JP5146359B2 - Feedback control device - Google Patents

Description

  The present invention relates to a feedback control device that performs feedback control so that an output of a controlled object follows a target value.

  As a variable valve operating device of an engine, it has a control shaft that operates by driving a motor, and by operating this control shaft against the reaction force of a valve spring, the valve lift characteristics are variably controlled, and One having a relationship between the driving force of the motor and the operating angle of the control shaft having a non-linear characteristic is known.

  In such a variable valve system, as the camshaft rotates, a so-called cam reaction force causes the control shaft to swing through the link, as well as disturbances such as static friction and cogging when using a brushed DC motor. Having an element.

  When feedback control is performed so that the control shaft operating angle coincides with the target value using such a variable valve device as a control target, the control shaft swings due to the cam reaction force as described above. A signal that vibrates in synchronism with engine rotation is input to the control shaft operating angle detector. When the detected value of the control shaft operating angle with this vibration signal added is used as a feedback control signal, the motor current, which is the operation amount, vibrates to increase the power consumption, and the actuator is subjected to unnecessary load fluctuations to improve durability. There is also the possibility of damage. On the other hand, when the idle control is realized by the valve lift amount control, a very small operation angle control is required, and thus the disturbance as described above has a great influence on the target value followability of the operation angle control. In some cases, the idle control performance may be impaired.

  In order to solve these problems, a band elimination filter that reduces the output gain of the frequency band according to the vibration signal is provided, and a signal calculated by the band elimination filter is input as the operating angle signal of the control axis. A control device is disclosed in Patent Document 1.

JP 2000-282900 A

  However, in the configuration of the control device disclosed in Patent Document 1, when the cut frequency band of the band elimination filter is changed according to the engine rotation speed, for example, when the engine rotation speed is low, the frequency band of vibration noise becomes low. Since the control frequency band required for control response (frequency to be controlled determined from step response) is very close, the signal to be controlled is set when the width of the filter cut frequency is set wide. There is also a problem that the control performance is remarkably deteriorated because the detection itself of the control shaft operating angle is delayed because it is also removed by the filter. In addition, if the bandwidth of the filter cut frequency is set narrow, the oscillation frequency will deviate from the bandwidth of the filter cut frequency during acceleration / deceleration (transition time) when the engine speed changes, and the DC motor is caused by the oscillation. The current command value for driving the motor does not vibrate greatly, power consumption increases, and an unnecessary load may be applied to the DC motor.

The filter described above is a method that uses a band elimination filter directly on the control detection signal, but the control target is modeled and the deviation between the model output and the actual detection value is treated as all disturbances. A method that does not reflect the fluctuation due to the cam reaction force in the feedback control by passing a signal having a frequency higher than the predetermined frequency as noise including the fluctuation frequency with a high-pass filter and subtracting the filtered signal from the actual detection value. There is also.
Describe more specifically. First, the response output of the controlled object model is calculated when a predetermined input is made to the controlled object model in which the controlled object is modeled by a secondary transfer function. On the other hand, the actual output when the same input is made to the actual control target is detected. The difference between the response output and the detected value can be treated as a disturbance that cannot be expressed by the control model.
Of these disturbances, a high-pass filter is set so as to cut the cam reaction force fluctuation, and the fluctuation caused by this fluctuation is subtracted from the detected control shaft operating angle as a noise component to eliminate the fluctuation. Is also possible.

  However, in the configuration in which the swing due to the reaction force torque T1 of the control shaft operating angle is passed by the high-pass filter, the cutoff frequency of the high-pass filter is determined by the reaction force torque T1 in order to pass the swing due to the reaction force torque T1 with certainty. It is necessary to set the frequency as low as about 1/2 to 1/3 of the oscillation frequency. For this reason, even disturbances that should be suppressed by control, such as electrical noise or model error due to reaction torque change due to the lubrication state of the valve mechanism, other than noise to be removed from the detected value of the control shaft operating angle are detected as noise. Thus, although the vibration of the drive current of the DC motor can be suppressed, there is a problem that the disturbance resistance is deteriorated.

  Therefore, an object of the present invention is to provide a control device that enables feedback control with high disturbance resistance and stability.

  The control device of the present invention is a control device that performs feedback control for causing the actual output of the control target to follow the target value based on the control target value and the actual output of the control target. And detecting a disturbance to the controlled object based on a model output when a predetermined input is made to the controlled object model obtained by modeling the controlled object and an actual output of the controlled object when the input is made A band-pass filter that passes fluctuation components that are not reflected in feedback control among disturbances is provided, and the fluctuation component is added to feedback control by removing the band-pass filter passing signal that has passed through the band-pass filter from the actual output. It is a feedback control device that does not reflect. Furthermore, it has a pass bandwidth adjusting means for changing the pass bandwidth of the band pass filter based on the state of the controlled object, and the pass bandwidth adjusting means widens the pass bandwidth as the state change of the controlled object becomes larger. The pass bandwidth is narrowed as the state change of the controlled object becomes smaller.

  According to the present invention, it is possible to perform feedback control that suppresses the vibration amplitude of the drive current without greatly deteriorating the disturbance resistance and stability even when the state of the controlled object changes greatly.

