JP2001132483A - Sliding mode control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To carry out a sliding mode control having reduced the effect by an operation dead zone, in the valve timing control device of an internal combustion engine by a hydraulic control. SOLUTION: The feedback correction amount (UDTY) of an electromagnetic selector valve for hydraulic controlling the rotary phase against the crankshaft of a camshaft is feedback controlled by a sliding mode control, while adding and calculating the linear term control amount including the term proportional to the error amount which is the deviation between the target opening of the rotary phase and a real angle to the non-linear term control amount using a changeover function S. Thereby, as the linear term control amount is given, even at the entering time to the operation dead zone, a control system is introduced to the changeover line (S=0) with a proper moving speed from an initial state and is converged to the target angle while sliding on the changeover line.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an apparatus for controlling a controlled object having an operation dead zone for a control amount of a hydraulic control system or the like in a sliding mode.

[0002]

2. Description of the Related Art Japanese Patent Application Laid-Open No. 10-14022 discloses a valve timing control apparatus having a structure in which the rotational phase of a camshaft with respect to a crankshaft of an internal combustion engine is continuously variably controlled by hydraulic control using a switching valve. There is such a vane type valve timing control device.

In this apparatus, a concave portion is formed in an inner peripheral surface of a cylindrical housing fixed to a cam sprocket, and a wing portion (vane) of an impeller fixed to a camshaft is accommodated in the concave portion. The camshaft is configured to be rotatable relative to the cam sprocket within a range in which the wing can move within the recess.

The wings supply and discharge oil relatively to a pair of hydraulic chambers formed by partitioning the recess in front and rear in the rotation direction, so that the wings are positioned at an intermediate position of the recess. And the rotary phase is continuously variable controlled.When the hydraulic pressure of the pair of hydraulic chambers is adjusted to the hydraulic pressure at which the target rotational phase is obtained, the hydraulic passage is closed by the switching valve. And the supply and discharge of oil is stopped.

[0005]

In the above-described hydraulic control system, the hydraulic passage for supplying and discharging oil from the hydraulic chamber is closed by a switching valve, and the hydraulic chamber is normally operated by a reaction force from the intake and exhaust valves. In order to prevent oil leakage from the valve, the closing margin of the valve body of the switching valve (spool valve) is set large. For this reason, the amount of oil in the hydraulic chamber and, consequently, the rotational phase of the camshaft to be controlled have an operation dead zone with respect to the control amount of the switching valve (duty ratio or the like in the electromagnetic drive type).

As a method of controlling the camshaft rotation phase, PID control or the like is generally adopted.
It is difficult to perform feedback control with good response to the dead zone only by the D control. For this reason, there is a type in which dither control is performed by adding a dither amount separately from the PID. However, it is necessary to make a determination of addition of the dither based on an error amount, and the control becomes complicated.
In order to secure the control accuracy by reducing the variation of the dead zone width of each component, the processing accuracy of the component must be increased, and the processing cost is increased.

In order to execute the PID control with good responsiveness, the viscosity of the oil changes according to the oil temperature and the oil pressure. Therefore, it is desirable to set the feedback gain variably. Not easy.

On the other hand, in recent years, a sliding mode control has been attracting attention as a feedback control having a small effect on disturbance and a high robustness. Therefore, it is conceivable to apply the sliding mode control to the hydraulically controlled valve timing control device, and when the sliding mode control is designed according to a textbook, the effects of disturbances such as changes in the oil temperature and oil pressure are suppressed. However, it has been found that it does not function effectively for the dead zone and does not replace or assist the dither control.

SUMMARY OF THE INVENTION The present invention has been made in view of such a conventional problem, and suppresses a decrease in responsiveness due to the dead zone for a controlled object having an operation dead zone for a control amount of a hydraulic control system or the like. It is an object of the present invention to execute a sliding mode control with high robustness.

[0010]

According to the present invention, there is provided an apparatus for controlling a controlled object having an operation dead zone with respect to a control amount in a sliding mode, wherein a linear term of the control amount is controlled. It is characterized in that it is set as a function of the deviation between the target position of the object and the actual position.

According to the first aspect of the present invention, the linear term of the control amount in the sliding mode control is set as a function of the deviation between the target position of the controlled object and the actual position.

