JP2001164964A - Sliding mode control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve control performance of a valve timing device, etc., of an internal combustion engine. SOLUTION: A valve timing control device performs feedback-control by sliding mode control. In such a device, when deviation between a target angle and an actual angle is within a specified range of control dead band causing no fluctuation of the target angle for a specified time (S1 to S3), a mean value of the deviation during the specified time is calculated (S4). When the mean value exceeds a threshold value on an angle-retarding side, a base duty ratio is corrected to an angle-advancing side (S7, S8). It is thus possible to correct dislocation of the base duty ratio in respect to a center value in the operation dead band.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a sliding mode control, and more particularly, to a sliding mode control device used for feedback-controlling a rotational phase of a camshaft with respect to a crankshaft of an internal combustion engine to a target value.

[0002]

2. Description of the Related Art Japanese Patent Laid-Open No. H10-14022 discloses a valve timing control apparatus having a structure in which the opening / closing timing of intake / exhaust valves is continuously variably controlled by changing the rotation phase of a camshaft with respect to a crankshaft. There is a vane type valve timing control device as disclosed in the publication.

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. 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 control valve. And the supply and discharge of oil is stopped.

[0005]

By the way, the feedback control method of the camshaft rotation phase includes P
ID control or the like is generally employed. In this case, the control amount is calculated using only a deviation (error amount) between the actual angle of the camshaft to be controlled and the target angle as a single variable. .

However, in order to perform the PID control with good responsiveness, it is desirable to set the feedback gain variably because the viscosity of the oil changes according to the oil temperature and the oil pressure. Not easy.

In hydraulic control, there is a large operation dead zone of a switching valve (spool valve) for switching between supply and discharge of oil.
In order to overcome the dead zone, dither control is performed by adding dither separately from the PID.
It is necessary to set the dither addition judgment finely, which makes the control complicated and takes up the capacity of the ROM and RAM.In addition, in order to reduce the variation of the dead zone width of each component and secure the control accuracy, The processing accuracy of the parts must be increased, and the processing cost has increased.

For this reason, transition from general PID control to sliding mode control, which is less affected by disturbance, is being studied. When the sliding mode control is applied to the one having the above-mentioned operation dead zone, as shown in FIG. 11, a basic control amount is set corresponding to the median value of the operation dead zone (for example, the control duty is 50%). When the feedback control is started by adding the feedback control amount that goes over the operation dead zone to the control amount, and the deviation between the target angle of the valve timing control device and the actual angle (actual angle) falls within a predetermined control dead zone. Then, the feedback control is stopped (feedback control amount = 0). That is, if feedback control is performed until the deviation between the target angle and the actual angle becomes zero, hunting occurs due to overshoot due to a response delay, so the control dead zone is provided. Thus, the amount of oil supplied to the hydraulic chamber of the valve timing control device between the time when the feedback control is stopped after entering the control dead zone and the time when the switching valve (spool valve) enters the operation dead zone and is completely closed. Thereby, the actual angle can be sufficiently converged to the target angle.

However, there is a case where the basic control amount does not correspond to the center position of the dead zone of the operation of the switching valve and shifts due to parts variation, deterioration over time, or the like. For example, in the case of deviation as shown in FIG. 12, when control is performed in the advance direction at the time of feedback control, the control duty required to get over the operation dead zone becomes large. Decrease, and responsiveness deteriorates.

Conversely, when the control is performed in the retard direction, the control duty required for overcoming the operation dead zone is reduced, and the duty provided for the advance operation is increased by that amount and exceeds the target value. Undershoots to the retard side.

Further, during the advance angle control, the opening of the switching valve when the feedback control is stopped after entering the control dead zone is small (even when a linear term proportional to the deviation is provided, the duty for the deviation is small). The amount of oil supplied to the hydraulic chamber after the feedback control is stopped also decreases since the feedback control is stopped, so that the steady-state deviation is retarded with respect to the target angle. Occurs. On the other hand, during the retard control, the opening degree of the switching valve during the feedback control is large, so that the effect of the undershoot on the retard side is left as it is, and a steady-state error is also generated on the retard side from the target angle.

[0012] On the other hand, in the case where it is shifted in the opposite direction from FIG.
Contrary to the above, during the advance control, overshoot occurs in the advance direction from the target angle, and during the retard control, the responsiveness deteriorates, and a steady-state error occurs on the advance side of the target angle.