It is a top view of the variable valve apparatus to which this embodiment is applied. It is principal part sectional drawing of a variable valve apparatus. It is a block diagram which shows the drive part of a variable valve apparatus. It is a figure for demonstrating the effect | action of a variable valve apparatus. It is a block diagram which shows the control shaft and control cam of a variable valve apparatus. It is a figure for demonstrating the effect | action in each rotation angle of a control shaft. It is a characteristic view which shows the relationship between a control shaft working angle and a drive current. It is a block diagram of the control axis positioning controller of this embodiment. It is a block diagram of a linearization compensator and a control object. It is a map of a linearization compensator. It is a block diagram of a rocking | swiveling remover (the 1). It is a block diagram of a rocking | swiveling remover (the 2). It is a figure which shows the simulation result in the case by this embodiment, and the case without rocking | fluctuation removal, (a) is a control shaft working angle, (b) has shown about drive current. It is a figure which shows the simulation result at the time of hold | maintaining a control axis to a fixed angle, (a) is a control axis working angle, (b) is a drive current, (c) has shown about electric power consumption. It is a figure which shows the simulation result in case the fluctuation | variation by the reaction force torque T1 of a control-shaft operating angle is estimated with a high-pass filter, (a) has shown the control-shaft operating angle, (b) has shown about drive current. It is a figure which shows the frequency distribution of the fluctuation | variation of a control-shaft operating angle at the time of applying this embodiment to a V type 6 cylinder engine. It is a time chart at the time of idling control when a pass band width is wide and narrow. It is a time chart when drive shaft rotational speed _NCAM changes in a state where a pass band width is narrow. It is a time chart in the case of switching the attenuation coefficient ζ based on the change rate of the drive shaft rotation speed. It is an attenuation coefficient table. It is a figure which shows the non-linear reaction force characteristic and the reaction force characteristic after linearization compensation.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 to FIG. 3 show an embodiment of a variable valve gear that can change the valve lift amount and the valve operating angle without changing the crank angle position at which the intake valve reaches the maximum lift.

  FIG. 1 is a plan view of the variable valve operating apparatus, FIG. 2 is a cross-sectional view of the main part of the variable valve operating apparatus, and FIG. 3 is a configuration diagram showing a drive unit of the variable valve operating apparatus. In FIG. 1, the configuration on the exhaust valve side (lower side in FIG. 1) is not shown.

  A drive shaft 11 (cam shaft) that is continuous over all cylinders is provided on the upper portion of the cylinder head 10. The drive shaft 11 has a sprocket attached to one end (not shown) and rotates in conjunction with a crankshaft of an engine (engine) via a timing chain or the like.

  A cylindrical bearing portion 18 a of a swing cam 18 that drives an intake valve (or exhaust valve) 19 is fitted on the outer periphery of the drive shaft 11 so as to be relatively rotatable. The swing cam 18 has a thin plate shape having a tip (cam nose) 18b, and a cam surface 18c is formed on the outer periphery of the swing cam 18 so as to be in sliding contact with an upper surface 19b of a valve lifter 19a as a transmission member provided at the upper end of the intake valve 19. ing.

A ring-shaped eccentric cam 12 is fixed to the outer periphery of the drive shaft 11 by press fitting or the like. The center (axial center) C ₂ of the eccentric cam 12 is eccentric by a predetermined amount with respect to the center (axial center) C ₁ of the drive shaft 11. A base portion 13a of a ring-shaped link 13 is fitted on the outer periphery of the eccentric cam 12 so as to be relatively rotatable via a bearing or the like. Note that the swing center (axial center) of the swing cam 18 coincides with the center C ₁ of the drive shaft 11.

  As shown in FIG. 1, a control shaft 14 extends in the cylinder row direction substantially parallel to the drive shaft 11, obliquely above the drive shaft 11. The control shaft 14 is rotated and held in a predetermined rotational speed range by a drive unit 20 described later according to the operating state of the engine.

A ring-shaped control cam 15 is fixed to the outer periphery of the control shaft 14 by press fitting or the like. The center (axial center) C ₄ of the control cam 15 is eccentric by a predetermined amount with respect to the center (axial center) C ₃ of the control shaft 14. A cylindrical central base of the rocker arm 16 is fitted on the outer periphery of the control cam 15 so as to be relatively rotatable. The one end portion 16a of the rocker arm 16 and the small-diameter tip portion 13b of the ring-shaped link 13 are connected to each other via a first pin 29a that passes through the both ends 16a and 13b.

  The other end 16 b of the rocker arm 16 and the swing cam 18 are linked by a rod-shaped link 17. More specifically, the other end portion 16b of the rocker arm 16 and the one end portion 17a of the rod-like link 17 are connected to each other via a second pin 29b that passes through both the portions 16b and 17a so as to be relatively rotatable. Further, the other end 17b of the rod-shaped link 17 and the swing cam 18 are coupled to each other via a third pin 29c through which both the ends 17b and 18 are inserted.

  Next, the configuration of the drive unit 20 that rotates and holds the control shaft 14 will be described.

  As shown in FIG. 1, the control shaft 14 extends into a case 22 fixed to the cylinder head 10, and a worm wheel 21 is fixed to one end thereof. A DC motor 26 driven by a control signal from the control axis positioning controller 50 is attached to the case 22, and an output shaft 26 a of the DC motor 26 can be rotated into the case 22 via a roller bearing 25. It is extended. A worm gear 24 that meshes with the worm wheel 21 is fixed to the output shaft 26a. In order to increase the motor torque between the worm gear 24 and the worm wheel 21, the gear ratio is set appropriately large. The case 22 is provided with a control shaft operating angle sensor 23 for detecting the rotation angle (control shaft operating angle) of the control shaft 14 (worm wheel 21). The output of the control shaft operating angle sensor 23 is The DC motor 26 is feedback-controlled based on the operating angle of the control shaft 14 that is input to the control shaft positioning controller 50 and detected by the control shaft operating angle sensor 23.