As a result, even when the actual position of the controlled object does not change after entering the operation dead zone, there is a deviation (≠ 0) between the target position and the actual position, so that it is set as a function of the deviation. By the linear term, the moving speed to the switching line for switching the feedback gain is appropriately given,
Thus, it is possible to converge on the target position with good responsiveness.

Further, the invention according to claim 2 is characterized in that the controlled object is a hydraulic control system. According to the invention of claim 2, the sliding mode control according to the present invention is applied to the hydraulic control system.

Thus, in the hydraulic control system having a large operation dead zone due to the switching valve and the like, control with good responsiveness in which the influence of the dead zone is suppressed is executed.
The invention according to a third aspect is configured such that the controlled object is configured to continuously and variably control a rotation phase of a camshaft with respect to a crankshaft by hydraulic control, and to supply and discharge oil to and from the hydraulic actuator that is hydraulically controlled. Is selectively controlled by a switching valve to control the valve timing of the internal combustion engine.

According to the third aspect of the present invention, the sliding mode control according to the present invention is applied to the valve timing control device for an internal combustion engine having the above configuration.

Thus, in the valve timing control device of the hydraulic control type internal combustion engine having a large dead zone due to the switching valve, control with good responsiveness in which the influence of the dead zone is suppressed is executed.

According to a fourth aspect of the present invention, the linear term of the control amount is divided into a term proportional to the deviation between the target position and the actual position of the controlled object, and a term proportional to the operating speed of the controlled object. It is characterized by adding and setting.

According to the present invention, the linear term of the control amount in the sliding mode control is a term proportional to the deviation between the target position of the controlled object and the actual position, and a term proportional to the operating speed of the controlled object. Is set by adding

[0019] With this, a function is provided that allows the moving speed to the switching line to be appropriately given even when entering the operation dead zone by means of a term proportional to the deviation between the target position of the controlled object and the actual position. When deviating from the dead zone, a function of adjusting the moving speed to the switching line by a term proportional to the operating speed of the controlled object is added, so that the moving speed is adjusted to a more appropriate one, and the responsiveness is further improved.

Further, the invention according to claim 5 is characterized in that the linear term of the control amount is set only by a term proportional to the deviation between the target position of the controlled object and the actual position.

According to the fifth aspect of the invention, the linear term of the control amount in the sliding mode control is set only by the term proportional to the deviation between the target position of the controlled object and the actual position.

Even when the linear term is set only by a term proportional to the deviation between the target position of the controlled object and the actual position,
Since it has a function of appropriately giving the moving speed to the switching line not only when the vehicle enters the operation dead zone but also when the vehicle deviates from the operation dead zone, the sliding mode control can be realized. By omitting the term proportional to the operating speed of the controlled object, control (calculation)
Can be simplified and the program capacity can be saved.

[0023]

Embodiments of the present invention will be described below. 1 to 6 show a mechanical portion of a valve timing control device for an internal combustion engine according to the embodiment, and show a mechanism applied to an intake valve side.

The valve timing control device shown in FIG. 1 is a cam sprocket 1 (timing sprocket) that is rotationally driven by a crankshaft (not shown) of an engine via a timing chain, and is rotatable relative to the cam sprocket 1. A camshaft 2 provided, a rotating member 3 fixed to an end of the camshaft 2 and rotatably housed in the cam sprocket 1, and rotating the rotating member 3 relative to the cam sprocket 1. And a lock mechanism 10 for selectively locking a relative rotation position between the cam sprocket 1 and the rotating member 3 at a predetermined position.

The cam sprocket 1 has a rotating portion 5 having teeth 5a on its outer periphery with which a timing chain (or a timing belt) meshes, and a rotating member 3 disposed in front of the rotating portion 5 so as to be rotatable. A housing 6, a disk-shaped front cover 7 serving as a lid for closing a front end opening of the housing 6, and a substantially disk disposed between the housing 6 and the rotating part 5 to close a rear end of the housing 6. The rotating part 5, the housing 6, the front cover 7, and the rear cover 8 are integrally connected by four small-diameter bolts 9 in the axial direction.

