JP2013257857A - Controller of mechanical device, mechanical system, and method of controlling mechanical device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress self-excited oscillation of a mechanical device caused by presence of a dead band between a drive shaft and a load shaft, in a simple configuration and reliable manner.SOLUTION: A controller 14 controls a mechanical device 12 in which a drive shaft and a load shaft engage with each other, and force of the drive shaft is transmitted to the load shaft, as a control object. The controller 14 calculates a phase difference between the drive shaft and load shaft through a phase difference calculator 62, calculates a compensation value for compensating for the calculated phase difference by using a proportional element and a differentiation element through a compensator 64, and performs subtraction of a torque command value τfor a motor by using the compensation value through a compensation value subtractor 66.

Description

  The present invention relates to a control device for a mechanical device, a mechanical system, and a control method for the mechanical device.

As shown in FIG. 14, the multi-inertia torsional mechanical device in which the drive shaft and the load shaft are engaged with each other by a gear or the like and the force of the drive shaft is transmitted to the load shaft. That is, there may be a dead zone (also referred to as mechanical backlash or backlash) between the gears 204 and 206 coupled to the drive shaft 200 and the load shaft 202. This dead zone discontinuously changes the transmission of force from the drive shaft to the load shaft.
For this reason, the presence of a dead zone may cause self-excited vibration in which the controlled load vibrates intermittently. As shown in FIG. 15, the self-excited vibration hardly occurs when the load is operated slowly, and easily occurs when the load is operated quickly.

  The dead zone is a non-linear element. For example, Patent Document 1 calculates the equivalent inertia obtained by adding the backlash portion inertia to the motor inertia, calculates the estimated speed from the torque command and the equivalent inertia, integrates the deviation between the motor speed and the estimated speed, and gives it to the linear element. A backlash vibration suppressing method is disclosed in which a backlash estimated torque is calculated by multiplying a linear element output by an equivalent spring constant, and the backlash estimated torque is feedback-compensated before the torque command is obtained. The backlash vibration suppressing method described in Patent Document 1 is a method of handling a nonlinear element of backlash (dead zone) by continuously linearizing it.

Japanese Patent No. 3580632

Here, modeling the dead zone, which is a non-linear element, causes complexity of the algorithm. In addition, the use of nonlinear elements for control leads to vulnerability to control modeling errors.
Furthermore, although it is not the control which suppresses self-excited vibration actively, there is also control which is suppressing only the self-excited vibration as a result, and the suppression effect of self-excited vibration is not compensated theoretically.
As described above, there is no established control that can suppress self-excited vibration caused by the presence of the dead body while having a simple design, simple adjustment, high robustness, and the like.

  The present invention has been made in view of such circumstances, and the self-excited vibration of the mechanical device due to the presence of a dead zone between the drive shaft and the load shaft can be more reliably detected with a simple configuration. An object of the present invention is to provide a control device for a mechanical device, a mechanical system, and a control method for the mechanical device that can be suppressed.

  In order to solve the above-mentioned problems, the control device, the mechanical system, and the control method of the mechanical device of the present invention employ the following means.

  The control device for a mechanical device according to the first aspect of the present invention is a control device that controls a mechanical device that engages a drive shaft and a load shaft and transmits the force of the drive shaft to the load shaft, Compensation for compensating for the deviation calculated by the position deviation calculating means using at least one of a proportional element and a differential element, and a position deviation calculating means for calculating a deviation between the position of the drive shaft and the position of the load shaft. Compensating means for calculating a value, and subtracting means for subtracting a control value for the drive shaft by the compensation value calculated by the compensating means.

  According to this configuration, the control device targets a mechanical device that engages the drive shaft and the load shaft and transmits the force of the drive shaft to the load shaft. The engagement of the drive shaft and the load shaft means that gears are provided on the drive shaft and the load shaft, for example, and the gears are combined to transmit force. The driving means for driving the driving shaft is, for example, a motor.

