JP2007292001A - Control device for apparatus of hydraulic drive vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide both of responsiveness in operation position change control and holding stability at an operation position, in an apparatus for a hydraulic drive vehicle. <P>SOLUTION: A hydraulic drive type variable valve timing mechanism for switching advance and delay angle control direction by switching a hydraulic path by a hydraulic control valve has a structure for improving holding stability by improving control responsiveness by eliminating a dead zone of the hydraulic control valve and making an operation amount/oil flow rate characteristics linear, while applying a vibration component calculated in a vibration component calculation block to a feedback correction calculated in a feedback compensator block, and converging it close to a target angle. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a control device for a vehicular device that is hydraulically driven, and more particularly to a technique that achieves both responsiveness when changing and controlling the operating position of a device, and stable stability at the operating position.

  Some vehicle-mounted internal combustion engines include a variable valve timing mechanism that continuously and variably controls the valve timing (opening / closing timing) of intake and exhaust valves by changing the rotational phase of the camshaft with respect to the crankshaft. In the hydraulically driven variable valve timing mechanism, the operation direction of the hydraulic drive mechanism is switched by switching the hydraulic path supplied to the hydraulic drive mechanism by the switching control valve (spool valve), and the hydraulic supply state to the hydraulic drive mechanism is changed. Control is performed to maintain the rotational phase at the target position.

In this way, in a system that switches the operating direction of the hydraulic drive mechanism by switching the hydraulic path with the hydraulic control valve, the switching control valve has a dead zone near the neutral position in order to keep the operating position in a stable state at the target position. I have.
However, the provision of the dead zone improves the retention of the operating position, but when changing the operating position, the start of the supply of hydraulic pressure is delayed until the dead zone is passed, resulting in a decrease in responsiveness.

In patent document 1, the responsiveness fall by the said dead zone is suppressed by vibrating a valve body with the amplitude more than half of the width | variety of a dead zone.
JP 2003-336529 A

However, even the thing of patent document 1 does not change that it has a dead zone, It is necessary to give the vibration of a big amplitude, and there existed a limit in the response improvement.
In order to improve responsiveness, it is conceivable to use a hydraulic control valve that eliminates the dead zone. However, when trying to control a hydraulic control valve that has no dead zone to the neutral position with a predetermined operation amount (duty ratio), the dead zone is As a result, the oil flow rate control performance at the neutral position deteriorates, and oil leakage occurs and deviates from the target value.By controlling the convergence with feedback control, the actual rotation phase will fluctuate around the target value. This causes a problem that the convergence stability (holding stability) is lowered.

  The present invention has been made paying attention to such conventional problems, and an object thereof is to achieve both responsiveness and holding stability of vehicle control equipment such as a hydraulically driven variable valve timing mechanism.

  Therefore, the invention according to claim 1 switches the operation direction of the vehicle equipment by switching the hydraulic path for supplying the hydraulic pressure from the hydraulic source to the hydraulic drive mechanism of the vehicle equipment by operating the hydraulic control valve. A control device for a hydraulically driven vehicle device that controls a hydraulic pressure supply state to the hydraulic drive mechanism and maintains an operating position of the vehicle device, wherein an operation amount-oil flow rate characteristic of the hydraulic control valve is substantially linear. On the other hand, the operation signal of the hydraulic control valve is vibrated to maintain the hydraulic pressure supply state to the hydraulic drive mechanism.

The invention according to claim 2 further comprises feedback control means for detecting an actual operation position and performing feedback control so that the operation position of the vehicular equipment is the target position, and the target position and the actual operation position are provided. The operation signal of the hydraulic control valve is oscillated when the deviation is less than a predetermined value.
The invention according to claim 3 is configured such that the level of vibration given to the operation signal is set to be within a control deviation in which the operating position change amount of the hydraulic drive mechanism due to the vibration is allowed. .

  According to a fourth aspect of the present invention, the vehicle equipment is a variable valve mechanism for an internal combustion engine mounted on the vehicle.