The rotating portion 5 has a substantially annular shape, and each small-diameter bolt 9 is screwed at an equal interval of about 90 ° in the circumferential direction.
One of the female screw holes 5b is formed to penetrate in the front-rear direction.
A fitting hole 11 having a stepped diameter is formed to penetrate. Further, a disc-shaped fitting groove 12 into which the rear cover 8 is fitted is formed in the front end face.

The housing 6 has a cylindrical shape with front and rear ends formed with openings, and four partition walls 13 are protruded from the inner peripheral surface at 90 ° in the circumferential direction. The partition 13 has a trapezoidal cross-section, and each has a housing 6.
Are provided along the axial direction of the
And four bolt insertion holes 14 through which the small-diameter bolt 9 is inserted are formed in the base end side in the axial direction. Further, a U-shaped sealing member 15 and a leaf spring 16 for pressing the sealing member 15 inward are fitted into a holding groove 13a which is cut out along the axial direction at the center position of the inner end surface of each partition 13. Is held.

Further, the front cover 7 has a relatively large-diameter bolt insertion hole 17 at the center.
Four bolt holes 18 are formed in the housing 6 at positions corresponding to the bolt insertion holes 14.

The rear cover 8 has a disk portion 8a which is fitted and held in the fitting groove 12 of the rotating member 5 on the rear end face, and a small-diameter annular portion 2 of the sleeve 25 at the center.
An insertion hole 8c into which 5a is inserted is formed, and four bolt holes 19 are also formed at positions corresponding to the bolt insertion holes 14.

The camshaft 2 comprises a cylinder head 2
A cam (not shown) for rotatably supporting an intake valve via a valve lifter is integrally provided at a predetermined position on the outer peripheral surface at a top end of the second end via a cam bearing. The part is provided integrally with a flange part 24.

The rotating member 3 is fixed to the front end of the camshaft 2 by a fixing bolt 26 inserted from the axial direction through the sleeve 25 whose front and rear portions are fitted into the flange portion 24 and the fitting hole 11, respectively. An annular base 27 having a bolt insertion hole 27a through which the fixing bolt 26 is inserted at the center, and four vanes 28a, 28b, 2 integrally provided at a position at 90 ° in the circumferential direction of the outer peripheral surface of the base 27.
8c and 28d.

The first to fourth vanes 28a to 28d are:
Each cross section has a substantially inverted trapezoidal shape, is disposed in a concave portion between the partition portions 13, and separates the concave portion before and after in the rotational direction, between the both sides of the vanes 28a to 28d and both side surfaces of each partition portion 13. In addition, an advance hydraulic chamber 32 and a retard hydraulic chamber 33 are configured. Further, the inner peripheral surface 6 of the housing 6 is inserted into a holding groove 29 which is notched in the axial direction at the center of the outer peripheral surface of each of the vanes 28a to 28d.
and a U-shaped seal member 30 slidably contacting the
The leaf springs 31 that press 0 outward are respectively fitted and held.

The lock mechanism 10 includes an engagement groove 20 formed at a predetermined position on the outer peripheral side of the fitting groove 12 of the rotating portion 5,
An engagement hole 21 formed through a predetermined position of the rear cover 8 corresponding to the engagement groove 20 and having a tapered inner peripheral surface;
A sliding hole 35 penetratingly formed along the inner axial direction at a substantially central position of the one vane 28 corresponding to the engagement hole 21.
A lock pin 34 slidably provided in the slide hole 35 of the one vane 28; and a coil spring 3 which is a spring member elastically mounted on the rear end side of the lock pin 34.
9 and a pressure receiving chamber 40 formed between the lock pin 34 and the sliding hole 35.

The lock pin 34 is formed with a middle-diameter main body 34a on the center side, an engaging portion 34b formed in a tapered conical shape on the front end side of the main body 34a, and a rear end side of the main body 34a. And a stopper portion 34c having a large-diameter stepped portion, and is engaged by the spring force of the coil spring 39 elastically mounted between the bottom surface of the internal concave groove 34d of the stopper portion 34c and the inner end surface of the front cover 7. In the pressure receiving chamber 40 formed between the main body 34a and the stopper 34c and between the main body 34a and the stopper 34c and between the main body 34a and the inner peripheral surface of the sliding hole 35. It slides in the direction of coming out of the engagement hole 21 by the hydraulic pressure.
The pressure receiving chamber 40 communicates with the retard side hydraulic chamber 33 through a through hole 36 formed in a side portion of the vane 28. The engaging portion 34b of the lock pin 34 engages with the engaging hole 21 at the rotation position on the maximum retard side of the rotating member 3.