  In such a mechanical device, there is a dead zone where the drive shaft and the load shaft do not mesh with each other, and self-excited vibration occurs due to the influence of the dead zone. Since the dead zone element is a nonlinear component, it is difficult to accurately model the dead zone element.

  In this configuration, the deviation between the position of the drive shaft and the position of the load shaft is calculated by the position deviation calculation means. Note that the deviation is a phase difference between the rotating drive shaft and the load shaft.

Then, a compensation value for compensating for the deviation calculated by the position deviation calculating means using at least one of the proportional element and the differential element is calculated by the compensating means, and the control value for the drive shaft is subtracted by the compensation value by the subtracting means. Is done. The proportional element in the compensation means is a virtual spring element, and the differential element is a virtual damper element. Both the spring element and the damper element have a function of suppressing vibration.
For this reason, the compensation value calculated by the compensation means has the effect of equivalently increasing the torsional rigidity when viewed from the drive shaft side, and suppresses self-excited vibration.

  Therefore, this configuration can reduce the self-excited vibration of the mechanical device due to the presence of the dead zone between the drive shaft and the load shaft without modeling the dead zone element, which is a non-linear component, with a simple configuration. It can be reliably suppressed.

  In the first aspect, it is preferable that the control value is a torque command value for the drive shaft.

  According to this configuration, since the torque command value for the drive shaft is fed back as a compensation value, self-excited vibration can be suppressed with a simple configuration.

  In the first aspect, it is preferable that the control value is a speed command value for the drive shaft, and includes a differentiation unit that differentiates the deviation calculated by the position deviation calculation unit and outputs the differentiated value to the compensation unit. .

  According to this configuration, since the speed command value for the drive shaft is fed back as a compensation value, self-excited vibration can be suppressed with a simple configuration.

  In the first aspect, it is preferable that the value of the proportional element is set in advance so as to be equal to the value of the spring stiffness of the mechanical device.

  According to this configuration, the force of the virtual spring is smoothly connected to the reaction force of the shaft torsion torque received by the drive shaft, and self-excited vibration can be further suppressed.

  In the first aspect, it is preferable that the value of the differential element is set in advance so as to be a damping rate larger than a damping rate of vibration in the mechanical device.

  According to this configuration, the vibration suppressing effect is further increased, and the self-excited vibration can be further suppressed.

  In the first aspect, the vehicle includes an estimation unit that estimates a position of the load shaft based on a state quantity of the drive shaft, and the position deviation calculation unit is configured to estimate the position of the drive shaft and the load estimated by the estimation unit. Calculate the deviation from the axis position.

  According to this configuration, the position of the load shaft is estimated by the estimation unit based on the state quantity of the drive shaft. The state quantity of the drive shaft is the speed of the drive shaft, the position of the drive shaft, or a torque command value for driving the drive shaft. Then, the deviation between the position of the drive shaft and the estimated position of the load shaft is calculated by the position deviation calculating means. As described above, this configuration can calculate the deviation by estimating the position of the load shaft even in a device configuration in which it is difficult to detect the position of the load shaft.

  Therefore, this configuration has a dead zone between the drive shaft and the load shaft without modeling the dead zone element that is a non-linear component, even if the device configuration cannot detect the state quantity of the load shaft. The self-excited vibration of the mechanical device resulting from the above can be more reliably suppressed with a simple configuration.

  The mechanical system according to the second aspect of the present invention includes a mechanical device that engages a drive shaft and a load shaft, and transmits the force of the drive shaft to the load shaft, and the control described above that controls the mechanical device. An apparatus.

  A control method for a mechanical device according to a third aspect of the present invention is a control method for controlling a mechanical device that engages a drive shaft and a load shaft and transmits the force of the drive shaft to the load shaft. A first step of calculating a deviation between the position of the drive shaft and the position of the load shaft; a second step of calculating a compensation value for compensating the calculated deviation using at least one of a proportional element and a differential element; And a third step of subtracting the control value for the drive shaft by the calculated compensation value.