  According to the first aspect of the present invention, when the control of changing the operating position of the vehicle equipment is performed by eliminating the dead zone by making the operation amount-oil flow rate characteristic of the hydraulic control valve substantially linear, high responsiveness is ensured. In addition, it is possible to stably maintain the hydraulic pressure supply state while preventing the displacement of the operation position due to oil leakage by giving a slight vibration to the operation signal, so that the operation position of the vehicle device is stabilized. Can be maintained.

According to the invention of claim 2, it is possible to reduce the calculation load by applying vibration to the operation amount only when the deviation between the target position of the operation position where the holding stability is a problem and the actual position is within a predetermined value, Power consumption can be reduced, and when the deviation exceeds a predetermined value, the application of vibration is prohibited, and adverse effects on responsiveness can be avoided.
According to the invention which concerns on Claim 3, the shift | offset | difference from the target position of the operation position of the apparatus for vehicles is stopped in the tolerance | permissible_range, and it can maintain in a stable state.

  According to the invention of claim 4, by applying to a hydraulically driven variable valve mechanism that varies the operating characteristics of the engine valve, such as a variable valve timing mechanism, the control response of the variable valve mechanism, Both holding stability can be maintained well.

Hereinafter, an embodiment in which the present invention is applied to a hydraulically driven variable valve timing mechanism provided in a vehicle-mounted internal combustion engine as a hydraulically driven vehicle device will be described.
1 to 6 show the configuration of the variable valve timing mechanism.
In the figure, a cam sprocket 1 (timing sprocket) that is rotationally driven via a timing chain by a crankshaft (not shown) of an engine (internal combustion engine), and a camshaft provided to be rotatable relative to the cam sprocket 1. 2, a rotating member 3 fixed to the end of the camshaft 2 and rotatably accommodated in the cam sprocket 1, and a hydraulic circuit 4 for rotating the rotating member 3 relative to the cam sprocket 1, And a lock mechanism 10 that selectively locks the relative rotational position of the cam sprocket 1 and the rotating member 3 at a predetermined position.

  The cam sprocket 1 includes a rotating portion 5 having a tooth portion 5a meshing with a timing chain (or timing belt) on the outer periphery, and a housing 6 disposed in front of the rotating portion 5 and rotatably accommodating the rotating member 3. A disc-shaped front cover 7 serving as a lid for closing the front end opening of the housing 6, and a substantially disc-shaped rear cover disposed between the housing 6 and the rotating portion 5 to close the rear end of the housing 6. 8, and the rotating portion 5, the housing 6, the front cover 7, and the rear cover 8 are integrally coupled from the axial direction by four small-diameter bolts 9.

  The rotating portion 5 has a substantially annular shape, and has four female screw holes 5b through which the small-diameter bolts 9 are screwed in the circumferentially equidistant positions of about 90 ° in the front-rear direction. A step-diameter fitting hole 11 into which a passage-forming sleeve 25 described later is fitted is formed through. Further, a disc-shaped fitting groove 12 into which the rear cover 8 is fitted is formed on the front end surface.

  The housing 6 has a cylindrical shape in which both front and rear ends are formed with openings, and four partition walls 13 project from the circumferential position of the inner peripheral surface at 90 °. The partition wall 13 has a trapezoidal shape in cross section, is provided along the axial direction of the housing 6, and both end edges are flush with the both end edges of the housing 6. Four bolt insertion holes 14 through which the small-diameter bolts 9 are inserted are formed penetrating in the axial direction. Further, a U-shaped seal member 15 and a leaf spring 16 that presses the seal member 15 inward are fitted into a holding groove 13a that is cut out along the axial direction at the center position of the inner end face of each partition wall 13. Is retained.

Further, the front cover 7 has a relatively large-diameter bolt insertion hole 17 at the center, and four bolt holes 18 at positions corresponding to the bolt insertion holes 14 of the housing 6. ing.
The rear cover 8 has a disc portion 8a fitted and held in the fitting groove 12 of the rotating member 5 on the rear end surface, and a small-diameter annular portion 25a of the sleeve 25 is fitted in the center. A fitting hole 8c is formed, and four bolt holes 19 are also formed at positions corresponding to the bolt insertion holes 14.