The hydraulic circuit 4 includes a first hydraulic passage 41 for supplying and discharging hydraulic pressure to and from the advance hydraulic chamber 32 and a second hydraulic passage 42 for supplying and discharging hydraulic pressure to the retard hydraulic chamber 33. The two hydraulic passages 41 and 42 have
The supply passage 43 and the drain passage 44 are connected to each other via an electromagnetic switching valve 45 for switching the passage. An oil pump 47 for pumping oil in an oil pan 46 is provided in the supply passage 43, while a downstream end of the drain passage 44 communicates with the oil pan 46.

The first hydraulic passage 41 has a first hydraulic passage 41 formed inside the cylinder head 22 and inside the axis of the camshaft 2.
A passage portion 41a, a first oil passage 41b branched and formed in the head portion 26a through the axial direction inside the fixing bolt 26 and communicating with the first passage portion 41a, a small-diameter outer peripheral surface of the head portion 26a, and a rotating member An oil chamber 41c formed between the inner peripheral surface of the bolt insertion hole 27a provided in the base 27 of the third rotating member 3 and communicating with the first oil passage 41b; It is composed of a chamber 41c and four branch passages 41d communicating with each advance-side hydraulic chamber 32.

On the other hand, the second hydraulic passage 42 is formed in the cylinder head 22 and one side inside the camshaft 2 and a second passage portion 42a formed in the sleeve 25 and bent in a substantially L-shape. A second oil passage 42b communicating with the second passage portion 42a, and four oil passage grooves 42 formed at the outer peripheral side edge of the fitting hole 11 of the rotating member 5 and communicating with the second oil passage 42b.
c and four oil holes 42d formed at about 90 ° in the circumferential direction of the rear cover 8 and communicating each oil passage groove 42c and the retard side hydraulic chamber 33.

The electromagnetic switching valve 45 controls the relative switching between the hydraulic passages 41 and 42, the supply passage 43 and the drain passages 44a and 44b by an internal spool valve body. The switching operation is performed by the control signal of (1).

Specifically, as shown in FIGS. 4 to 6, a cylindrical valve body 51 inserted and fixed in a holding hole 50 of a cylinder block 49 and a valve hole 52 in the valve body 51 are slid. The spool valve 53 is provided movably and switches a flow path, and includes a proportional solenoid type electromagnetic actuator 54 for operating the spool valve 53.

In the valve body 51, a supply port 55 communicating the downstream end of the supply passage 43 and the valve hole 52 is formed at a substantially central position of the peripheral wall so as to penetrate therethrough. First and second hydraulic passages 41 and 42
A first port 56 and a second port 57 which communicate the other end of the valve hole 52 with the valve hole 52 are respectively formed through. Further, both drain passages 44a and 44b and the valve hole 5 are provided at both ends of the peripheral wall.
Third and fourth ports 58, 59 communicating with the second port 2 are formed through.

The spool valve element 53 has a substantially cylindrical first valve part 60 for opening and closing the supply port 55 at the center of the small diameter shaft part, and has third and fourth ports 58, 5 at both ends.
9 has a substantially cylindrical second and third valve portions 61 and 62 for opening and closing the valve 9. Further, the spool valve element 53 is connected to the front end shaft 5.
A conical valve spring 63 elastically mounted between an umbrella portion 53b provided on one end edge of the valve 3a and a spring seat 51a provided on an inner peripheral wall on the front end side of the valve hole 52, to the right in the drawing, that is, the first valve portion 60. Urged in a direction to connect the supply port 55 with the second hydraulic passage 42.

The electromagnetic actuator 54 includes a core 6
4, a moving rod 65a that includes a moving plunger 65, a coil 66, a connector 67, and the like, and that presses an umbrella portion 53b of the spool valve body 53 is fixed to an end of the moving plunger 65.