With such a configuration, when the drive shaft 11 rotates in conjunction with the rotation of the engine, the ring-shaped link 13 moves in translation via the eccentric cam 12, and the rocker arm 16 corresponds to the center C _{4 of the} control cam 15. And the swing cam 18 swings through the rod-shaped link 17. At this time, the cam surface 18c of the swing cam 18 is in sliding contact with the upper surface of the valve lifter 19a as a transmission member provided at the upper end of the intake valve 19, and the valve lifter 19a is pressed against the reaction force of a valve spring (not shown). As a result, the intake valve 19 opens and closes in conjunction with the rotation of the engine.

Further, when the output shaft 26a of the DC motor 26 is rotationally driven by a predetermined angle according to the operating state of the engine, the control shaft 14 is rotated by a predetermined angle (operating angle) via the worm gear 24 and the worm wheel 21. As a result, the position of the center C ₄ of the control cam 15 serving as the rocking center of the rocker arm 16 changes, and the lift characteristic of the intake valve 19 changes. More specifically, as the operating angle of the control shaft 14 is rotated to the larger side and the distance between the center C ₄ of the control cam 15 and the center C ₁ of the drive shaft 11 is closer, the valve lift characteristic is displaced. The lift amount and valve operating angle increase.

  Next, the operating characteristics of the variable valve operating device will be discussed with reference to FIGS.

A reaction force F1 generated by a valve spring reaction force of the intake valve 19 acts on the other end portion 16b of the rocker arm 16 via the swing cam 18, the rod-shaped link 17, the second pin 29b, and the like. A reaction force F2 generated as a reaction acts on the one end 16a of the rocker arm 16 via the eccentric cam 12, the ring-shaped link 13, the first pin 29a, and the like. Therefore, the swing center C ₄ of the rocker arm 16, the synthesis reaction force F3 substantially reaction force F1, F2 acts.

  As a result, a torque T1 that is the product of the arm length r1 from the center C3 of the control shaft 14 to the direction line of the combined reaction force F3 and the combined reaction force F3 acts on the control shaft 14. Therefore, in order for the drive unit 20 to hold the control shaft 14 at a predetermined angle, at least a reverse torque that matches the torque T1 is required.

In a state where the control shaft 14 is held at a predetermined rotation angle (operating angle), as shown in FIG. 4, when the swing cam 18 is pushed down to the highest valve lift side, that is, counterclockwise in FIG. The combined reaction force F3 is maximized when it swings most. Direction of the resultant reaction force F3 at this time is substantially parallel to the first line L1 connecting the center C ₃ of the center C ₁ and the control shaft 14 of the drive shaft 11.

  Here, as shown in FIG. 6, the combined reaction force F3 increases as the valve lift amount (valve operating angle) increases (left side in FIG. 6: minimum valve operating angle, middle in FIG. 6, middle valve position, FIG. 6 right side: maximum valve operating angle), the arm length r1 increases from the minimum valve operating angle to the intermediate position as the eccentric cam 12 rotates, but decreases thereafter. Therefore, the torque T1, which is the product of the combined reaction force F3 and the arm length, has a non-linear characteristic with respect to the control shaft operating angle θ.

  FIG. 7 shows the relationship between the operating angle θcs of the control shaft 14 (the valve lift amount of the intake valve 19 is correlated with the valve operating angle) and the driving current ics (proportional to the torque T1) flowing through the DC motor 26. In FIG. 7, when the control shaft operating angle θcs is zero, the valve lift amount (valve operating angle) is minimum, and when the control shaft operating angle θcs is 60 degrees, the valve lift amount (valve operating angle) is maximum.

  Further, when the drive shaft 11 rotates in synchronization with the engine rotation, the reaction force torque T <b> 1 varies with the rotation of the drive shaft 11. Therefore, the operating angle of the control shaft 14 swings due to the fluctuation of the reaction force torque T1. The oscillation cycle has a frequency synchronized with the engine rotation speed, that is, the rotation speed of the drive shaft 11.

  By the way, when the control shaft operating angle sensor 23 detects the control shaft operating angle, the swing caused by the reaction force torque T1 is detected in a superimposed manner. There is a problem that the current command value for driving the motor 26 oscillates greatly, power consumption increases, and an unnecessary load is applied to the DC motor 26.

  As a means for solving this problem, there is a method in which a fluctuation component due to cam reaction force is removed using a band elimination filter and is not reflected in feedback control. Specifically, since this oscillation frequency changes according to the engine speed, for example, band elimination that removes a signal in a frequency band corresponding to the engine speed (reducing the output) as described above. A method of taking out the control axis operating angle signal using a filter is conceivable.

  However, as described above, when the cut frequency band of the band elimination filter is changed according to the engine speed, for example, when the engine speed is low, the frequency band of vibration noise is lowered, which is required from the control responsiveness. Control frequency band (frequency to be controlled determined by the step response) is very close, so if the filter cut frequency is set wide, the signal to be controlled itself is also removed by the filter and the control axis There is a problem that the control performance is significantly deteriorated because the detection of the operating angle is delayed. Also, if the bandwidth of the filter cut frequency is set narrow, the oscillation frequency will deviate from the bandwidth of the filter cut frequency during acceleration / deceleration (transition) when the engine rotation speed changes, and the oscillation will be reduced. The current command value for driving the motor 26 oscillates greatly, power consumption increases, and an unnecessary load may be applied to the DC motor 26.

  The filter described above is a method that uses a band elimination filter directly on the control detection signal, but the control target is modeled and the deviation between the model output and the actual detection value is treated as all disturbances. A method that does not reflect the fluctuation due to the cam reaction force in the feedback control by passing a signal having a frequency higher than the predetermined frequency as noise including the fluctuation frequency with a high-pass filter and subtracting the filtered signal from the actual detection value. There is also.