The camshaft 2 is rotatably supported at the upper end portion of the cylinder head 22 via a cam bearing 23, and a cam (not shown) that opens the intake valve via a valve lifter is integrated at a predetermined position on the outer peripheral surface. In addition, a flange portion 24 is integrally provided at the front end portion.
The rotating member 3 is fixed to the front end portion 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 in the flange portion 24 and the fitting hole 11, respectively. An annular base portion 27 having a bolt insertion hole 27a through which the fixing bolt 26 is inserted, and four vanes 28a, 28b, 28c, 28d integrally provided at 90 ° positions in the circumferential direction of the outer peripheral surface of the base portion 27; It has.

  Each of the first to fourth vanes 28a to 28d has a substantially inverted trapezoidal cross section, and is disposed in a recess between the partition walls 13, and separates the recess in the front and rear in the rotation direction. An advance side hydraulic chamber 32 and a retard side hydraulic chamber 33 are formed between both sides and both side surfaces of each partition wall portion 13. In addition, a U-shaped seal member 30 slidably contacting the inner peripheral surface 6a of the housing 6 and the seal member 30 are pressed outwardly into a holding groove 29 cut in the axial direction at the center of the outer peripheral surface of each of the vanes 28a to 28d. The leaf springs 31 are fitted and held.

  The lock mechanism 10 is formed through an engagement groove 20 formed at a predetermined position on the outer peripheral side of the fitting groove 12 of the rotating portion 5 and a predetermined position of the rear cover 8 corresponding to the engagement groove 20. An engagement hole 21 whose inner peripheral surface is tapered, a sliding hole 35 formed through the substantially central position of the one vane 28 corresponding to the engagement hole 21 along the internal axis direction, and the one A lock pin 34 slidably provided in the sliding hole 35 of the vane 28, a coil spring 39 that is a spring member elastically mounted on the rear end side of the lock pin 34, and the lock pin 34 and the sliding hole 35 is formed with a pressure receiving chamber 40 formed between them.

  The lock pin 34 includes a middle-side main body 34a on the center side, an engaging portion 34b formed in a substantially tapered shape on the front end side of the main body 34a, and a large step formed on the rear end side of the main body 34a. The engaging hole 21 is formed 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. And the hydraulic pressure in the pressure receiving chamber 40 formed between the outer peripheral surface between the main body 34a and the stopper portion 34c and the inner peripheral surface of the sliding hole 35. It slides in the direction of coming out of the engagement hole 21. The pressure receiving chamber 40 communicates with the retard angle side hydraulic chamber 33 through a through hole 36 formed in a side portion of the vane 28. Further, the engaging portion 34 b of the lock pin 34 is configured such that the engaging portion 34 b is engaged with the engaging hole 21 at the rotation position on the maximum retard angle side of the rotating member 3.

  The hydraulic circuit 4 includes two systems, a first hydraulic passage 41 that supplies and discharges hydraulic pressure to the advance side hydraulic chamber 32 and a second hydraulic passage 42 that supplies and discharges hydraulic pressure to the retard side hydraulic chamber 33. The hydraulic passage 41 and the hydraulic passage 41 are connected to a supply passage 43 and a drain passage 44 via an electromagnetically driven hydraulic control valve 45, respectively. The supply passage 43 is provided with an oil pump 47 that pumps the oil in the oil pan 46, while the downstream end of the drain passage 44 communicates with the oil pan 46.