  According to the present invention, there is an excellent effect that the self-excited vibration of the mechanical device caused by the dead zone existing between the drive shaft and the load shaft can be more reliably suppressed with a simple configuration.

1 is a schematic configuration diagram of a mechanical system according to a first embodiment of the present invention. It is a graph which shows the example of the time change of the load angle command value, the detected value of a load angle, and the detected motor angle value when self-excited vibration occurs due to the presence of a dead zone. It is a functional block diagram which shows the structure of the mechanical system which concerns on 1st Embodiment of this invention. It is a schematic diagram which shows the effect | action of the compensator which concerns on 1st Embodiment of this invention. It is a block diagram of a backlash-less gear. It is a functional block diagram which shows the structure of the mechanical system using a backlash-less gear. It is a figure which shows the rigidity of the mechanical apparatus using a backlash-less gear. It is a figure which shows the rigidity of the mechanical apparatus which concerns on this 1st Embodiment. It is a graph which shows the time change of the load angle command value which concerns on 1st Embodiment of this invention, a load angle, and a motor angle. It is a graph which shows the state of the self-excited vibration at the time of changing the virtual spring coefficient which concerns on 1st Embodiment of this invention. It is a functional block diagram which shows the structure of the mechanical system which concerns on 2nd Embodiment of this invention. It is a functional block diagram which shows the structure of the mechanical system which concerns on 3rd Embodiment of this invention. It is a functional block diagram which shows the structure of the load position estimation part which concerns on 3rd Embodiment of this invention. It is a schematic diagram which shows the dead zone which exists between a drive shaft and a load shaft. It is a graph which shows the time change of the self-excited vibration by presence of a dead zone.

  Hereinafter, an embodiment of a control device, a mechanical system, and a control method for a mechanical device according to the present invention will be described with reference to the drawings.

[First Embodiment]
The first embodiment of the present invention will be described below.

FIG. 1 is a schematic configuration diagram of a mechanical system 10 according to the first embodiment.
The mechanical system 10 includes a mechanical device 12 and a control device 14 that controls the mechanical device 12.

  In the mechanical device 12, the drive shaft 20 and the load shaft 22 mesh with each other, and the force of the drive shaft 20 is transmitted to the load shaft 22. The drive shaft 20 and the load shaft 22 mesh with each other. As shown in FIG. 1, the drive shaft 20 and the load shaft 22 are provided with gears 24 and 26, respectively, and the gears 24 and 26 are combined to transmit force. Is Rukoto. For example, the drive shaft 20 is a rotating shaft (so-called motor shaft) of the motor 28, and the load shaft 22 is a rotating shaft of the load 30. As described above, the mechanical device 12 is a two-inertia system including the motor 28 having the drive shaft 20 and the load 30 having the load shaft 22.

  The control device 14 includes a position controller 32 and a speed controller 34, and is a fully closed control system that controls the position of the load shaft 22 and the speed of the motor 28 (drive shaft 20).

The position controller 32 includes a load angle command value θ _L ^* indicating the position of the load shaft 22 (which is an angle in the first embodiment, and hereinafter referred to as “load angle θ _L ”), and a detected value of the load angle θ _L. Motor speed command value ω _M ^* indicating the speed of the drive shaft 20 (angular speed in the first embodiment) is calculated based on the deviation (hereinafter referred to as “load angle deviation”) δθ _L. Note that the load angle deviation δθ _L is calculated by the subtractor 36. Further, the position controller 32 calculates the motor speed command value ω _M ^* by proportional control using the load angle deviation δθ _L as an example.

The speed controller 34 determines the deviation between the motor speed command value ω _M ^* calculated by the position controller 32 and the detected value of the motor speed ω _M that is the angular speed of the drive shaft 20 (motor 28) (hereinafter referred to as “motor speed deviation”). The torque command value τ _M ^* indicating the torque of the drive shaft 20 is calculated based on δω _M. The motor speed deviation [delta] [omega _M is calculated by the subtractor 38. Moreover, the speed controller 34 calculates the torque command value τ _M ^* by proportional-integral control using the motor speed deviation δω _M as an example.