In addition, if there is a variation in the size of the motion dead zone, for example, as shown in FIG. 13, when the motion dead zone is small, the nonlinear term of the feedback control amount set by the sliding mode control becomes small. It becomes relatively large and causes chattering. That is, since the opening degree of the switching valve when the feedback control is stopped after entering the control dead zone is large, the amount of oil supplied to the hydraulic chamber increases until the switching valve closes, and the undershoot to the retard side occurs. In the case of deviating from the control dead zone, feedback control is performed on the advance angle side, and chattering that overshoots on the advance angle side is generated.

Conversely, as shown in FIG. 14, when the motion dead zone is large, the nonlinear term becomes relatively small, and the nonlinear term alone cannot overcome the motion dead zone at the time of feedback control, resulting in poor response. Getting worse.

The present invention has been made in view of such circumstances, and in a sliding mode control of a control object having an operation dead zone, the effects of variations in parts and changes over time can be avoided by learning, and high-precision control is performed. The purpose is to be able to.

[0016]

For this reason, the present invention according to claim 1 provides control by adding a feedback control amount that goes over the operation dead zone to a basic control amount set corresponding to the median value of the operation dead zone. Start feedback control to the target value of the target, stop the feedback control when the deviation between the target value and the actual value enters a control dead zone within a predetermined range, and calculate the feedback control amount by sliding mode control. In the sliding mode control device, the basic control amount is corrected based on a steady-state deviation between a target value and an actual value when the control enters the control dead zone.

According to the first aspect of the present invention, as described above, if a deviation occurs in the correspondence between the basic control amount and the median value of the operation dead zone, a steady-state deviation between the target value and the actual value will occur in the controlled object. appear.

Therefore, the basic control amount is corrected based on the steady-state error so as to correspond to the median value of the operation dead zone, thereby suppressing deterioration of the response, occurrence of overshoot and undershoot, and occurrence of the steady-state error. be able to.

Further, according to a second aspect of the present invention, the feedback to the target value of the control object is performed by adding a feedback control amount that goes over the operation dead zone to the basic control amount set corresponding to the median value of the operation dead zone. Start the control, and when the deviation between the target value and the actual value enters a control dead zone within a predetermined range, stop the feedback control,
In the sliding mode control device that calculates the feedback control amount by the sliding mode control, when the control enters the control dead zone, the feedback control is performed using only the nonlinear term while changing the gain of the nonlinear term of the feedback control amount. The gain is adjusted so that the deviation between the target value and the actual value falls within a set range.

According to the second aspect of the present invention, when entering the control dead zone, while performing the feedback control using only the nonlinear term while changing the gain of the nonlinear term of the feedback control amount, the target value and the actual value are controlled. By adjusting the gain so that the deviation from the value falls within a set range, the nonlinear term is adjusted to a size that slightly exceeds the operation dead zone, thereby preventing chattering and ensuring responsiveness. .

Further, the invention according to claim 3 is characterized in that the switching function S in the sliding mode control is calculated as a function of a deviation between a target value to be controlled and an actual value.

According to the third aspect of the present invention, since the gain of the nonlinear term is switched according to the state of the deviation, the gain converges smoothly to the target value.

The invention according to claim 4 is characterized in that the switching function S is calculated by the following equation. S = γ × PERR + d (PERR) / dt γ: Slope PERR: Deviation between target value and actual value of control target d (PERR) / dt: Differential value of the deviation According to the invention according to claim 4, the switching coefficient By giving a differential value d (PERR) / dt of the deviation in addition to the deviation PERR between the target value and the actual value of the control target as S, the sliding mode along the switching line is made smoother. be able to.

The invention according to claim 5 is characterized in that the switching function S is calculated by the following equation. S = γ × PERR + d (NOW) / dt γ: Slope PERR: Deviation between the target value and the actual value of the controlled object d (NOW) / dt: Actual speed of the controlled object Even if an actual speed, which is a differential value of the position of the control target, is given instead of the differential value d (PERR) / dt of the deviation PERR amount in the term 4, the sliding mode along the switching line is similarly smooth. It can be.

Further, the invention according to claim 6 is characterized in that the control amount U in the sliding mode control is calculated by the following equation.