  The first hydraulic passage 41 is branched and formed in the head portion 26a from the cylinder head 22 through the first passage portion 41a formed in the axial center of the camshaft 2 and the axial direction in the fixing bolt 26. A first oil passage 41b that communicates with the first passage portion 41a, a small-diameter outer peripheral surface of the head portion 26a, and an inner peripheral surface of the bolt insertion hole 27a that is provided in the base portion 27 of the rotating member 3 are the first. An oil chamber 41c that communicates with the oil passage 41b, and four branch passages 41d that are formed substantially radially in the base portion 27 of the rotating member 3 and communicate with the oil chamber 41c and each advance-side hydraulic chamber 32. Yes.

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

  In the hydraulic control valve 45, the internal spool valve element switches the hydraulic passages 41 and 42, the supply passage 43, and the drain passages 44a and 44b relative to each other, whereby the advance side hydraulic chamber 32 in the hydraulic drive mechanism of the VTC. , Switching the supply and discharge of the hydraulic pressure to the retard side hydraulic chamber 33 to switch the operation direction of the VTC, that is, the advance direction operation and the retard direction operation, and controlling the hydraulic pressure supply state to control the operation position of the VTC (camshaft (Rotation angle). The hydraulic control valve 45 is driven and controlled by a control signal from an engine control unit (ECU) 48.

Specifically, as shown in FIGS. 4 to 6, a cylindrical valve body 51 inserted and fixed in the holding hole 50 of the cylinder block 49 and a valve hole 52 in the valve body 51 are slidable. The spool valve body 53 is provided to switch the flow path, and a proportional solenoid type electromagnetic actuator 54 that operates the spool valve body 53.
The valve body 51 is formed with a supply port 55 penetrating the downstream end of the supply passage 43 and the valve hole 52 at a substantially central position of the peripheral wall, and the first and the second on both sides of the supply port 55. A first port 56 and a second port 57 that communicate with the other end of the second hydraulic passages 41 and 42 and the valve hole 52 are formed penetratingly. Further, third and fourth ports 58 and 59 are formed through both ends of the peripheral wall so as to communicate the drain passages 44a and 44b with the valve hole 52.

  The spool valve body 53 has a substantially cylindrical first valve portion 60 that opens and closes the supply port 55 at the center of the small diameter shaft portion, and opens and closes the third and fourth ports 58 and 59 at both ends. It has substantially cylindrical second and third valve portions 61 and 62. The spool valve body 53 is a conical valve spring 63 elastically mounted between an umbrella portion 53b provided at one end edge of the support shaft 53a on the front end side and a spring seat 51a provided on the inner peripheral wall of the front end side of the valve hole 52. Therefore, the supply valve 55 and the second hydraulic passage 42 are urged in the right direction in FIG.

The electromagnetic actuator 54 includes a core 64, a moving plunger 65, a coil 66, a connector 67, and the like, and a driving rod 65 a that presses the umbrella portion 53 b of the spool valve body 53 is fixed to the tip of the moving plunger 65.
The ECU 48 detects the current operating state (load, rotation) from signals from the rotation sensor 101 that detects the engine rotation speed and the air flow meter 102 that detects the intake air amount, and from the crank angle sensor 103 and the cam sensor 104. The relative rotation position between the cam sprocket 1 and the camshaft 2, that is, the rotational phase of the camshaft 2 with respect to the crankshaft is detected by the signal.

The ECU 48 controls the energization amount for the electromagnetic actuator 54 based on a duty control signal.
For example, when a control signal (OFF signal) with a duty ratio of 0% is output from the ECU 48 to the electromagnetic actuator 54, the spool valve element 53 moves to the position shown in FIG. As a result, 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. Therefore, the hydraulic oil pressure-fed from the oil pump 47 is supplied to the retarded hydraulic chamber 33 through the supply port 55, the valve hole 52, the second port 57, and the second hydraulic passage 42, and at the advanced side. The hydraulic oil in the hydraulic chamber 32 is discharged from the first drain passage 44a into the oil pan 46 through the first hydraulic passage 41, the first port 56, the valve hole 52, and the third port 58.