The motor 28 drives the drive shaft 20 based on the torque command value τ _M ^* calculated by the speed controller 34.

  The control device 14 includes, for example, a CPU (Central Processing Unit), a RAM (Random Access Memory), and a computer-readable recording medium. A series of processes for realizing various functions of the control device 14 is recorded on a recording medium or the like in the form of a program as an example, and the CPU reads the program into a RAM or the like to process and calculate information. Various functions are realized by executing the processing.

  Here, between the drive shaft 20 and the load shaft 22, there is a dead zone (also referred to as mechanical backlash or backlash) between the gears 24 and 26 coupled to the drive shaft 20 and the load shaft 22. . This dead zone discontinuously changes the transmission of force from the drive shaft to the load shaft.

FIG. 2 is a graph showing an example of a time change of the load angle command value θ _L ^* , the load angle θ _L , and the motor angle θ _M that is the rotation angle of the drive shaft 20 (motor 28) in self-excited vibration due to the presence of the dead zone. It is. FIG. 2 shows the result of simulation.
As shown in FIG. 2, in the conventional control, even if the load angle command value θ _L ^* is constant, the load angle θ _L and the motor angle θ _M vibrate due to the influence of the dead zone.

  FIG. 3 is a functional block diagram showing the configuration of the mechanical system 10 according to the first embodiment. In FIG. 3, the mechanical device 12 is represented by modeling for each element of the transfer function.

Drive shaft 20 (motor 28), the motor inertia element 40, integral element 42, and the motor viscous resistance element 44 acts to rotate at a motor speed omega _M. Then, the motor speed omega _M is integrated (integral element 46), the motor angle theta _M is obtained. The motor angle theta _M is differentiated (differentiation element 60), the detection value of the motor speed omega _M is obtained.

The force output from the drive shaft 20 is transmitted to the load shaft 22 under the influence of the dead zone Nx (dead zone element 48) and the torsional rigidity K _R (torsional rigidity element 50).

The load shaft 22 rotates at the load angular velocity ω _L by the action of the load inertia element 52, the integration element 54, and the load viscous resistance element 56. Then, the load angular velocity omega _L is integrated (integral element 58), the angle of the drive shaft 20 load angle theta _L is obtained.

  Here, the dead zone element 48 is difficult to accurately model because it cannot accurately grasp the width of the dead body. Note that the dead zone element 48 is a non-linear component.

  Therefore, the control device 14 according to the first embodiment includes a phase difference calculator 62, a compensator 64, and a compensation value subtractor 66.

The phase difference calculator 62 calculates the deviation between the position of the drive shaft 20 and the position of the load shaft 22. In the rotating drive shaft 20 and load shaft 22, the deviation is the phase difference δθ _ML between the motor angle θ _M and the load angle θ _L. When there is no dead zone in the mechanical device 12, the phase difference δθ _ML does not occur, so the phase difference δθ _ML represents the influence of the dead zone.

The compensator 64 calculates a compensation value X that compensates the phase difference δθ _ML calculated by the phase difference calculator 62 using a proportional element and a differential element.

  The compensation value subtractor 66 subtracts the control value for the motor 28 by the compensation value X calculated by the compensator 64.

The control value for the motor 28 according to the first embodiment is the torque command value τ _M ^* output from the speed controller 34, and the motor 28 is subtracted by the compensation value X by the compensation value subtractor 66. Torque command value τ _M ^* is input. That is, the torque command value τ _M ^* is fed back with the compensation value X.

  In the compensator 64 according to the first embodiment, the proportional element is a virtual spring element, and the differential element is a virtual damper element. Both the spring element and the damper element have a function of suppressing vibration.

The compensator 64 according to the first embodiment calculates the compensation coefficient C _ML by the sum of the proportional term having the virtual spring coefficient K _C and the differential term having the virtual damper coefficient T _C as expressed in the following equation (1). calculate. In the following formula (1), s is a Laplace operator.
_{C ML (s) = K C} × (1 + T C × s) ··· (1)

Then, the compensation value X is calculated by multiplying the phase difference δθ _ML by the compensation coefficient C _ML .

