting the electrical inertia to a negative value. In the following, "setting the electric inertia value Ja to _a positive value" may be referred to as "positive electric inertia". Further, "setting the electric inertia value Ja to _a negative value" may be referred to as "negative electric inertia".

In discrete system control, the phase lag also affects the negative electrical inertia. FIG. 8 is a diagram showing the relationship between the phase lag and each torque when the electric inertia is negative. As shown in FIG. 8, even when the electric inertia is negative, the component τ _JD in the electric friction direction appears as in the case where the electric inertia is positive. However, when the electric inertia is negative, the component _τJD in the electric friction direction acts in the direction opposite to the case where the electric inertia is positive. In FIG. 8, the component _τJD in the electric friction direction is the frictional force in the acceleration direction, which is the frictional force that makes the system unstable. Therefore, in order to cancel this frictional force, it is necessary to correct it by a positive frictional force component. As for the correction amount, if the true electric friction torque τ _D2 is positive in the equation (29), at least the system is not unstable. Since this is equivalent to the case where the true electric friction value D _a2 is positive, the following equation (37) can be obtained from the equation (36).

Equation (37) can be modified as the following equation (38).

Equation (38) is more if the angular frequency ω of the rotation speed ω _G of the load motor is large or the delay time ΔT _s due to the sampling period T _s is large when the electric inertia value J _a is set to a negative value. It shows that the correction by a large electric friction value D _a is required.

Further, the electric inertia value J _a and the electric friction value D _a change to the true electric inertia value J a ₂ and the true electric friction value D a ₂ due to the phase delay φ. Therefore, it is necessary to design so that the true electric inertia value J _a2 and the true electric friction value D _a2 are equal to the original electric inertia value J _a and the electric friction value D _a , respectively. By modifying the equations (35) and (36), the following equations (39) and (40) are obtained.

In equations (39) and (40), when controlling electric inertia and electric friction in a discrete control system, the electric inertia value J _a is smaller than the true electric inertia value J a ₂ in consideration of the phase delay φ. Further, it is shown that the electric friction value D _a needs to be set to a value larger than the true electric friction value D a ₂ .

Based on the above-mentioned study, the inventors of the present application constructed a bi-inertial system including electric inertia and electric friction using the simulation software MATLAB (registered trademark) and performed a simulation. In the simulation, assuming the engine test device 1 shown in FIG. 1, a bi-inertial system was constructed so that the vibration frequency of the torque generated by the engine 2 matches the resonance frequency. The block diagrams of the constructed bi-inertial system are shown in FIGS. 9A and 9B. FIG. 9A is a block diagram showing a configuration of a bi-inertia resonance system in which continuous system control is performed without phase delay of electric inertia and electric friction. Further, FIG. 9B is a block diagram showing a configuration of a bi-inertia resonance system in which discrete system control is performed, in which there is a phase delay in electric inertia and electric friction. In FIGS. 9A and 9B, a high-pass filter is provided on the electric friction side so that the electric friction acts only on the vibration component of the rotation speed ω _G of the load motor. T _d is the time constant of the high-pass filter.

In the bi-inertial resonance system shown in FIGS. 9A and 9B, the sampling frequency T _s is set to 500 μs, the phase delay φ sets the resonance frequency in the bi-inertial resonance system shown in FIG. 2 as ω, and ΔT _s is the sampling frequency T _s . The simulation was performed with a value 1.5 times that of (1.5 T _s ). In addition, J _E = 3.5 × 10 ^-3 kgm ² , J _G = 1.5 × 10 ^-2 kgm ² , _DE = 5.2 × 10 ^-3 Nm (rad / s), _DG = 1. 1 × 10 ⁻² Nm (rad / s), K _s = 800 Nm / rad, and T _d = 0.1 s.

Using one of the combinations of the electric inertia value J _a and the electric friction value D _a (J _a , D _a = (-4.72 × 10 ^-3 , 4.31) considering the phase delay φ of the discrete system control). The relationship from the engine input τ _E to the shaft torsional torque τ _s in the bi-inertia resonance system shown in FIGS. 2 and 9A and 9B is shown by the Bode diagram of FIG. 10. In FIG. 10, the bi-inertia shown in FIG. 2 is shown. The simulation result in the resonance system is shown by a solid line, the simulation result in the bi-inertia resonance system shown in FIG. 9A is shown by a one-point chain line, and the simulation result in the bi-inertia resonance system shown in FIG. 9B is shown by a dotted line.

