JP7426319B2 - Control device - Google Patents

Description

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

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

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

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

As shown in FIG. 1, the engine testing device 1 includes an engine 2, a dynamometer 3, a shaft, and a coupling 4 that connects them. The engine testing apparatus 1 shown in FIG. 1 can be modeled as a two-inertial resonant system using the torsion element included in the coupling 4. FIG. 2 is a block diagram in which the engine testing apparatus 1 shown in FIG. 1 is modeled as a two-inertia resonance system.

In FIG. 2, J _E is the moment of inertia of the engine 2. D _E is the viscous friction coefficient of the engine 2. J _G is the moment of inertia of a load motor included in the dynamometer 3 that reproduces the load of the engine 2 . D _G is the viscous friction coefficient of the load motor. K _s is the spring constant of coupling 4. τ _E is the torque of engine 2 (engine input). 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 rotational speed of the engine 2. ω _G is the rotation speed of the load motor. s is a Laplace operator.

When the engine testing apparatus 1 is defined as a two-inertia resonant system _shown in _FIG . It has a resonant frequency according to _s . When the vibration frequency of the torque generated by the engine 2 and the resonance frequency match, the torque vibrations applied to the coupling 4 and the shaft are amplified, resulting in a large load being applied. Therefore, it is required to perform resonance suppression control so that torque vibration is not amplified at the resonance frequency.

Conventional techniques for resonance suppression control in a two-inertial resonant system include control design using Manabe polynomials, resonance ratio control, H∞ control, and the like. However, in the engine testing apparatus 1 shown in FIG. 1, the torque generated by the engine 2 and the rotational speed of the engine 2 cannot be obtained, and only the torque command of the load motor can be operated. Therefore, conventional techniques such as those described above cannot be used.

SUMMARY OF THE INVENTION An object of the present invention, which has been made in view of the above problems, is to provide a control device that can suppress vibrations caused by resonance between a drive system and a load motor.

In order to solve the above problems, a control device according to the present invention is a control device that is connected to a drive system via a torsion element and controls a torque command of a load motor that reproduces the load of the drive system at a predetermined sampling period. an electric inertia torque calculation unit that calculates an electric inertia torque of the load motor based on a rotational speed of the load motor and an 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 electric friction value of the electric friction value; 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 that corrects 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 respectively J _a and D _a , and the electric inertia value and the electric friction value that change depending on the delay time due to the sampling period are J _a2 and D a, respectively. _a2 , the angular frequency of the input torque of the drive system is ω, the sampling period is T _s , and the delay time due to the sampling period T _s is 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 equation.

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.

1 is a diagram illustrating a configuration example of an engine testing device. FIG. 2 is a block diagram showing the configuration of a two-inertial resonance system that models the engine testing apparatus shown in FIG. 1. FIG. 3 is a block diagram in which the electric inertia and electric friction of a load motor are added to the two-inertia resonance system shown in FIG. 2. FIG. FIG. 3 is a diagram describing the relationship between electric inertia torque and electric friction torque using vectors when electric inertia and electric friction are discretized. FIG. 3 is a diagram describing the relationship between electric inertia torque and electric friction torque using vector elements when electric inertia and electric friction are discretized. FIG. 3 is a diagram showing the relationship between phase delay and torque command of a load motor. FIG. 3 is a diagram showing the relationship between electric inertia torque τ′ _J , phase component τ _JJ without phase lag, and phase lag component τ _JD . FIG. 3 is a diagram showing the relationship between electric friction torque τ′ _D , phase component τ _DD without phase lag, and phase lag component τ _DJ . FIG. 3 is a diagram showing the relationship between phase delay and each torque in positive electric inertia. FIG. 3 is a diagram showing the relationship between phase delay and each torque in negative electric inertia. FIG. 2 is a block diagram showing the configuration of a two-inertial resonant system in which continuous system control is performed. FIG. 2 is a block diagram showing the configuration of a two-inertial resonant system in which discrete system control is performed. FIG. 9 is a Bode diagram showing the relationship from engine input to shaft torsion torque in the two-inertia system shown in FIGS. 2, 9A, and 9B. FIG. 1 is a diagram illustrating a configuration example of a control device according to an embodiment of the present invention.