U = c × PERR + d × ｛d (NOW) /
dt] −K [S / (| S | + δ)] d (NOW) / dt: actual speed of the controlled object c, d: constant δ: chattering prevention coefficient According to the invention according to claim 6, the c × PERR + d ×
[D (NOW) / dt] linear term control quantity UL
Has a role of adjusting the speed at which the state of the control system approaches the switching line (S = 0), and the nonlinear term control amount UNL represented by -K [(S / (| S | + δ)] It has a role to generate a sliding mode along a line.

According to a seventh aspect of the present invention, the control target is a rotational phase of a camshaft with respect to a crankshaft of the internal combustion engine, and the rotational phase is feedback-controlled to a target value, thereby controlling the intake and exhaust valves. The opening and closing timing is continuously variably controlled.

According to the seventh aspect of the present invention, in a configuration in which the valve timing is continuously changed by changing the rotation phase of the camshaft with respect to the crankshaft, the sliding mode control is applied to the valve timing ( (Substantially controlled object) can be controlled with high accuracy as a target.

The invention according to claim 8 is characterized in that the rotation phase of the camshaft is controlled by selectively controlling the supply and discharge of oil to and from a hydraulic actuator that is hydraulically controlled by a switching valve. And

According to the eighth aspect of the invention, by selectively controlling the supply and discharge of oil to and from the hydraulic actuator by the switching valve, the driving direction of the hydraulic actuator is switched and the amount of oil to the hydraulic chamber is adjusted. By doing so, the rotational phase of the camshaft is continuously variably controlled.

The effect of the present invention can be sufficiently exhibited by applying the present invention to a hydraulically controlled object having an operation dead zone by the switching valve.

[0032]

Embodiments of the present invention will be described below. FIGS. 1 to 6 show a mechanical portion of a valve timing control device for an internal combustion engine that performs feedback control using sliding mode control in the present embodiment, and shows a mechanism applied to an intake valve side.

The valve timing control device shown in the figure is a cam sprocket 1 (timing sprocket) driven to rotate via a timing chain by a crankshaft (not shown) of an engine, 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 to rotatably house the rotating member 3. 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 part 5 has a substantially annular shape, and each of the small-diameter bolts 9 is screwed at an equidistant position 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 open front and rear ends, 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 is provided with 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 fitted and held in the fitting groove 12 of the rotary member 5 on the rear end surface, 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 engagement 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 on one side inside the camshaft 2 and a substantially L-shaped bent portion inside the sleeve 25. 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).

More 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 for 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, and the supply port 55 is provided on both sides of the supply port 55. 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 third and fourth ports 58 and 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 operation state (load, rotation) based on signals from a rotation sensor 101 for detecting the engine 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 current supplied 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 hydraulic chamber 33 becomes high and the internal pressure of the advance hydraulic chamber 32 becomes low, and the rotating member 3 rotates in one direction at a 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. A feedback correction value UDTY for matching a value (target advance value) with
The result of the addition of the base duty ratio BASEDTY and the feedback correction amount UDTY is used as a final duty ratio VTCTY, and a control signal of the duty ratio VTCTY is output to the electromagnetic actuator 54. The base duty ratio BASEDTY is
This is a value corresponding to the basic control amount of the present invention.
5, the third port 58 and the fourth port 59 are both closed,
The center value (for example, 50) of the duty ratio range (operation dead zone) in which oil is not supplied or discharged in any of the hydraulic chambers 32 and 33
%).

That is, when it is necessary to change the relative rotation position (rotational phase) in the retard direction, the duty ratio is reduced by the feedback correction 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 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.

FIG. 7 is a block diagram showing a state of duty control of the electromagnetic actuator 54 by the controller 48 to which the sliding mode control designed as described above is applied.

The deviation PERR between the VTC target angle VTCTRG and the VTC actual angle VTCNOW is calculated, and the deviation PERR is calculated.
A proportional control amount Up obtained by multiplying RR by a P gain c, and VT
VTC actual speed U which is a differential value of C actual angle VTCNOW
The linear term control amount UL is calculated by adding the speed control amount UN obtained by multiplying n ′ by the speed gain d.

Further, a value obtained by multiplying the deviation PERR by the slope γ and a differential value d (PERR) / dt of the deviation PERR are added to calculate a switching function S, and a smoothing function using the switching function S is calculated. The nonlinear term control amount UNL is calculated as −kS / (| S | + δ).