Describe more specifically. First, the response output of the controlled object model is calculated when a predetermined input is made to the controlled object model in which the controlled object is modeled by a secondary transfer function. On the other hand, the actual output when the same input is made to the actual control target is detected. The difference between the response output and the detected value can be treated as a disturbance that cannot be expressed by the control model.
Of these disturbances, a high-pass filter is set so as to cut the cam reaction force fluctuation, and the fluctuation caused by this fluctuation is subtracted from the detected control shaft operating angle as a noise component to eliminate the fluctuation. Is also possible.

  However, in the configuration in which the swing due to the reaction force torque T1 of the control shaft operating angle is passed by the high-pass filter, the cutoff frequency of the high-pass filter is determined by the reaction force torque T1 in order to pass the swing due to the reaction force torque T1 with certainty. It is necessary to set the frequency as low as about 1/2 to 1/3 of the oscillation frequency. For this reason, even disturbances that should be suppressed by control, such as electrical noise or model error due to reaction torque change due to the lubrication state of the valve mechanism, other than noise to be removed from the detected value of the control shaft operating angle are detected as noise. Thus, as shown in FIGS. 15A and 15B, although the vibration of the drive current of the DC motor 26 can be suppressed, the disturbance resistance is deteriorated. FIG. 15 shows the result of simulation with the variable valve operating apparatus as the control target when the control shaft operating angle detection value is used as it is as the F / B signal and when the estimation is performed using the high-pass filter. 15A shows the control shaft operating angle, and FIG. 15B shows the drive current. A PID compensator is used as the feedback compensator, and the engine rotational speed is set to 800 rpm, and a step disturbance corresponding to the driving current of 3 A is applied at the time of 1 second.

  Therefore, in the present embodiment, the following control axis positioning control is performed.

  FIG. 8 is a control block diagram of positioning control of the control shaft operating angle θcs of the variable valve mechanism, which is executed by the control shaft positioning controller 50. The control axis positioning controller 50 includes a linearization compensator 61, a feedback compensator 62, and a fluctuation remover 63 that removes fluctuations (noise) caused by the reaction force torque T1.

  The variable valve apparatus having the DC motor 26 shown in FIGS. 1 to 3 is the control object 60, the input of the control object 60 is the current command value icsc given to the DC motor 26, and the output (control amount) is the control shaft operating angle. θcs. This control shaft operating angle θcs can be detected by a control shaft operating angle sensor 23 as response detection means for detecting the output of the controlled object 60.

  The linearization compensator 61 will be described with reference to FIG. FIG. 9 is a detailed block diagram of the linearization compensator 61 and the controlled object 60. Kt is a motor torque constant, J is an inertia around the axis of the DC motor 26, D is a viscous friction coefficient of the lubricating oil, T1 is a non-linear (not first-order) reaction torque that changes according to the control shaft operating angle, and s is Laplace operator. The controlled object 60 has a characteristic represented by a second-order transfer function having a nonlinear reaction torque that changes according to the control shaft operating angle.

In terms of control management, the nonlinear model has problems such as complicated logic when designing the controller. Therefore, the linearization compensator 61 is a compensator in order to make the nonlinear characteristic of the torque reaction force virtually linear. Normally, a compensator is created using a map of the nonlinear characteristic of torque reaction force according to the operating angle, but here, the linearization compensation amount I according to the control shaft operating angle θcs created by measurement in advance is used. _{Use the LC} map. The reason for using the linearization compensation amount ILC _instead of the torque reaction force is that the torque generated by the DC motor 26 with respect to the current value of the DC motor 26 is proportional to the motor torque constant Kt, which is a proportional coefficient.
The map used here is shown in FIG. In FIG. 10, the linearization compensation amount is zero when the control shaft operating angle is zero to 40 degrees, and when it exceeds 40 degrees, the linearization compensation amount I _LC decreases as the control shaft operating angle increases. FIG. 21 shows a nonlinear reaction force characteristic and a reaction force characteristic after linearization compensation.

By adding the linearization compensation amount I _LC to the current command value I _CMD and giving it to the controlled object, the controlled object model including the nonlinear reaction force torque T1 becomes a linear model. As a result, the linearized transfer function from the current command value I _CMD of the controlled object 60 to the control shaft operating angle θcs is as shown in Equation (1), which is a quadratic transfer function. K is the slope of the control axis reaction force after linearization compensation.

The feedback compensator 62 will be described. The feedback compensator 62 is a PID control compensator. The current command value I _CMD that is the output of the PID compensator is calculated by equation (2).

K _P is a proportional gain, K _I is an integral gain, K _D is a differential gain, and T _D is a time constant of approximate differentiation. θCMD is an angle command value and changes according to the operation of the driver and the operating state of the engine. θcs2 is a feedback signal calculated by a swing remover 63 described later.

  The swing remover 63 will be described with reference to FIG. FIG. 11 is a block diagram showing the configuration of the swing remover 63. The swing remover 63 includes a control target model 63a, a high-pass filter 63b, band-pass filters 63c to 63e, and subtractors 63f, 63g, 63h, and 63i.

The controlled object model 63a has a transfer function G _P (s) similar to that of the equation (1), and is provided to output an ideal signal when there is no disturbance with respect to the current command value. The transfer function of the high pass filter 63b is expressed by Expression (3). The band-pass filters 63c to 63e, which will be described later, want to pass only the signal of the oscillation frequency of the cam reaction force. However, because of the design of the band-pass filters 63c to 63e, signals around the passing frequency band are also passed. The signal other than the cam reaction force frequency is allowed to pass through the band-pass filter 63b by removing in advance the signal on the lower frequency side than the frequency to be passed from the signal input to the band-pass filter by the high-pass filter 63b. This is restrained. The high-pass filter 63b is variably set according to the engine speed (linked to the oscillation frequency) in the same manner as the band-pass filters 63c to 63e.