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

A control device according to an embodiment of the present invention is provided in an engine testing device 1 shown in FIG. This suppresses vibrations caused by resonance between the engine 2 and the load motor. First, resonance suppression control in a two-inertial resonance system shown in FIG. 2, which models the engine testing apparatus 1, will be described.

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

The theoretical value ω _n0 of the resonant angular frequency shown in equation (3) indicates that in the two-inertial resonant system shown in Fig. 2, when a signal with an angular frequency equal to the resonant angular frequency is input, the input/output characteristic becomes ∞. It shows. However, if the moment of inertia J _E and the viscous friction coefficient D _E of the engine 2 cannot be ignored, the damping coefficient ξ _n and the time constant α in equation (2) exist. By comparing the coefficients of the denominator polynomials in equations (1) and (2), the following equations (4), (5), and (6) are obtained.

From equations (4), (5), and (6), the damping coefficient ξ _n and time constant α are calculated as shown in equations (7) and (8) below.

Further, from the above-mentioned formula, the substantial resonance angular frequency ω _n in consideration of the viscous friction coefficients D _E and D _G is expressed by the following formula (9).

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

As a result of extensive studies, the inventors of the present application found that in the two-inertial resonant system shown in FIG. 2, the resonant angular frequency ω _n can be controlled by controlling the electrical inertia and electrical friction of the load motor. In the following, control of the resonance angular frequency ω _n using the electric inertia and electric friction of the load motor will be explained.

FIG. 3 is a block diagram in which electric inertia and electric friction of a load motor are added to the two-inertia resonance system shown in FIG. 2. In FIG. 3, J _a is the electrical inertia value of the load motor. D _a is the electric friction value of the load motor. τ _J is the electric inertia torque as electric inertia. τ _D is the electric friction torque as electric friction. τ _in is the torque command of the load motor. Note that in the following, it is assumed that the rotational speed ω _G of the load motor does not have a speed offset component.

As shown in Fig. 3, in a two-inertia resonant system in which the electric inertia of the load motor and electric friction are added, the electric inertia torque τ J of the load motor is determined 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 rotational 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 two-inertial resonant system shown in FIG. 3, the transfer function from the engine input τ _E to the shaft torsion torque τ _s is expressed by the following equation (10).

In equation (10), if the viscous friction coefficients D _E , D _G and electric friction value D _a are ignored, the resonance angular frequency ω' _no is determined from the denominator polynomial in equation (10), and the following equation (11) is obtained. shown.

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

The damping coefficient ξ' _n in a two-inertial resonant system with electric inertia and electric friction added is expressed by the following equation (12), and the effective resonance angular frequency ω' _n considering the viscous friction coefficients D _E and D _G is It is represented by the following equation (13).

When controlling the torque command τ _in at a predetermined sampling period T _s , electric inertia (electric inertia torque τ _J ) and electric friction (electric friction torque τ _D ) are sampled at the sampling period T _s . In this case, electrical inertia and electrical friction are discretized. In the following, the influence of discretization of electric inertia and electric friction will be explained.

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

Assuming that the engine input τ _E is a sinusoidal torque of amplitude Aτ _E with a constant angular frequency ω, as shown in equation (17) below, the rotational speed ω _G of the load motor is the same as the engine input τ _E. Assume that it is a sinusoidal oscillation with an angular frequency, an amplitude Aω _G and a phase lag Ψ.

As shown in equation (15), electric inertia torque τ _J includes a differential element (s: Laplace operator), so equations (15) and (16) can be transformed into equations (19) and (20) below. Can be rewritten.

From equations (19) and (20), it can be seen that the electric inertia torque τ _J is a manipulated variable having a phase lead of 90 degrees with respect to the electric friction torque τ _D. The relationship between electric inertia torque τ _J and electric friction torque τ _D is shown in FIGS. 4A and 4B. FIG. 4A is a diagram describing the relationship between electric inertia torque τ _J and electric friction torque τ _D using vectors. FIG. 4B is a diagram describing the relationship between electric inertia torque τ _J and electric friction torque τ using vector elements shown in 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 electric inertia and electric friction are discretized at 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 that the engine input τ _E has an angular frequency component as shown in equation (17), the phase delay φ is expressed by the following equation (21).

In the following, the electric inertia torque, electric friction torque, and torque command when the phase delay φ occurs are respectively τ' _J , τ' _D , and τ' _in . When electric inertia torque τ' _J , electric friction torque τ' _{D and} torque command τ' _in are added to the relationship between electric inertia torque τ J and electric friction torque τ _D shown in FIG. 4A, the result is as shown in FIG. 5.