Thus, the compensator 64 is input a phase difference .delta..theta _ML, since the compensation value X to be output corresponding to the torque, the torsional rigidity K _R O relationship between 2 inertia of the drive shaft 20 and the load shaft 22 Is equivalent to For this reason, the compensation value calculated by the compensator 64 has the effect of increasing the torsional rigidity equivalently when viewed from the drive shaft 20 side, and suppresses self-excited vibration.

FIG. 4 is a schematic diagram illustrating the operation of the compensator 64. The compensator 64 virtually includes a spring and a dead band that is opposite to the rotation direction of the gear 24 of the drive shaft 20 as shown in FIG. This corresponds to the provision of a damper.
As a result, the teeth of the gear 24 of the drive shaft 20 are located in the dead zone opposite to the rotation direction, and contact with the teeth of the gear 26 of the load shaft 22 in the dead zone is suppressed, and the dead zone exists. The self-excited vibration of the mechanical device 12 is more reliably suppressed with a simple configuration.

  As described above, the compensator 64 is provided in consideration of the dynamics between the two inertias of the drive shaft 20 and the load shaft 22. Here, a mechanical system using a conventionally known backlashless gear in which a torsion spring is mechanically provided between two inertias and a dead zone element is removed will be compared with the mechanical system 10 according to the first embodiment. To do.

FIG. 5 is an example of a configuration diagram of the backlash-less gear 70.
As shown in FIG. 5, the backlash-less gear 70 includes a gear 72 provided on the drive shaft 20 and a gear 74 provided on the load shaft 22 coupled via a torsion spring 76. The gear 72 is twisted and meshed with the gear 74 of the load shaft 22. As a result, the restoring force of the torsion spring 76 stably presses the teeth of the gear 72 in one direction of the gear 74 so that no dead zone is generated.

FIG. 6 is a functional block diagram showing the configuration of the mechanical device 82 and the control device 84 of the mechanical system 80 using the backlash-less gear 70. In FIG. 6, the same components as those in FIG. 3 are denoted by the same reference numerals as those in FIG.
Since the mechanical system 80 uses the backlash-less gear 70, the dead zone element 48 does not exist as described above. Instead, there is a backlashless gear element 86. As shown by the force flow indicated by the double line in FIG. 6, the backlashless gear 70 gives torque (action and reaction) to both the drive side and the load side.

On the other hand, since the mechanical system 10 according to the first embodiment subtracts the torque command value τ _M ^* by the compensation value X as described with reference to FIG. 3, the torque by compensation is given only to the drive side.

FIG. 7 is a diagram showing the rigidity of the mechanical device 82 using the backlash-less gear 70 with the vertical axis representing torque and the horizontal axis representing a torsion angle (difference between input angle and output angle).
As shown in FIG. 7, the rigidity (torsion spring rigidity) of the torsion spring 76 is weaker than the rigidity (axial rigidity) of the drive shaft 20 and the load shaft 22 indicated by solid lines.

On the other hand, FIG. 8 is a diagram showing the rigidity of the mechanical device 12 according to the first embodiment.
A one-dot chain line in FIG. 8 indicates rigidity when the compensator 64 does not act. When the torsion angle is in the dead body, no force is transmitted, and as a result, the shaft rigidity is discontinuous.

On the other hand, the solid line in FIG. 8 shows the rigidity when ideal compensation is performed by the compensator 64, and the force of the virtual spring element is smoothly applied to the reaction force of the shaft torsion torque received by the drive shaft 20. As a result, the rigidity of the mechanical device 12 does not change.
In order to perform such ideal compensation, the value of the virtual spring coefficient K _C is set in advance so as to be equal to the value of the torsional rigidity K _R (spring rigidity) of the mechanical device 12.

The value of the virtual damper coefficient T _C is preferably than the damping rate of vibration in machinery 12 is preset so that a large damping factor. For example, the attenuation rate is set to be 0.1 or more. Thereby, the suppression effect of a vibration becomes larger and it becomes possible to suppress self-excited vibration more.