T _H is a time constant of the high-pass filter, and its value is set so that the gain is zero dB in order to pass the signal as it is in the pass frequency band of band-pass filters 63c to 63e described later.

  The signal θ1 input to the bandpass filters 63c to 63e is calculated by Expression (4) by using the current command value ICMD and the control operating angle θcs as inputs by the subtractor 63f.

  The bandpass filters 63c to 63e have a transfer function of Expression (5).

In the above equation (5), n is 1 for the bandpass filter 63c, 2 for the bandpass filter 63d, and 3 for the bandpass filter 63e. ζ is a damping coefficient, ωn is a natural frequency, and is variably adjusted based on Expression (6) using the drive shaft rotational speed N _CAM . The unit here ωn is rad / s, the unit of N _CAM is Hz.

  That is, each of the bandpass filters 63c to 63e is set to pass the same frequency as the drive shaft rotation speed and twice or three times the drive shaft rotation speed.

  Next, the attenuation coefficient ζ will be described.

As the attenuation coefficient ζ, an attenuation coefficient ζ1 in which the passband widths of the bandpass filters 63c to 63e are relatively small and an attenuation coefficient ζ2 (that is, ζ2> ζ1) that is also relatively large are set in advance. Either one is selected and used based on the estimation result of the change rate of the drive shaft rotational speed _NCAM .

Here, damping coefficients ζ1 are pass-band width of the filter 63c~63e is, the drive shaft rotational speed N _CAM is very narrow includes a frequency of the swing of the control shaft operating angle due to the reaction force torque T1 in the case of a range Is such a value.

On the other hand, the attenuation coefficient ζ2 is such that the passband width of the band-pass filters 63c to 63e is wider than that of the attenuation coefficient ζ1, and the drive shaft rotational speed N _CAM that the filter width is desired to be widened as a measure against the deviation at the time of transition. At the time of fluctuation, a value is set so that the pass band width of the filter is widened so that the vibration width of the drive current can be allowed as a countermeasure against the deviation at the time of transition (that is, when the change in the rotation speed is large, the bandpass filters 63c to 63e). Increases the passband width).

  Next, the method for estimating the drive shaft rotation speed change rate and the method for selecting the damping coefficient will be described with specific examples.

(First method)
Detected by calculating a change rate of the drive shaft rotation speed N _CAM, if this rate of change is smaller than a preset threshold selects the damping coefficient .zeta.1, if it is greater, it selects the damping coefficient ?? 2. This threshold value is set by monitoring the drive current vibration generated with the change in the drive shaft rotational speed N _CAM when the damping coefficient ζ1 is used, and experimentally determining an allowable upper limit value for the drive current vibration. The change rate of the drive shaft rotational speed N _CAM when the allowable upper limit value is reached is set as a threshold value.

  When the pass band width of the bandpass filters 63c to 63e is narrowed and set near the oscillation frequency of the control shaft operating angle in question, the frequency component to be removed from the feedback signal is changed into the oscillation of the control shaft operating angle in question. The frequency can be almost limited. As a result, the frequency band necessary for feedback control is not removed, and higher disturbance resistance and stability can be realized to improve the control performance. In particular, in the case of idle control in which the throttle valve is opened and the low air volume is controlled by adjusting the operating angle with a variable valve system, high resolution is required because the change in the air volume with respect to the change in the operating angle is large. And In such a region, it is necessary to control the operating angle by a minute angle, and since it is affected by static friction and cogging of the DC motor, the pass band width should be narrowed so that only the oscillation frequency due to the cam reaction force can pass. It is necessary to improve the high disturbance resistance.

  FIG. 17 is a time chart when the pass bandwidth is wide and when it is narrow during idle control. The solid line in the figure indicates the case where the passband width is wide, and the broken line indicates the case where it is narrow. When stepwise disturbance is applied at the time point t1, the disturbance of the driving current is reduced because the disturbance is cut as noise, as compared with the case where the passband width is wide. However, the control shaft operating angle is slower to converge when the pass band width is wide than when it is narrow.

On the other hand, the oscillation frequency of the control shaft changes according to the drive shaft rotational speed N _CAM , so the center frequencies of the bandpass filters 63c to 63e must be set according to the drive shaft rotational speed N _CAM . The drive shaft rotational speed NCAM has a deviation from the actual value at least due to the influence of detection delay and the like. For this reason, in a situation where the pass band widths of the bandpass filters 63c to 63e are set to be narrow and the drive shaft rotational speed detection value is delayed from the actual value, the drive shaft rotational speed N _CAM changes greatly. The actual oscillation frequency of the control shaft operating angle and the center frequency of the band-pass filters 63c to 63e do not coincide with each other, and as a result, the vibration of the drive current increases.

FIG. 18 is a time chart when the drive shaft rotational speed N _CAM changes in a state where the pass bandwidth is narrow. Specifically, a disturbance is applied stepwise at the time t1, the drive shaft rotational speed NCAM increases from t3 to t4, and the drive shaft rotational speed decreases from t7 to t8.

  Even if the disturbance is applied at t1, the control shaft operating angle and the drive current have already converged to the original values at a time before t2. However, at t3 to t4 and t7 to t8 where the drive shaft rotational speed NCAM changes greatly, as described above, the oscillation frequency of the control shaft operating angle and the center frequency of the bandpass filters 63c to 63e become inconsistent, and the drive current Vibrates greatly.

  On the other hand, in this embodiment, when the change rate of the drive shaft rotational speed NCAM is small, the passband width of the bandpass filters 63c to 63e is narrowed, and when the change rate is large, the passbands of the bandpass filters 63c to 63e are used. Switch the filter characteristics to widen the width.