As shown in FIG. 10, a bi-inertia resonance system (FIG. 9A, 9A, which considers electric inertia and electric friction) as compared with a resonance frequency (about 84 Hz) in a bi-inertia resonance system (FIG. 2) which does not consider electric inertia and electric friction. The resonance frequency (about 87 Hz) in 9B) was shifted to the high frequency side. Further, the gain of the resonance frequency in the bi-inertial resonance system shown in FIGS. 9A and 9B was lower than the gain of the resonance frequency in the bi-inertia resonance system shown in FIG. Therefore, by considering the electric inertia and the electric friction, the vibration due to the resonance was suppressed. In addition, a combination of electric inertia value J _a and electric friction value D _a (J _a , D _a = (-7.50 × 10 ^-3 , 3.00) designed in a continuous system without discrete system control was used. The simulation result in the bi-inertia resonance system shown in FIG. 9A and the simulation result in the bi-inertia resonance system shown in FIG. 9B were in agreement. Therefore, by correcting the electric inertia value J _a and the electric friction value D _a , the discrete system It was found that control similar to continuous system control is possible.

Next, with reference to FIG. 11, the configuration of the control device 10 according to the embodiment of the present invention will be described. In the control device 10 according to the present embodiment, a drive system such as an engine and a load motor that reproduces the load of the drive system, such as the engine test device 1 shown in FIG. 1, are twisted elements via a coupling 4. In the connected system, the torque command τ _in , which indicates the output torque of the load motor, is controlled by the sampling period T _s . That is, the control device 10 according to the present embodiment controls the torque command τ _in by discrete system control.

As shown in FIG. 11, the control device 10 according to the present embodiment includes an electric inertia torque calculation unit 11, an electric friction torque calculation unit 12, a torque command generation unit 13, and a correction unit 14.

The rotation speed ω _G of the load motor is input to the electric inertia torque calculation unit 11. The electric inertia torque calculation unit 11 calculates the electric inertia torque τ _J of the load motor based on the input rotational speed ω _G of the load motor and the electric inertia value J _a of the load motor. Specifically, the electric inertia torque calculation unit 11 calculates the electric inertia torque _τJ based on the above-mentioned equation (15). The electric inertia torque calculation unit 11 outputs the calculated electric inertia torque _τJ to the torque command generation unit 13.

The rotation speed ω _G of the load motor is input to the electric friction torque calculation unit 12. The electric friction torque calculation unit 12 calculates the load motor electric friction torque τ _D based on the input rotation speed ω _G of the load motor and the electric friction value D _a of the load motor. Specifically, the electric friction torque calculation unit 12 calculates the electric friction torque τ _D based on the above-mentioned equation (16). The electric friction torque calculation unit 12 outputs the calculated electric friction torque τ _D to the torque command generation unit 13.

The torque command generation unit 13 generates the torque command τ _in based on the electric inertia torque τ _J calculated by the electric inertia torque calculation unit 11 and the electric friction torque τ _D calculated by the electric friction torque calculation unit 12. Specifically, the torque command generation unit 13 generates the torque command τ _in based on the equation (14).

The correction unit 14 corrects the electric inertia value J _a used by the electric inertia torque calculation unit 11 for the calculation and the electric friction value D _a used by the electric friction torque calculation unit 12 for the calculation based on the sampling period T _s . Specifically, the correction unit 14 corrects the electric inertia value J _a and the electric friction value D _a based on the above-mentioned equations (39) and (40).

As described above, vibration due to resonance can be suppressed by considering the electric inertia and electric friction of the load motor. Therefore, according to the control device 10 according to the present embodiment, vibration due to resonance can be suppressed. Further, by correcting the electric inertia value J _a and the electric friction value D _a based on the sampling period T _s , it is possible to perform the same control as in the case of continuous system control even when the discrete system control is performed. Therefore, according to the control device 10 according to the present embodiment, it is possible to control the load motor in the same manner as in the case of performing continuous system control even when performing discrete system control.

Although the above embodiments have been described as representative examples, it will be apparent to those skilled in the art that many modifications and substitutions are possible within the spirit and scope of the invention. Therefore, the invention should not be construed as limiting by the embodiments described above, and various modifications and modifications can be made without departing from the claims.

1 Engine test device 2 Engine (drive system)
3 Dynamometer 4 Coupling 10 Control device 11 Electric inertia torque calculation unit 12 Electric friction torque calculation unit 13 Torque command generation unit 14 Correction unit

Claims (2)

It is a control device that is connected to the drive system via a twisting element and controls the torque command of the load motor that reproduces the load of the drive system at a predetermined sampling cycle.
An electric inertia torque calculation unit that calculates the electric inertia torque of the load motor based on the rotation speed of the load motor and the electric inertia value of the load motor.
An electric friction torque calculation unit that calculates the electric friction torque of the load motor based on the rotation speed of the load motor and the electric friction value of the load motor.
A torque command generator that generates the torque command based on the electric inertia torque and the electric friction torque,
A control device including a correction unit that corrects the electric inertia value and the electric friction value based on the sampling period. In the control device according to claim 1,
The electric inertia value and the electric friction value are set to J _a and D _a , respectively, and the electric inertia value and the electric friction value changed by the delay time due to the sampling cycle are set to J a ₂ and D a ₂ , respectively, and the input torque of the drive system is set. Let ω be the angular frequency, T _s be the sampling period, and ΔT _s be the delay time due to the sampling period T _s .
The correction unit corrects the electric inertia value J _a and the electric friction value D _a so as to satisfy the following equations.

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