As shown in FIG. 5, the torque command τ' _in when electric inertia and 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 electric friction torque of the torque command τ′ _in on a continuous time basis in which the electric inertia and electric friction are not discretized are assumed to be τ _J2 and τ _D2 .

As shown in equations (19), (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 . The electric inertia value when the electric inertia torque τ _J2 is set to J _a2 , and the electric friction value when the electric friction torque τ _D2 is set to D _a2 . In other words, the torque command τ' _in of the load motor when the electric inertia and electric friction are discretized at the sampling period T _s is the torque command based on the electric inertia value J _a2 and the electric friction value D _a2 when discretization is not performed. is equal to In the following, τ _J2 is defined as the true electric inertia torque, and τ _D2 is defined as the true electric friction torque. Moreover, below, J _a2 is defined as a true electric inertia value, and D _a2 is defined as a true electric friction value.

Calculation of the true electric inertia value D _a2 and the true electric friction value J _a2 will be explained. First, the electric inertia torque τ' _J and the electric friction torque τ' _D with a phase lag φ are respectively decomposed into an electric inertia torque without a phase lag and an electric friction torque. For the electric inertia torque τ' _J and the electric friction torque τ' _D considering the phase lag φ, let the phase components without phase lag be τ _JJ and τ _DD , respectively. Furthermore, let τ _JD and τ _DJ be phase lag components of 90 degrees with respect to the phase components τ _JJ and τ _DD with no phase lag, respectively. Then, the electric inertia torque τ' _J and the electric friction torque τ' _D in consideration of the phase delay φ have the relationships shown in FIGS. 6A and 6B, respectively. FIG. 6A is a diagram showing the relationship between electric inertia torque τ′ _J , phase component τ _JJ without phase lag, and 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 without phase lag and the phase lag component τ _DJ .

Expressing each vector in FIGS. 6A and 6B using equations, 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 lag φ have the same vector magnitude as the electric inertia torque τ _J and the electric friction torque τ _D without phase lag, respectively. do.

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

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 relationships shown in equations (32) to (34).

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

From the above, discretizing the electric inertia and electric friction at the sampling period T _s means converting the electric inertia value J _a and the electric friction value D _a into the true electric inertia value J _a2 and the true electric friction value D _a2 , respectively. It turns out that it is equivalent to changing. In addition, the true electric inertia value J _a2 and the true electric friction value D _a2 depend on 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 calculations can be made based on Note that the direction of the phase delay φ and the direction of θ are opposite.

By introducing electric inertia, equation (11) can also be considered to change the moment of inertia J _G of the load motor by changing the electric inertia value J _a . It is also possible to increase the resonance angular frequency ω _n by changing the electrical inertia value J _a . Therefore, consider setting the electrical inertia to a negative value. Hereinafter, "setting the electrical inertia value J _a to a positive value" may be referred to as "positive electrical inertia." Furthermore, "setting the electrical inertia value J _a to a negative value" is sometimes referred to as "negative electrical inertia."

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

Equation (37) can be transformed as shown in Equation (38) below.

Equation (38 _{) is expressed as} This indicates that correction using a large electric friction value D _a is required.

Moreover, the electric inertia value J _a and the electric friction value D _a change to the true electric inertia value J _a2 and the true electric friction value D _a2 due to the phase delay φ. Therefore, it is necessary to design the true electric inertia value J _a2 and the true electric friction value D _a2 to be equivalent to the original electric inertia value J _a and electric friction value D _a , respectively. By transforming equations (35) and (36), the following equations (39) and (40) are obtained.

Equations (39) and (40) indicate that 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 _a2 in consideration of the phase delay φ, It also indicates that the electric friction value D _a needs to be set to a larger value than the true electric friction value D _a2 .

Based on the above-mentioned study, the inventors of the present application constructed a two-inertia system including electric inertia and electric friction and conducted a simulation using the simulation software MATLAB (registered trademark). The simulation assumed the engine testing apparatus 1 shown in FIG. 1, and constructed a two-inertia system so that the vibration frequency of the torque generated by the engine 2 coincided with the resonance frequency. Block diagrams of the constructed two-inertial system are shown in FIGS. 9A and 9B. FIG. 9A is a block diagram showing the configuration of a two-inertia resonant 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 the configuration of a two-inertial resonant system in which discrete system control is performed, in which there is a phase delay of electrical inertia and electrical friction. In addition, 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 rotational speed ω _G of the load motor. T _d is the time constant of the high-pass filter.