FIG. 19 is a time chart when the attenuation coefficient ζ is switched based on the change rate of the drive shaft rotation speed. The damping coefficient ζ2 is set for t3 to t4 and t7 to t8 at which the drive shaft rotational speed greatly changes, and the damping coefficient ζ1 is set for others. Thereby, the vibration of the drive current at t3 to t4 and t7 to t8 is suppressed compared to the case of FIG. As described above, in the region where the change in the drive shaft rotational speed N _CAM is small, it is possible to suppress the amplitude of the drive current in the region where the change in the drive shaft rotational speed N _CAM is large while realizing high disturbance resistance.

(Second method)
As an index for determining the change rate of the drive shaft rotational speed, an idle control execution flag for determining whether or not the idle control is being executed is provided, and either the damping coefficient ζ1 or the damping coefficient ζ2 is determined based on the idle control execution flag. Select. Specifically, if the idle control is being executed, the rate of change of the drive shaft rotation speed is small, so the attenuation coefficient ζ1 is selected so that the pass bandwidth is narrow, and otherwise the pass bandwidth is widened. A damping coefficient ζ2 is selected. Thereby, the target value followability of the valve lift amount control of the intake valve 19 can be improved, and the engine speed fluctuation can be suppressed.

(Third method)
In the case of a vehicle having a so-called auto-cruise control function that controls to maintain the vehicle speed set by the driver, auto-cruise indicating whether or not auto-cruise control is being performed as an index for determining the change rate of the drive shaft rotation speed A control execution flag is used. If the auto-cruise control is being executed, the rate of change of the drive shaft rotational speed is small. Therefore, the attenuation coefficient ζ1 is selected so that the pass bandwidth is narrow, and otherwise, the attenuation coefficient is set so that the pass bandwidth is wide. Select ζ2. Thereby, the target value followability of the valve lift amount control of the intake valve 19 is improved as in the second method, so that it is possible to reduce the vehicle speed fluctuation at the time of load fluctuation due to a change in road gradient or the like.

(Fourth method)
As the index for determining the change rate of the drive shaft rotation speed, the change rate of the acceleration operation amount is used. Acceleration opening change rate or throttle opening change rate is calculated as the change rate of the acceleration operation amount, and when either opening change rate is larger than the threshold value, the change rate of the drive shaft rotational speed _NCAM becomes large. Since the possibility is high, the attenuation coefficient ζ2 is selected so as to widen the pass bandwidth, and otherwise, the attenuation coefficient ζ1 is selected so as to narrow the pass bandwidth.

The threshold value used here defines an allowable upper limit value for the drive current vibration generated with the change of the drive shaft rotational speed N _CAM when the damping coefficient ζ1 is used, and the accelerator opening or the throttle at which the allowable upper limit value is reached. Set the rate of change of opening.

In this way, when it can be determined that the drive shaft rotational speed N _CAM fluctuates clearly, it is possible to reliably reduce the vibration width of the drive current when the drive shaft rotational speed fluctuates by widening the pass band width in advance.

(Fifth method)
When the transmission is a continuously variable automatic transmission, gear ratio information is used as an index for determining the rate of change of the drive shaft rotation speed. The current gear ratio is used as the gear ratio information. If the gear ratio is large, the change in the drive shaft rotational speed N _CAM when the engine torque varies or the vehicle state changes increases. Further, in the case of a continuously variable transmission, the gear ratio reflects the driver's intention to accelerate such as the accelerator pedal depression amount.

  It should be noted that the speed ratio here is relative, and the speed ratio for idle operation or low speed travel is assumed to be “large”.

  Therefore, when the gear ratio is large, the damping coefficient ζ2 that increases the passband width is selected, and when not, the damping coefficient ζ1 is selected. As a result, it is possible to reliably suppress vibration of the drive current when the drive shaft rotational speed fluctuates. Further, even when engine torque or vehicle state changes, it is possible to avoid a decrease in control performance of the variable valve gear without unnecessarily widening the passband width. Instead of the current gear ratio, the change ratio of the gear ratio may be used.

  Moreover, you may use together the 1st-5th method mentioned above with another method, respectively.

Bandpass filter 63c~63e the damping coefficient ζ is switched as described above is not used when the drive shaft rotational speed N _CAM is in a predetermined condition, the output of the band-pass filter becomes zero when not in use.

  The non-use condition is determined according to the frequency component of the swing of the control shaft operating angle at each drive shaft rotational speed measured in advance. The frequency distribution of the oscillation of the variable valve system changes according to the rotational speed of the drive shaft (camshaft) depending on the mechanical configuration (link mechanism and mechanical natural frequency) of the variable valve system. Therefore, here, as an example, a case where the present embodiment is applied to a V-type 6-cylinder engine will be described with reference to FIG.

  FIG. 16 shows the frequency distribution analysis result of the oscillation when the rotational speed of the drive shaft (camshaft) is 100 rpm and 1200 rpm. In both cases of 100 rpm and 1200 rpm, there is oscillation having the same frequency as the drive shaft (camshaft) rotation speed and a frequency three times as high. However, a frequency component that is twice the rotational speed of the drive shaft (camshaft) that exists at 1200 rpm does not exist at 100 rpm. That is, in the case of 100 rpm, even if the band-pass filter 63d that passes twice the frequency of the drive shaft (camshaft) rotation speed is used, there is no effect in suppressing vibration of the drive current.

  Therefore, in the case of 100 rpm, the bandpass filter 63d is not used, thereby suppressing an unnecessary decrease in the control performance (disturbance resistance and stability).