The controller 48 detects the current operating state (load, rotation) based on signals from a rotation sensor 101 for detecting the engine rotation speed and an air flow meter 102 for detecting the intake air amount, and detects the crank angle sensor 1.
03 and a signal from the cam sensor 104, the relative rotation position between the cam sprocket 1 and the cam shaft 2, that is,
The rotational phase of the camshaft 2 with respect to the crankshaft is detected.

The controller 48 controls the amount of power to the electromagnetic actuator 54 based on a duty control signal. For example, a control signal (OF) having a duty ratio of 0% is sent from the controller 48 to the electromagnetic actuator 54.
When the F signal is output, the spool valve body 53
The position shown in FIG. 4, that is, the maximum rightward direction is moved by the spring force. Accordingly, the first valve portion 60 opens the open end 55a of the supply port 55 to communicate with the second port 57, and at the same time, the second valve portion 61 opens the open end of the third port 58, and the fourth The valve part 62 closes the fourth port 59. For this reason, the hydraulic oil pumped from the oil pump 47 is supplied to the supply port 55, the valve hole 52, the second port 57,
The hydraulic oil is supplied to the retard hydraulic chamber 33 through the hydraulic passage 42 and the hydraulic oil in the advance hydraulic chamber 32 is supplied to the first hydraulic passage 4.
The oil is discharged from the first drain passage 44a into the oil pan 46 through the first port 56, the valve hole 52, and the third port 58.

Accordingly, the internal pressure of the retard side hydraulic chamber 33 becomes high and the internal pressure of the advance side hydraulic chamber 32 becomes low, so that the rotating member 3 rotates in one direction at maximum through the vanes 28a to 28b. As a result, the cam sprocket 1 and the camshaft 2 relatively rotate to one side to change the phase. As a result, the opening timing of the intake valve is delayed, and the overlap with the exhaust valve is reduced.

On the other hand, when a control signal (ON signal) having a duty ratio of 100% is output from the controller 48 to the electromagnetic actuator 54, the spool valve body 53 moves leftward against the spring force of the valve spring 63 as shown in FIG. To the maximum, the third valve portion 61 closes the third port 58, and at the same time, the fourth valve portion 62 opens the fourth port 59,
The first valve section 60 makes the supply port 55 communicate with the first port 56. Therefore, the hydraulic oil is supplied to the advance hydraulic chamber 32 through the supply port 55, the first port 56, and the first hydraulic passage 41, and the hydraulic oil in the retard hydraulic chamber 33 is supplied to the second hydraulic chamber 32. Hydraulic passage 42, second port 57, fourth port 5
9. The oil is discharged to the oil pan 46 through the second drain passage 44b, and the pressure in the retard hydraulic chamber 33 becomes low.

For this reason, the rotating member 3 includes the vanes 28a to 28a.
Rotate in the other direction to the maximum through 28d,
The cam sprocket 1 and the camshaft 2 relatively rotate to the other side to change the phase. As a result, the opening timing of the intake valve is advanced (advanced), and the overlap with the exhaust valve is increased.

The controller 48 is configured such that the first valve portion 60 closes the supply port 55, the third valve portion 61 closes the third port 58, and the fourth valve portion 62 closes the fourth port 5.
9 is set to a base duty ratio BASEDTY while the crank angle sensor 10
3 and a relative rotational position (rotational phase) between the cam sprocket 1 and the camshaft 2 detected based on a signal from the cam sensor 104, and a target of the relative rotational position (rotational phase) set in accordance with an operation state. The feedback correction amount UDTY for matching the value (target advance value) is set by sliding mode control as described later, and the addition result of the base duty ratio BASEDTY and the feedback correction amount UDTY is used as the final duty ratio VT.
CDTY and a control signal of the duty ratio VTCDTY is output to the electromagnetic actuator 54.
Note that the base duty ratio BASEDTY is substantially the center value of the duty ratio range in which the supply port 55, the third port 58, and the fourth port 59 are all closed and oil is not supplied or discharged in any of the hydraulic chambers 32 and 33. (For example, 50%).