FIG. 9 is a graph showing an example of a time change of the load angle command value θ _L ^* , the load angle θ _L , and the motor angle θ _M according to the first embodiment. Incidentally, FIG. 9, as an example, the value equivalent to the torsional stiffness K _R machinery 12 values of the virtual spring coefficient K _{C (spring} stiffness), when the value of the virtual damper coefficient T _C 0 (zero) It is the result of simulation.
As shown in FIG. 9, in the control according to the first embodiment, the load angle command value θ _L ^* and the load angle θ _L are substantially equal, and the motor angle θ _M together with the load angle θ _L exhibits self-excited vibration. No longer occurs.

Figure 10 is a graph showing the state of the self-excited vibration of the case of changing the virtual spring coefficient K _C according to the first embodiment. In the simulation shown in FIG. 10, not the value of the virtual damper coefficient T _C is varied.
As shown in FIG. 10, depending on the value of the virtual spring coefficient K _C, there is a case where the waveform collapses at the rising or if the response self-excited vibration can not be suppressed. For this reason, the values of the virtual spring coefficient K _C and the virtual damper coefficient T _C need to be set appropriately so that the self-excited vibration is suppressed and the waveform of the response rise is not collapsed. The method of setting the value, as described above, the value of the virtual spring coefficient K _C combined with torsional rigidity K _R of the machinery 12 _(spring stiffness), than the damping rate of the machine 12 the value of the virtual damper coefficient T _C A large value (attenuation rate of about 0.1) is preferable.

As described above, the control device 14 according to the first embodiment controls the mechanical device 12 that engages the drive shaft 20 and the load shaft 22 and transmits the force of the drive shaft 20 to the load shaft 22. . Then, the control device 14 calculates a phase difference between the drive shaft 20 and the load shaft 22, calculates a compensation value for compensating the calculated phase difference using a proportional element and a differential element, and uses the compensation value as a compensation value. The torque command value τ _M ^* for the motor 28 is subtracted.

  Therefore, the control device 14 according to the first embodiment can reduce the self-excited vibration of the mechanical device 12 due to the presence of the dead zone without modeling the dead zone element 48 which is a nonlinear component with a simple configuration. It can be reliably suppressed. In addition, since the control device 14 according to the first embodiment does not model the dead band element 48 that is a non-linear component, the control apparatus 14 has high robustness to the estimation error (modeling error) of the dead band element.

  Further, the compensator 64 is independent of the position controller 32 and the speed controller 34 that calculate the command value for the motor, and therefore the design is simple.

Further, parameters to be adjusted in order to calculate a compensation value X, for a virtual spring coefficient K _C, and a virtual damper coefficient T _C, the fine adjustment for obtaining a proper compensation value X is simple.

[Second Embodiment]
Hereinafter, a second embodiment of the present invention will be described.

  The schematic configuration of the mechanical system 10 according to the second embodiment is the same as the schematic configuration of the mechanical system 10 according to the first embodiment shown in FIG.

  FIG. 11 is a functional block diagram showing the configuration of the mechanical system 10 according to the second embodiment. In FIG. 11, the same components as those in FIG. 3 are denoted by the same reference numerals as those in FIG.

The control device 14 according to the second embodiment sets the control value for the motor 28 fed back by the compensation value X as the speed command value ω _M ^* output from the position controller 32.
For this reason, the control device 14 differentiates the phase difference δθ _ML calculated by the phase difference calculator 62 by the differentiator 90 and outputs the result to the compensator 92. That is, the speed difference δω _ML between the drive shaft 20 and the load shaft 22 is calculated by the differentiator 90.

The compensator 92 calculates a compensation value X that compensates the speed difference δω _ML calculated by the differentiator 90 using a proportional element and a differential element.