Therefore, the internal pressure of the retard side hydraulic chamber 33 is high and the internal pressure of the advance side hydraulic chamber 32 is low, and the rotating member 3 rotates in one direction at the maximum via the vanes 28a to 28b. As a result, the cam sprocket 1 and the camshaft 2 rotate relative 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) with a duty ratio of 100% is output from the ECU 48 to the electromagnetic actuator 54, the spool valve element 53 is maximally leftward against the spring force of the valve spring 63 as shown in FIG. At the same time as the third valve portion 61 closes the third port 58 by sliding, the fourth valve portion 62 opens the fourth port 59, and the first valve portion 60 includes the supply port 55 and the first port 56. To communicate with. Therefore, the hydraulic oil is supplied into the advance side 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 side hydraulic chamber 33 is second. The oil is discharged to the oil pan 46 through the hydraulic passage 42, the second port 57, the fourth port 59, and the second drain passage 44b, and the retard side hydraulic chamber 33 becomes low pressure.

For this reason, the rotating member 3 rotates to the maximum in the other direction via the vanes 28a to 28d, whereby the cam sprocket 1 and the camshaft 2 are relatively rotated to the other side and the phase is changed. The opening timing of the intake valve is advanced (advanced), and the overlap with the exhaust valve is increased.
The ECU 48 has a position in which the first valve unit 60 closes the supply port 55, the third valve unit 61 closes the third port 58, and the fourth valve unit 62 closes the fourth port 59. The relative duty position (rotation phase) between the cam sprocket 1 and the camshaft 2 detected based on signals from the crank angle sensor 103 and the cam sensor 104, and the operating state The feedback correction amount UDTY for matching the target value (target advance value) of the relative rotation position (rotation phase) set accordingly is set as described later, and the base duty ratio BASEDTY and the feedback correction amount UDTY are set. Is the final duty ratio VTCDTY, and the control signal for the duty ratio VTCDTY is electromagnetic actuator. It is to be output to the data 54.

The base duty ratio BASEDTY is a substantially median 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 no oil is 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 retarding direction, the duty ratio is reduced by the feedback correction UDTY, and the hydraulic oil pumped from the oil pump 47 is retarded. While being supplied to the hydraulic chamber 33, the hydraulic oil in the advance side hydraulic chamber 32 is discharged into the oil pan 46. Conversely, the relative rotation position (rotation phase) is changed in the advance direction. When it is necessary to increase the duty ratio, the duty ratio is increased by the feedback correction UDTY, the hydraulic oil is supplied into the advance hydraulic chamber 32, and the hydraulic oil in the retard hydraulic chamber 33 is supplied to the oil pan 46. Will be discharged. When the relative rotation position (rotation phase) is maintained in the current state, the absolute value of the feedback correction UDTY is decreased to return to a duty ratio near the base duty ratio. The internal pressures of the hydraulic chambers 32 and 33 are controlled by closing the 55, the third port 58, and the fourth port 59 (stopping the supply and discharge of hydraulic pressure).

  The feedback correction amount UDTY is calculated by, for example, normal PID control. That is, when the detected relative rotation position (rotation phase) between the cam sprocket 1 and the camshaft 2 is the actual angle VTCNOW of the variable valve timing mechanism (VTC), and the target value is the VTC target angle VTCTRG, The proportional component P, the integral component I, and the differential component D are calculated and controlled with respect to the deviation VTCERR (= VTCNOW−VTCTRG).

Here, in the case of the normal hydraulic control valve 45, the axial length of the first valve portion 60 of the spool valve body 53 is formed to be larger than the axial length of the supply port 55, whereby the spool valve A dead zone is provided in which the first valve portion 60 closes the supply port 55 even when the body 53 is moved by a predetermined amount in the axial direction and the state where oil is not supplied or discharged is maintained.
However, in the present invention, the axial length of the first valve portion 60 of the spool valve body 53 is formed to be substantially equal to (slightly larger than) the axial length of the supply port 55. As a result, when the spool valve body 53 is moved even a little from the state where the first valve portion 60 is in the neutral position where the supply port 55 is closed and the oil is not supplied or discharged, the supply port 55 is opened and the first port is opened. The oil is supplied and discharged in communication with the port 56 or the second port 57 so that there is substantially no dead zone, and the operation amount (duty ratio) of the hydraulic control valve 45-oil as shown in FIG. The flow rate characteristic is a substantially linear characteristic.