That is, when it is necessary to change the relative rotation position (rotation phase) in the retard direction, the duty ratio is reduced by the feedback correction amount UDTY, and the hydraulic oil pumped from the oil pump 47 is discharged. While being supplied to the retard side hydraulic chamber 33, the hydraulic oil in the advance side hydraulic chamber 32 is discharged into the oil pan 46, and conversely, the relative rotation position (rotation phase) is advanced. When it is necessary to change in the direction, the feedback correction U
The duty ratio is increased by DTY, and the hydraulic oil is supplied into the advance hydraulic chamber 32 and the hydraulic oil in the retard hydraulic chamber 33 is discharged to the oil pan 46. When the relative rotation position (rotational phase) is maintained in the current state, the absolute value of the feedback correction UDTY is reduced, so that the duty ratio is controlled to return to a duty ratio near the base duty ratio. By closing the 55, the third port 58, and the fourth port 59 (stopping the supply and discharge of the hydraulic pressure), the internal pressures of the hydraulic chambers 32 and 33 are controlled to be maintained.

Here, the feedback correction amount UDT
Y is calculated by the sliding mode control as follows. In the following, the detected relative rotational position (rotation phase) between the cam sprocket 1 and the camshaft 2 will be referred to as the actual angle of the valve timing control device (VTC).
The target value will be described as a VTC target angle.

1. Calculation of mathematical model In sliding mode control, the parameters of the controller are determined by the mathematical model to be controlled.
First, a mathematical model of the VTC is calculated.

The mathematical model can be obtained by, for example, establishing an equation of motion or obtaining from the identification of the system.
Here, system identification was used. As a result of system identification when input u (k): duty and output y (k): actual angle of VTC, the following transfer function was obtained.

G (s) = b / (s ² + a ₂ · s + a ₁ ) Simplification of transfer function The transfer function is simplified in order to simplify the configuration of the controller because the model obtained by system identification may be a higher-order model.

G (s) = b / [s (s + a ₂ )] (2.1) Calculation of State Equation From the obtained transfer function, the differential equation of VTC is given as follows. Where x: actual angle of VTC, u: input (duty)

[0055]

(Equation 1) The equation of state is

[0056]

(Equation 2) Substituting the differential equation (3.1) into (3.2) gives the following:

[0057]

(Equation 3) 4. Switching function design Sliding mode control depends on the state of the system.
Since the feedback gain is switched, this switching function S is set as follows.

[0058]

(Equation 4) The design of the switching function is very important because the sliding mode may not occur depending on the parameters of the switching function. There are mainly the following design methods.

Design Method Using Pole Assignment Method Design Method of Optimal Switching Hyperplane Design Method Using System Zeros Design Method of Hyperplane by Frequency Shaping α ₁ and α ₂ are obtained by applying the above design method. α ₁ : α ₂ =
When γ that satisfies γ: 1 is obtained, the result is as follows.

[0060]

(Equation 5) However, as described above, the switching function designed according to a normal textbook is a function of the actual position of the control target, that is, the function of the actual angle x of the VTC. It will be unsuitable.

First, with respect to the term γx, when the target angle of the VTC is other than 0 °, a positive value is always added and the value is irrelevant to the target angle and the actual angle. Does not converge.

Regarding the term dx / dt, when the solenoid-operated directional control valve is in the dead zone, the VTC does not operate. Therefore, the actual speed dx / dt does not change. Is bad.

In the design according to the textbook, it is recommended to add an integral term of the error amount. However, when the camshaft is set at the target angle, the integral term of the error amount is left as a non-zero value. And functions to prevent convergence to the target angle.

Therefore, the switching function S is set as a function of the following error amount.

[0065]

(Equation 6) Here, the switching function was designed using a design method using zeros of the system. The zeros of the system are (S,
This is a method of setting the zero of (A, B) in the left half plane on the complex plane (S: switching function, A, B: constant of equation (3.2)).

5. Calculation of Sliding Condition The simplest condition for establishing sliding is S · dS
/ Dt <0. The above condition is satisfied only when S decreases. S means that the error decreases and converges to the target value when the above condition is satisfied because the error and the differential value of the error are used as variables.

First, an expression required for the expansion of S is obtained. The control amount u is set as follows.

[0068]

(Equation 7) Substituting this into equation (3.1) gives

[0069]

(Equation 8) Next, the hat u is developed.

Since hat u is an input when sliding, S = dS / dt = 0.

[0071]

(Equation 9) If bu = hat u,

[0072]

(Equation 10) Consider a sliding condition S · dS / dt <0.