The compensation value subtracter 94 subtracts the speed command value ω _M ^* by the compensation value X calculated by the compensator 92.
Then, the speed command value ω _M ^* subtracted by the compensation value X is input to the subtractor 38. The subtractor 38 calculates a motor speed deviation δω _M that is a deviation between the speed command value ω _M ^* subtracted by the compensation value X and the motor speed ω _M and outputs the motor speed deviation δω _M to the speed controller 34.

Therefore, since the control device 14 according to the second embodiment feeds back the compensation value X to the speed command value ω _M ^* even when the compensation value X cannot be fed back to the torque command value τ _M ^* , it is simple. The configuration can suppress self-excited vibration.

[Third Embodiment]
Hereinafter, a third embodiment of the present invention will be described.

  The schematic configuration of the mechanical system 10 according to the third embodiment is the same as the schematic configuration of the mechanical system 10 according to the first embodiment shown in FIG.

  FIG. 12 is a functional block diagram showing the configuration of the mechanical system 10 according to the third embodiment. In FIG. 12, the same components as those in FIG. 3 are denoted by the same reference numerals as those in FIG.

The mechanical system 10 according to the third embodiment is configured as a semi-closed control system. That is, the mechanical system 10 according to the third embodiment does not detect the load angle θ _L or the load angular velocity ω _L that is the state quantity of the load shaft 22.

Therefore, the mechanical system 10 according to the third embodiment includes a load position estimation unit 100 that estimates the position of the load shaft 22 based on the speed that is the state quantity of the drive shaft 20.
Incidentally, the machine system 10 according to the third embodiment uses the detection value of the motor speed omega _M as the speed of the drive shaft 20 to be input to the load position estimating unit 100. However, not limited thereto, the load position torque command value indicating the torque of the drive shaft 20 in the estimation unit 100 tau _M ^* is input, it may be the motor speed is used which is calculated based on the torque command value tau _M ^* .
In the following description, the position of the load shaft 22 estimated by the load position estimation unit 100 is referred to as an estimated load angle θ _L ′.

FIG. 13 is a functional block diagram illustrating a configuration of the load position estimation unit 100.
Load position estimating unit 100 simulates a process of motor speed omega _M reaches the load angle theta _L.
Specifically, the load position estimation unit 100 calculates the motor angle indicating the position of the drive shaft 20 by integrating the input motor speed ω _M by the integration element 110, and subtracts it by the estimated load angle θ _L ′. , The dead zone Nx (dead zone element 112) and the torsional stiffness K _R (torsional stiffness element 114).

Furthermore, the load position estimation unit 100 calculates the load angular velocity ω _L by causing the load inertia element 116, the integration element 118, and the load viscous resistance element 120 simulating the load shaft 22 to act on the multiplied values. Then, the load angular velocity omega _L is integrated (integral element 122), the estimated load angle theta _{L 'is} calculated.
Note that the configuration of the load position estimation unit 100 shown in FIG. 13 is an example, and another configuration may be used.

The phase difference calculator 62 according to the third embodiment calculates a phase difference δθ _ML that is a deviation between the position of the drive shaft 20 and the estimated load angle θ _L ′ estimated by the load position estimation unit 100.

Then, the compensator 64 calculates a compensation value X that compensates the phase difference δθ _ML calculated by the phase difference calculator 62 using a proportional element and a differential element.

Further, the estimated load angle θ _L ′ estimated by the load position estimating unit 100 is output to the subtractor 36. Thereby, the deviation between the load angle command value θ _L ^* and the estimated load angle θ _L ′ is calculated by the subtractor 36 as the load angle deviation δθ _L.

As described above, the mechanical system 10 according to the third embodiment calculates the phase difference δθ _ML by estimating the position of the load shaft 22 even in a device configuration in which the state quantity of the load shaft 22 cannot be detected. it can.

The mechanical system 10 according to the third embodiment may be applied to a fully closed control system. For example, the load position estimation unit 100 estimates the estimated load angle θ _L when a lot of noise is superimposed on the detected value of the state quantity (for example, the motor speed ω _M ) of the load shaft 22 even in the fully closed control system. 'Is estimated, and the phase difference δθ _ML is calculated.