The linear term control amount UL is controlled by a control system (VT
C) has a role of adjusting the speed at which the state of C) approaches the switching line (S = 0), and the nonlinear term controlled variable UNL has a role of generating a sliding mode along the switching line.

Then, the control amount (feedback correction amount) UDTY is calculated by adding the linear term control amount UL and the nonlinear term control amount UNL, and the feedback correction amount U is calculated.
DTY is added to the base duty ratio BASEDTY corresponding to the center position of the operation dead zone, and the addition result is output as a 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 (see FIG. 8). It is hardly affected by disturbances such as oil pressure, and control with high robustness can be performed.

The control according to the present invention is executed as follows in addition to the basic sliding mode control.
FIG. 9 shows a flowchart of control for learning and correcting the deviation of the base duty ratio BASEDTTY from the center position of the operation dead zone.

In step 1, it is determined whether or not the target angle VTCRG has changed. When it is determined that the target angle VTCRG has changed and the feedback control has been performed, the process proceeds to step 2.

In step 2, after the deviation PERR between the target angle VTCTRG and the actual angle VTCNOW enters a control dead zone within a predetermined range (for example, within ± 3 °), feedback control is stopped for a predetermined time x. Angle VT
It is determined whether the CTRG has not changed, that is, whether the CTRG is in a steady state.

If it is determined that the vehicle is in the steady state, the process proceeds to step 3, where it is determined whether the base duty ratio BASEDTY has been changed (corrected by learning in steps 6 and 8 described later) and not changed. If so, proceed to step 4.

In step 4, an average value (steady-state error) PERRAV of the error PERR at the predetermined time x is calculated. In step 5, it is determined whether or not the average value PERRAV of the deviation PERR exceeds a positive threshold value P0. If it is determined that the average value PERRAV exceeds the threshold value P0, a non-negligible steady-state deviation is generated on the retard side. Duty ratio BA
It is determined that SEDTY is shifted to the retard side with respect to the median value of the operation dead zone, and the routine proceeds to step 6, where a predetermined amount of positive correction duty HD is added to the base duty ratio BASEDTY.

If it is determined in step 5 that the average value PERRAV of the deviation PERR does not exceed the positive threshold value P0, the process proceeds to step 7, and the average value PERRAV of the deviation PERR falls below the negative threshold value -P0. Is not determined. And if it is determined to be below,
A non-negligible steady-state error occurs on the advance side, and as a result, it is determined that the base duty ratio BASEDTY is shifted to the advance side with respect to the median value of the operation dead zone. HD is subtracted.

In this manner, the base duty ratio BASEDTY is increased / decreased so that the average value PERRAV of the deviations PERR calculated in step 4 falls within the range of the threshold value ± P0. As a result, the steady-state deviation can be reduced to a negligible size, and the deviation between the base duty ratio BASEDTY and the dead zone can be sufficiently reduced. It is possible to suppress and quickly converge on the target angle.

Further, the second control according to the present invention is executed. FIG. 10 shows a flowchart of control for learning and correcting the gain K of the nonlinear term with respect to the variation in the size of the motion dead zone.

After judging a change in the target angle VTCTRG in Step 11, it is judged in Step 12 whether or not a control dead zone has been entered. When it is determined that the vehicle has entered the control dead zone,
Proceeding to step 13, the gain K of the nonlinear term control amount UNL is increased by a predetermined amount DK1, and only the increased nonlinear term control amount UNL is used as the feedback correction amount UDTY (the linear term control amount UL = 0). Perform

In step 14, after the above, a predetermined time x
The average value PEERAV2 of the deviation PEER in is calculated. In step 15, the average value PE of the deviation PEER is calculated.
It is determined whether or not ERAV2 exceeds a threshold value P1.

If it is determined that the value exceeds the threshold value P1, the routine proceeds to step 16, where the gain K is reduced by a predetermined amount DK2. If it is determined that the difference does not exceed the threshold value P1, the process jumps to step S17.

In step 17, the VTC target angle VTC
It is determined whether or not TRG has changed. If it is determined that TRG has not changed, the process returns to step 13 again to increase the gain K and perform feedback control.

In this manner, during the steady state in the control dead zone, the gain K is adjusted so that the average value PEERAAV2 of the deviation PEER is maintained near the threshold value P1 while increasing or decreasing the gain K.