[0073]

[Equation 11] From equations (5.2) and (5.3),

[0074]

(Equation 12) Therefore, if k is a positive value, sliding is established.

6. Design of control amount calculation formula The control amount (feedback correction amount) u is calculated by the formulas (5.1) and (5.3)
Then, it becomes as follows.

[0076]

(Equation 13) Using equation (2.1) which simplifies the transfer function of VTC, the state equation is as follows.

[0077]

[Equation 14] Using the state equation of (6.2), equation (6.1) is as follows.

[0078]

(Equation 15) Here, if α = b ⁻¹ (a−γ) and k ′ = b ⁻¹ k,

[0079]

(Equation 16) This equation is an equation that guarantees that the actuator moves while sliding on the switching line S = 0.

However, in the equation of the control amount designed in accordance with the textbook (according to the textbook), the linear term α · dx / dt is not supplied or discharged when the oil is in the dead zone.
The operation speed dx / dt = 0 → the linear term = 0, and does not function effectively.

Therefore, in the present invention, the following processing is performed so that the linear term functions effectively even when it enters the dead zone. That is, β · S (β is a constant) is added to the above expression of the control amount u. Here, when sliding on the switching line S = 0, β · S ≒ 0, so adding β · S to the control amount u has no effect on sliding.

[0082]

[Equation 17] Here, if β ′ = βγ and α ′ = α + β,

[0083]

(Equation 18) This expression is

[0084]

[Equation 19] In the form of

As described above, as a result of performing the above-described addition processing, the linear term of the control amount includes the error amount (PERR) of the VTC. Since the moving speed to the switching line is appropriately given and good sliding on the switching line is secured, it is possible to converge on the target angle with good responsiveness.

Note that the coefficients c and d are determined by using a normal linear control system design (determined from responsiveness and stability). For example, c can be determined from the 90% response time and overshoot of the actual valve timing controller. If the coefficient d is too large, it does not converge to the target angle and hunting occurs. Therefore, the coefficient d is set to an appropriate value so as not to diverge.

K sets a positive value. However, if it is too large, hunting may occur. Therefore, a maximum value that does not cause hunting is set. 7. When the nonlinear term UnL = −k · S / | S | = −ksgn (S) is used in a digital controller, the sampling period cannot be made infinitely small. Wake up.

Therefore, chattering is reduced by using a saturation function, a smoothing function and the like. FIG. 7 illustrates these functions. Either one may be used, but the smoothing function is easy to use because the operation formula is simpler than the saturation function (there is no conditional branch).

FIG. 8 is a block diagram showing the duty control of the electromagnetic actuator 54 by the controller 48 to which the sliding mode control designed as described above is applied.

[0090] calculating a VTC target angle VTCTRG and VTC error amount PERR is the deviation between the actual angle VTCNOW, a proportional part control quantity U _P multiplied by the P component gain c to the error amount PERR, differentiation of the VTC actual angle VTCNOW calculating the linear term control amount U _L by adding a a VTC actual speed U _N speed control amount multiplied by a velocity gain d to U _N 'value.

Further, a value obtained by multiplying the error amount PERR by the slope γ is calculated by calculating a differential value d (PERR) / differential value of the error amount PERR.
dt is added to calculate a switching function S, and the switching function S
Smoothing function was used -kS (| S | + δ) for calculating a non-linear term control amount U _NL as.

The linear term control amount _UL is controlled by a control system (VT
The role of adjusting the speed of bringing the state of C) closer to the switching line (S = 0) is adjusted, and the nonlinear term control amount _UNL is responsible for generating a sliding mode along the switching line.

[0093] Then, the linear term control amount U _L, by adding the non-linear term control amount U _NL, calculates a control amount (feedback correction amount) UDTY, the feedback correction amount UD
TY is added to the base duty ratio BASEDTY corresponding to the dead zone neutral position, and the result of the addition is output as the final duty ratio VTCTY.

As described above, the feedback correction amount is calculated by the sliding control, and the feedback gain is switched so as to guide the state of the control system on a preset switching line. It is hard to be affected and it is possible to perform control with high robustness (see FIG. 9).

In particular, by setting the linear term of the control amount as a function of the error amount, complicated dither control for overcoming the dead band of the switching valve is unnecessary or the dependency ratio can be reduced to simplify the matching. , ROM or R
AM capacity can also be saved.