Furthermore, the mechanical system 10 according to the third embodiment, the motor speed omega _M as the quantity of state of the load shaft 22 has been described the configuration input to the load position estimating unit 100, the present invention is limited thereto not be a form in which the detection value of the motor angle theta _M indicating the position of the load axis 22 as the quantity of state of the load shaft 22 is directly input to the load position estimating unit 100. In the case of this form, the load position estimation unit 100 does not require the integration element 110.

  As mentioned above, although this invention was demonstrated using said each embodiment, the technical scope of this invention is not limited to the range as described in the said embodiment. Various changes or improvements can be added to the above-described embodiments without departing from the gist of the invention, and embodiments to which the changes or improvements are added are also included in the technical scope of the present invention.

  For example, in each of the above-described embodiments, the compensation value X has been described as being calculated using a proportional element (virtual spring element) and a differential element (virtual damper element). It is not limited, and the compensation value X may be calculated using only one of the proportional element (virtual spring element) and the differential element (virtual damper element).

  In each of the above-described embodiments, the gear is used to transmit the force of the drive shaft 20 to the load shaft 22, but the present invention is not limited to this, and the drive shaft 20 and the load As long as the shaft 22 meshes with the driving shaft 20 and the force of the driving shaft 20 can be transmitted to the load shaft 22, other forms such as a ball screw may be used.

  In each of the above embodiments, the drive shaft 20 is driven by the motor 28. However, the present invention is not limited to this. Other driving means may be used.

DESCRIPTION OF SYMBOLS 10 Mechanical system 12 Mechanical apparatus 14 Control apparatus 20 Drive shaft 22 Load shaft 62 Phase difference calculator (position deviation calculation means)
64 Compensator (compensation means)
66 Compensation value subtracter (subtraction means)
90 Differentiator (differentiating means)
92 Compensator (compensation means)
94 Compensation value subtractor (subtraction means)
100 Load position estimating unit (estimating means)

Claims (8)

A control device that controls a mechanical device that engages a drive shaft and a load shaft and transmits the force of the drive shaft to the load shaft,
Position deviation calculating means for calculating a deviation between the position of the drive shaft and the position of the load shaft;
Compensation means for calculating a compensation value for compensating the deviation calculated by the position deviation calculation means using at least one of a proportional element and a differential element;
Subtracting means for subtracting the control value for the drive shaft by the compensation value calculated by the compensating means;
A control device for a mechanical device.   The control device for a mechanical device according to claim 1, wherein the control value is a torque command value for the drive shaft. The control value is a speed command value for the drive shaft,
The control device for a mechanical device according to claim 1, further comprising a differentiating unit that differentiates the deviation calculated by the position deviation calculating unit and outputs the differentiated value to the compensating unit.   The control device for a mechanical device according to any one of claims 1 to 3, wherein the value of the proportional element is set in advance so as to be equal to a value of a spring stiffness of the mechanical device.   The control device for a mechanical device according to any one of claims 1 to 4, wherein the value of the differential element is set in advance so as to be a damping rate larger than a damping rate of vibration in the mechanical device. An estimation means for estimating a position of the load shaft based on a state quantity of the drive shaft;
The control device for a mechanical device according to any one of claims 1 to 5, wherein the position deviation calculation means calculates a deviation between the position of the drive shaft and the position of the load shaft estimated by the estimation means. . A mechanical device that engages a drive shaft and a load shaft and transmits the force of the drive shaft to the load shaft;
The control device according to any one of claims 1 to 6, wherein the mechanical device is a control target;
A mechanical system comprising: A control method for controlling a mechanical device that meshes with a drive shaft and a load shaft and transmits the force of the drive shaft to the load shaft,
A first step of calculating a deviation between the position of the drive shaft and the position of the load shaft;
A second step of calculating a compensation value for compensating the calculated deviation using at least one of a proportional element and a differential element;
A third step of subtracting the control value for the drive shaft by the calculated compensation value;
A control method for a mechanical device.

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