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

Using one of the combinations of electric inertia value J _a and electric friction value D _a (J _a , D _a = (-4.72×10 ^-3 , 4.31) considering the phase delay φ of discrete system control The Bode diagram in FIG. 10 shows the relationship between the engine input τ _E and the shaft torsion torque τ _s in the two-inertia resonance system shown in FIG. 2 and FIGS. 9A and 9B. The simulation results for the resonant system are shown by a solid line, the simulation results for the two-inertial resonant system shown in FIG. 9A are shown by a dashed-dotted line, and the simulation results for the two-inertial resonant system shown in FIG. 9B are shown by a dotted line.

As shown in FIG. 10, compared to the resonant frequency (approximately 84 Hz) in the two-inertial resonant system (FIG. 2) that does not take into account electric inertia and electric friction, the two-inertial resonant system that takes into account electric inertia and electric friction (FIG. 9A, The resonant frequency (approximately 87 Hz) in 9B) was shifted to the higher frequency side. Moreover, compared to the gain of the resonant frequency in the two-inertial resonant system shown in FIG. 2, the gain of the resonant frequency in the two-inertial resonant system shown in FIGS. 9A and 9B was reduced. Therefore, by considering electric inertia and electric friction, vibrations due to resonance were suppressed. In addition, the combination of electric inertia value J _a and electric friction value D _a designed in a continuous system without discrete system control (J _a , D _a = (-7.50×10 ^-3 , 3.00) was used. The simulation results for the two-inertial resonant system shown in Fig. 9A and the simulation results for the two-inertial resonant system shown in Fig. 9B coincided. Therefore, by correcting the electric inertia value J _a and the electric friction value D _a , It was also 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 an embodiment of the present invention will be described. The control device 10 according to the present embodiment, like the engine testing device 1 shown in FIG. In the connected system, a torque command _τ in indicating the output torque of the load motor is controlled at a 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 section 11, an electric friction torque calculation section 12, a torque command generation section 13, and a correction section 14.

The electrical inertia torque calculation unit 11 receives the rotational speed ω _G of the load motor. 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 τ _J based on the above-mentioned equation (15). The electric inertia torque calculation unit 11 outputs the calculated electric inertia torque τ _J to the torque command generation unit 13.

The electric friction torque calculation unit 12 receives the rotational speed ω _G of the load motor. The electric friction torque calculation unit 12 calculates the load motor electric friction torque τ _D based on the input rotational 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 a 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 equation (14).

The correction unit 14 corrects the electric inertia value J _a used in the calculation by the electric inertia torque calculation unit 11 and the electric friction value _{D a} used in the calculation by the electric friction torque calculation unit 12 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 electrical inertia and electrical 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 , even when performing discrete system control, the same control as when performing continuous system control can be performed. Therefore, according to the control device 10 according to the present embodiment, even when performing discrete system control, it is possible to control the load motor in the same manner as when performing continuous system control.

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

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

Claims (2)

A control device that is connected to a drive system via a torsion element and controls a torque command of a load motor that reproduces the load of the drive system at a predetermined sampling period,
an electric inertia torque calculation unit that calculates the electric inertia torque of the load motor based on the rotational speed of the load motor and the electric inertia value of the load motor;
an electric friction torque calculation unit that calculates electric friction torque of the load motor based on a rotational speed of the load motor and an electric friction value of the load motor;
a torque command generation unit that generates the torque command based on the electric inertia torque and the electric friction torque;
A control device comprising: a correction unit that corrects the electric inertia value and the electric friction value based on the sampling period. The control device according to claim 1,
The electric inertia value and the electric friction value are respectively J _a and D _a , the electric inertia value and the electric friction value that change depending on the delay time due to the sampling period are J _a2 and D _a2 respectively, and the input torque of the drive system is If the angular frequency is ω, the sampling period is T _s , and the delay time due to the sampling period T _s is ΔT _s , then
The correction unit is a control device that corrects the electric inertia value J _a and the electric friction value D _a so as to satisfy the following equation.

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