Therefore, the nonlinear term control amount UNL can be adjusted to an appropriate level by adjusting the gain K (learning correction) with respect to the variation of the motion dead zone, and as a result, the nonlinear term control amount can be adjusted. The operation dead zone can be overcome only by the UNL, and chattering can be prevented from occurring.

[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 control block diagram of the valve timing control mechanism.

FIG. 8 is a time chart showing how the valve timing control mechanism converges to a target angle during sliding mode control.

FIG. 9 is a flowchart showing control for learning correction of the base duty ratio BASEDTY of the valve timing control mechanism.

FIG. 10 is a flowchart showing control for learning and correcting the gain of a nonlinear term of the valve timing control mechanism.

FIG. 11 is a diagram showing a state where a base duty ratio BASEDTY accurately corresponds to a median value in an operation dead zone.

FIG. 12 is a diagram showing a state in which a base duty ratio BASEDTY is shifted to a retard side with respect to a central value in an operation dead zone.

FIG. 13 is a diagram illustrating a state in which an operation dead zone is small with respect to a nonlinear term.

FIG. 14 is a diagram showing a state when the operation dead zone is large with respect to a nonlinear term.

[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 F term (reference) 3G016 AA06 AA19 BA23 BA38 CA13 CA24 CA33 CA36 CA48 CA59 DA06 DA22 GA00 3G084 BA03 BA23 DA05 DA08 EA07 EA12 EB14 EB15 EB16 EC04 EC06 FA07 FA33 FA38 FA39 3G092 AA11 DA01 DG04 DG04 DG04 EA11 EA13 EA19 EA22 EA28 EA29 EB02 EB03 EC02 EC03 EC05 EC06 EC08 FA09 FA36 FA48 HA01Z HA13Z HE01Z 3G301 HA19 JA03 JA07 LA07 LC08 NA01 NA09 NC04 ND05 ND15 ND41 NE23 NE25 PA01Z PE01Z PE03Z PE04Z

Claims (8)

[Claims] 1. A feedback control to a target value of a controlled object is started by adding a feedback control amount that goes over the operation dead zone to a basic control amount set corresponding to a median value of the operation dead zone. In a sliding mode control device that stops the feedback control when the deviation from the actual value enters a control dead zone within a predetermined range and calculates the feedback control amount by sliding mode control, a target when entering the control dead zone. A sliding mode control device, wherein the basic control amount is corrected based on a steady-state deviation between a value and an actual value. 2. A feedback control to a target value of a controlled object is started by adding a feedback control amount over the operation dead zone to a basic control amount set corresponding to a median value of the operation dead zone. When the deviation from the actual value enters a control dead zone within a predetermined range, the feedback control is stopped, and in a sliding mode control device that calculates the feedback control amount by sliding mode control, when entering the control dead zone, While performing the feedback control using only the nonlinear term while changing the gain of the nonlinear term of the feedback control amount, the gain is adjusted such that the deviation between the target value and the actual value is within a set range. Sliding mode control device. 3. The sliding mode according to claim 1, wherein the switching function S in the sliding mode control is calculated as a function of a deviation between a target value to be controlled and an actual value. Control device. 4. The sliding mode control device according to claim 3, wherein the switching function S is calculated by the following equation. S = γ × PERR + d (PERR) / dt γ: Slope PERR: Deviation between target value and actual value of control target d (PERR) / dt: Derivative value of the above deviation 5. The sliding mode control device according to claim 3, wherein the switching function S is calculated by the following equation. S = γ × PERR + d (NOW) / dt γ: Slope PERR: Deviation between target value and actual value of control target d (NOW) / dt: Change speed of control target 6. The sliding mode control device according to claim 4, wherein the feedback control amount U is calculated by the following equation. U = c × PERR + d × ｛d (NOW) / dt] -K
[(S / | S | + δ)] d (NOW) / dt: change speed of controlled object c, d: constant δ: chattering prevention coefficient 7. The control target is a rotational phase of a camshaft with respect to a crankshaft of an internal combustion engine, and the rotational phase is feedback-controlled to a target value, so that the opening / closing timing of an intake / exhaust valve is continuously variably controlled. The sliding mode control device according to any one of claims 1 to 6, wherein: 8. The control method according to claim 7, wherein the rotational phase of the camshaft is controlled by selectively controlling the supply and discharge of oil to and from a hydraulic actuator that is hydraulically controlled by a switching valve. Sliding mode control device.