Further, since the influence of the dead zone is reduced,
The dimensional tolerance of the dead band width of the switching valve is relaxed, and the processing cost can be reduced. Further, in the above-described embodiment, since the linear term is set by providing a term proportional to the operation speed of the VTC in addition to the term proportional to the error amount, the term proportional to the operation speed is used when deviating from the operation dead zone. In addition to the function of adjusting the moving speed to the switching line, the speed is adjusted to a more appropriate moving speed, and the responsiveness is further improved.

However, even when the linear term of the control amount is set only by a term proportional to the error amount, the moving speed to the switching line is not only when the operation enters the operation dead zone but also when the operation departs from the operation dead zone. Since it has a function given moderately, it is possible to realize sliding mode control,
In this case, by omitting the term proportional to the operation speed, the control (calculation) is simplified and the program capacity can be saved (see FIG. 10).

The present invention is not limited to the VTC using the vane type hydraulic actuator. For example, a linear type hydraulic actuator is used to convert a linear motion into a rotary motion to vary the rotational phase of a camshaft. VTC like
Needless to say, the present invention is not limited to the hydraulic control type as long as the control target has an operation dead zone.

[Brief description of the drawings]

FIG. 1 is a sectional view showing a valve timing control mechanism according to an embodiment.

FIG. 2 is a sectional view taken along line BB of FIG.

FIG. 3 is an exploded perspective view of the valve timing control mechanism.

FIG. 4 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.

FIG. 5 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.

FIG. 6 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.

FIG. 7 is a diagram showing a form of a function used for a nonlinear term control amount of the sliding mode control.

FIG. 8 is a control block diagram of the valve timing control mechanism.

FIG. 9 is a time chart showing how the valve timing control mechanism converges to a target angle during sliding mode control.

FIG. 10 is a control block diagram of a valve timing control mechanism according to another embodiment.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 2 ... Camshaft 4 ... Hydraulic circuit 32 ... Advance side hydraulic chamber 33 ... Tilt side hydraulic chamber 45 ... Electromagnetic switching valve 47 ... Oil pump 53 ... Spool valve element 101 ... Rotation sensor 102 ... Air flow meter 103 ... Crank angle sensor 104 … Cam sensor

Continued on the front page (51) Int.Cl. ⁷ Identification code FI Theme coat II (Reference) 9A001 F-term (Reference) 3G016 AA06 AA19 BA23 BA38 CA04 CA13 CA15 CA17 CA21 CA24 CA27 CA36 CA48 CA51 CA59 DA06 DA22 GA00 3G084 BA23 DA05 EA12 EB12 FA07 FA33 FA38 3G092 AA11 DA09 DF09 DG05 EA19 EC01 FA06 HA01Z HE01Z HE03Z 3G301 HA19 JA03 LA07 MA18 ND02 NE25 PA01A PE01A PE03A 5H004 GA17 GA35 GB12 HA07 HB07 HB08 KA22 KA65 KA74 KB02 KB03 LA13 K12 K03 K06 LA12 KK14 KK32 KK54

Claims (5)

[Claims] An apparatus for controlling a controlled object having an operation dead zone with respect to a controlled variable in a sliding mode, wherein a linear term of the controlled variable is set as a function of a deviation between a target position and an actual position of the controlled object. A sliding mode control device characterized in that: 2. The sliding mode control device according to claim 1, wherein the control target is a hydraulic control system. 3. The control target is configured to continuously and variably control the rotation phase of a camshaft with respect to a crankshaft by hydraulic control, and to select oil supply / discharge to a hydraulic actuator that is hydraulically controlled by a switching valve. 3. The sliding mode control device according to claim 2, wherein the valve timing control device is a valve timing control device for an internal combustion engine configured to perform control by performing dynamic control. 4. The method according to claim 1, wherein the linear term of the controlled variable is set by adding a term proportional to a deviation between a target position and an actual position of the controlled object and a term proportional to the operating speed of the controlled object. Sliding mode control device. 5. The sliding mode control device according to claim 1, wherein the linear term of the control amount is set only by a term proportional to the deviation between the target position of the controlled object and the actual position.