  As described above, when there is no distribution of fluctuation of the control shaft operating angle that any of the bandpass filters 63c to 63e passes within a certain range of the rotational speed of the drive shaft (camshaft), or when the power consumption or the DC motor 26 is increased. If it is small enough not to affect the load on the drive shaft, the rotational speed of the drive shaft is set as a range not using the band-pass filter.

  Θ2 which is the sum total of the operating angle fluctuation amounts that have passed through the bandpass filters 63c to 63e configured as described above is obtained by Expression (7).

  θ2 is the swing amount of the control shaft operating angle, and this is subtracted from the control shaft operating angle θcs by the subtractor 63i as shown in Expression (8), so that the feedback signal θcs2 becomes a signal from which the swing is removed.

  The swing remover 63 is not limited to the configuration shown in FIG. 11. For example, as shown in FIG. 12, the same effect can be obtained even when the bandpass filters 63 c to 63 e are connected in series. be able to. In this case, the high-pass filters 63b1 to 63b3 set time constants such that the gain is zero dB in the pass frequency bands of the bandpass filters 63c to 63e, respectively.

  FIG. 13 is a simulation result when the positioning control of the present embodiment is executed and when the detected value of the control shaft operating angle is used as it is as an F / B signal (hereinafter referred to as no swing removal). 13A shows the control shaft operating angle, and FIG. 13B shows the drive current.

  In this simulation, a PID compensator set to the same F / B gain as in the simulation of FIG. 15 is used, and a step disturbance equivalent to the driving current 3A is applied at the time of 1 second.

  When the fluctuation due to the reaction force torque T1 of the control shaft operating angle is estimated by the high-pass filter, the vibration of the drive current can be made smaller than that without the fluctuation removal as shown in FIG. However, the change of the control shaft operating angle after the step disturbance is applied is larger than that in the case of no swing removal as shown in FIG.

  On the other hand, according to the present embodiment, the change in the control operating angle after the step disturbance is applied is substantially the same as in the case of no swing removal as shown in FIG. As shown in FIG. 13B, the vibration of the drive current is smaller than that in the case where the oscillation is not removed and in the case where the oscillation due to the reaction force torque T1 of the control shaft operating angle is estimated by the high pass filter.

  That is, according to the present embodiment, it is possible to suppress the vibration amplitude of the operation amount input to the controlled object without deteriorating the control performance (disturbance resistance and stability). Therefore, a larger F / B gain can be set, power consumption can be reduced, and unnecessary load on the controlled object can be avoided.

  FIG. 14 shows the case where the positioning control of the present embodiment is executed and the case where no oscillation is removed with respect to the control shaft operating angle, the drive current, and the power consumption when the control shaft operating angle is held at a constant angle. FIG. 14 (a) shows the control shaft operating angle, FIG. 14 (b) shows the drive current, and FIG. 14 (c) shows the power consumption.

  As shown in FIG. 14 (a), the change in the control shaft operating angle is almost the same in any case, but as shown in FIG. 14 (b), the vibration of the drive current is suppressed according to this embodiment. ing. Therefore, as shown in FIG. 14C, the power consumption can be reduced, and an unnecessary load is not applied to the DC motor 26, and the durability of the DC motor 26 can be improved.

  As described above, according to the present embodiment, the following effects can be obtained.

The passband width of (1) a band-pass filter 63 c to 63 e, spread as a change in the drive shaft rotational speed N _CAM increases, the narrowed more change is small, even when the drive shaft rotational speed N _CAM is greatly changed, resistance The vibration amplitude of the drive current can be suppressed without greatly deteriorating the disturbance and stability.

  (2) Since a plurality of bandpass filters 63 are provided, the vibration amplitude of the manipulated variable can be effectively suppressed when there are a plurality of vibration noise frequency components. In addition, since the use / non-use of each of the bandpass filters 63c to 63e is switched according to the state of the control shaft 14, the control performance is deteriorated by using a bandpass filter that is not related to suppression of the vibration amplitude of the control output current. Necessary decline can be suppressed.

(3) When it is determined that the engine is in the idle control state, the pass bandwidth is narrowed, so that the target value followability of the lift amount control of the variable valve system is improved, and the change in the engine speed, that is, the drive shaft speed N It is possible to suppress engine speed fluctuations in the idle control state where the change in _CAM is small.

  (4) When it is determined that the auto-cruise control is being performed, the pass bandwidth is narrowed. Therefore, by improving the target value followability of the variable valve device, the vehicle speed fluctuation at the time of load fluctuation such as a change in road gradient is reduced. Can be reduced.

(5) A change in the acceleration operation amount is detected, and when the change is larger than the threshold, the pass bandwidth is widened. Therefore, when it is predicted that the drive shaft rotational speed N _CAM will obviously change, the pass bandwidth is prior to the change. Spread. As a result, the vibration amplitude of the drive current when the drive shaft rotational speed _NCAM varies can be reliably reduced.

  (6) The larger the gear ratio, the wider the pass band width, and the smaller the gear ratio, the narrower the pass band width. Therefore, in a state where the engine rotational speed fluctuation tends to increase due to engine torque fluctuation, etc. If not, the pass bandwidth is narrowed. Therefore, it is possible to reduce the vibration of the drive current that accompanies fluctuations in the engine rotational speed, and it is not necessary to unnecessarily widen the pass bandwidth, so that it is possible to suppress a decrease in controllability of the variable valve operating apparatus.

  Instead of switching the attenuation coefficient ζ of the bandpass filters 63c to 63e, a bandpass filter having an attenuation coefficient ζ1 and a bandpass filter having an attenuation coefficient ζ2 may be provided. In this case, it is determined based on the change rate of the drive shaft rotation speed which bandpass filter calculation result of which attenuation coefficient ζ is used.

  Furthermore, the calculation results of both bandpass filters may be linearly complemented.

  A second embodiment will be described.