When the hydraulic control valve 45 has a linear characteristic with no dead zone in this way, when the target angle of the variable valve timing mechanism is changed, the supply and discharge of the hydraulic pressure is quickly started by the movement of the spool valve body 53 due to the change in the operation amount. Therefore, control responsiveness is improved.
However, since the dead zone is eliminated, even if the spool valve body 53 is controlled to the closed position, oil leaks from the gap between the supply port 55 and the first valve portion 60 to the first port 56 or the second port 57. Even if the actual angle of VTC deviates from the target value and is converged by the above feedback control, feedback control is performed so that the deviation amount caused by oil leakage is detected and corrected. However, due to the delay, as shown in FIGS. 7B and 7C, the actual angle of the VTC greatly fluctuates around the target angle, and the holding stability decreases.

Therefore, in the present invention, the holding stability is compensated by giving a vibration component to the operation amount of the VTC (hydraulic control valve 45).
FIG. 8 shows a control block diagram of the present embodiment. As described above, the feedback compensator block calculates the proportional component P, the integral component I, and the differential component D based on the deviation between the actual angle VTCNOW and the target angle VTCTRG of the variable valve timing mechanism (VTC) to be controlled. These are added to calculate the feedback correction amount UDTY.

In the vibration component calculation block, a vibration component (amplitude and period) to be added to the feedback correction amount UDTY that is the operation amount is calculated.
Then, an operation amount (duty ratio) obtained by adding a vibration component to the feedback correction amount UDTY is supplied to a VTC (hydraulic control valve 45) that is a control target.
FIG. 9 shows the operation in the basic form of the present invention. By swinging the operation amount around the neutral position, the actual VTC angle is converged and held within a predetermined range around the target angle.

Here, the vibration component calculation block applies the vibration component only in the region where the absolute value of the deviation VTCERR is less than a predetermined value and the holding stability is a problem based on the target angle VTCTRG and the actual angle VTCNOW, for example. can do.
In this way, the calculation load can be reduced by applying vibration only when necessary, power consumption can be reduced, and when the absolute value of the deviation VTCERR exceeds a predetermined value, the application of the vibration component is prohibited, thereby providing a response. The adverse effect on sex can also be avoided.

Note that the same effect can be obtained even when the vibration component is added only when the feedback correction amount UDTY (absolute value) is within a predetermined range instead of the absolute value of the deviation VTCERR.
Further, the amplitude and period of the vibration component, for example, as shown in FIG. 10, the maximum oil flow rate from the hydraulic control valve (see the hatched portion in the figure) obtained as a value that correlates with the product of the amplitude and period, Even if it flows into either the advance-side hydraulic chamber 32 or the retard-side hydraulic chamber 33, the VTC angle change amount is set within the allowable control deviation of the variable valve timing mechanism.

  For higher accuracy, the maximum oil flow rate from the hydraulic control valve obtained based on the response time constant of the oil flow rate is added to the amplitude and cycle of the vibration component as shown in FIG. ) Is set such that even if it flows into either the advance side hydraulic chamber 32 or the retard side hydraulic chamber 33, the VTC angle change amount is within an allowable control deviation of the variable valve timing mechanism.

Further, since the oil flow characteristic of the hydraulic control valve 45 also varies depending on engine state parameters such as the engine rotational speed Ne, the water temperature Tw, and the oil temperature To, the vibration level set as described above is set to the engine rotational speed Ne. The correction may be made based on the water temperature Tw, the oil temperature To, or the like.
Thus, by appropriately setting the vibration level, the deviation of the angle VTCNOW from the target angle VTCTGR can be kept within an allowable range, and the holding stability can be improved as much as possible.