  In this embodiment, the configuration and control are basically the same as those in the first embodiment, but switching control of the pass bandwidth of the bandpass filters 63c to 63e, particularly, switching from a wide pass bandwidth to a narrow pass bandwidth. The control is different.

In the present embodiment, the attenuation coefficient ζ is switched from ζ2 to ζ1 when the determination that switching to a narrow passband width by the same method as in the first embodiment continues for a predetermined time set in advance. The predetermined time here is set to a time equivalent to or longer than the time until the drive current oscillation caused by the change in the drive shaft rotational speed converges after the change in the drive shaft rotational speed N _CAM has subsided.

  The drive current oscillation caused by changing the passband width may remain due to the transient characteristics of the bandpass filters 63c to 63e even after the state change of the control shaft operating angle becomes small. Therefore, by switching the attenuation coefficient ζ after the predetermined time has elapsed, it is possible to reduce the drive current oscillation due to the transient characteristics of the bandpass filters 63c to 63e as described above.

  As described above, according to the present embodiment, the following effects can be obtained in addition to the same effects as those of the first embodiment.

  (1) When narrowing the passband once widened, the passbandwidth is narrowed after the state where the passbandwidth should be narrowed continues for a predetermined time, so that the drive after the state change of the control shaft operating angle becomes small A decrease in controllability due to current vibration can be suppressed.

  A third embodiment will be described.

  The configuration and control of this embodiment are basically the same as those of the first embodiment, but the switching control of the pass bandwidth of the bandpass filters 63c to 63e is different. In the present embodiment, instead of selecting the attenuation coefficient of ζ1 or ζ2, for example, it is continuously changed using a table as shown in FIG. In FIG. 20, the vertical axis represents the attenuation coefficient, and the horizontal axis represents the drive shaft rotation speed change rate. Instead of the change rate of the drive shaft rotation speed, the change rate of the acceleration operation amount such as the accelerator opening or the throttle opening, or the gear ratio may be used.

  According to this, the control performance can be further improved.

  In addition, although the case where it applied to the control apparatus of the variable valve operating apparatus of an engine was mentioned as an example and demonstrated, it is not restricted to this, It applies widely to the control apparatus of the mechanism in which the noise near a control frequency band generate | occur | produces. can do.

  The present invention is not limited to the above-described embodiment, and it goes without saying that various modifications can be made within the scope of the technical idea described in the claims.

DESCRIPTION OF SYMBOLS 11 Drive shaft 12 Eccentric cam 13 Ring-shaped link 14 Control shaft 15 Control cam 16 Rocker arm 17 Rod-shaped link 18 Swing cam 19 Intake valve 20 Drive part 23 Control shaft operating angle sensor 26 DC motor 50 Control shaft positioning controller 60 Control object 61 Linearization Compensator 62 Feedback Compensator 63 Oscillator Remover 63b High Pass Filter 63c Band Pass Filter 63d Band Pass Filter 63e Band Pass Filter

Claims (8)

A control device that performs feedback control for causing the actual output of the control target to follow the target value based on the control target value and the actual output of the control target. Based on the model output when the input is made and the output of the actual control object when the input is made, the disturbance to the controlled object is detected, and the fluctuation component of the disturbance that is not reflected in the feedback control In the feedback control device that does not reflect the oscillation component in the feedback control by removing the band pass filter passing signal that has passed through the band pass filter from the actual output,
A pass bandwidth adjusting means for changing the pass bandwidth of the band pass filter based on the state of the control target;
The control device characterized in that the pass bandwidth adjusting means widens the pass bandwidth as the state change of the control target increases, and narrows the pass bandwidth as the state change of the control target decreases.   The control device according to claim 1, comprising a plurality of the band-pass filters, and switching use / non-use of each band-pass filter according to the state of the control target. A control device for a variable valve operating system of an engine,
The actual control target output is the operating angle of the control shaft,
The input to the control object is a drive current or voltage,
The controlled object model is a controlled object model of a variable valve gear,
Of the disturbance, the swing component that is not reflected in the feedback control is a swing component due to the operating angle cam reaction force of the control shaft,
The pass bandwidth adjusting means widens the pass bandwidth as the change in the drive shaft rotational speed of the variable valve operating device increases, and narrows the pass bandwidth as the change in the drive shaft rotational speed decreases. The control device according to claim 1 or 2. Comprising idle state determination means for determining whether or not the engine is in an idle control state;
4. The control device according to claim 3, wherein when it is determined that the control unit is in an idle control state, the pass bandwidth adjusting unit narrows the pass bandwidth. Constant speed control means for controlling the vehicle speed so that the vehicle runs at a constant speed at a vehicle speed set by the driver;
Constant speed control state determination means for determining whether or not the vehicle speed control by the constant speed control means is in progress;
With
5. The control device according to claim 3, wherein, when it is determined that the vehicle speed control by the constant speed control unit is being performed, the pass bandwidth adjustment unit narrows the pass bandwidth. 6.   6. The vehicle according to claim 3, further comprising acceleration intention detection means for detecting an acceleration intention of the driver, wherein the pass bandwidth adjustment means widens the pass bandwidth when the acceleration intention is detected. The control device according to one. Continuously variable automatic transmission,
Means for detecting the gear ratio of the continuously variable automatic transmission;
With
The control according to any one of claims 3 to 6, wherein the passband width adjusting means widens the passband width as the speed ratio is larger, and narrows the passband width as the speed ratio is smaller. apparatus.   The said pass bandwidth adjustment means narrows a pass bandwidth after the state which should narrow the said pass bandwidth continues for a predetermined time, when narrowing the pass bandwidth once widened, The said pass bandwidth is narrowed. The control device according to any one of the above.

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