Furthermore, the vibration level may be variably set according to the absolute value of the deviation VTCERR or the feedback correction amount UDTY. For example, the vibration level increases as the absolute value of the deviation VTCERR or the feedback correction amount UDTY decreases. It may be set to increase the holding stability.
The present invention is not limited to the variable valve timing mechanism described above, and the operation direction is switched by switching the hydraulic path supplied to the hydraulic drive mechanism by operating the hydraulic control valve, and the hydraulic supply state to the hydraulic drive mechanism is controlled. The present invention can be applied to any control device for hydraulically driven vehicle equipment configured to hold the operating position, and the same effect can be obtained.

Here, technical ideas other than the claims that can be grasped from the above embodiment will be described together with effects.
(A) In the control device for hydraulically driven vehicle equipment according to any one of claims 1, 3, and 4,
Feedback control means for detecting the actual operation position and performing feedback control so that the operation position of the vehicle equipment is the target position;
When the operation amount (absolute value) set in the feedback control is within a predetermined range, the operation signal of the hydraulic control valve is vibrated.

According to such a configuration, only when the feedback operation amount (absolute value) in which the holding stability is a problem is within a predetermined value, it is possible to reduce the calculation load by applying vibration to the operation amount, to reduce power consumption, and to reduce the operation amount. By inhibiting the application of vibration when the (absolute value) exceeds a predetermined value, adverse effects on responsiveness can be avoided.
(B) In the control device for hydraulically driven vehicle equipment according to any one of claims 2 to 4 and (a) above,
The level of vibration given to the operation signal is variably set according to the absolute value of the deviation between the target position and the actual operation position or the absolute value of the hydraulic control valve operation amount.

According to such a configuration, for example, it is possible to increase the holding stability by setting the vibration level to increase as the deviation between the target position of the vehicle device and the actual position or the absolute value of the feedback operation amount decreases.
(C) In the control device for hydraulically driven vehicle equipment according to any one of claims 3 and 4 and (a) and (b) above,
The maximum oil flow rate of the hydraulic control valve obtained based on the amplitude and cycle of vibration given to the operation signal of the hydraulic control valve or taking into account the response time constant of the oil flow rate of the hydraulic control valve in these amplitude and cycle is the vehicle The vibration level is set so that the amount of change in the operating position of the vehicle equipment when it is supplied to or discharged from the hydraulic drive mechanism of the equipment is within a steady deviation allowable value.

According to this configuration, the vibration level can be set to the minimum necessary, and the operation position can be maintained in a stable state with high accuracy.
(D) In the control device for a hydraulically driven vehicle device described in (c) above,
The vibration level is corrected and set based on at least one of a rotation speed of a vehicle-mounted engine, a water temperature, and an oil temperature.

  According to such a configuration, the oil flow rate characteristic of the hydraulic control valve also changes depending on the engine state parameters. Therefore, the vibration level can be set with higher accuracy by correcting the vibration level according to these engine state parameters.

Sectional drawing which shows the structure of the variable valve timing mechanism in embodiment. BB sectional drawing of FIG. The disassembled perspective view of the said variable valve timing mechanism. The longitudinal cross-sectional view which shows the hydraulic control valve in the said variable valve timing mechanism. The longitudinal cross-sectional view which shows the hydraulic control valve in the said variable valve timing mechanism. The longitudinal cross-sectional view which shows the hydraulic control valve in the said variable valve timing mechanism. The figure which shows the state at the time of performing simple feedback control with the operation amount-oil flow characteristic of the hydraulic control valve in the said embodiment, and this characteristic. The control block diagram of the said embodiment. The time chart which shows the basic effect | action of the said embodiment. The figure which shows an example which sets the vibration level provided to the operation amount in the said embodiment with high precision. The figure which shows another example which sets the vibration level provided to the operation amount in the said embodiment with high precision.

Explanation of symbols

    DESCRIPTION OF SYMBOLS 41 ... 1st hydraulic path 42 ... 2nd hydraulic path 48 ... ECU 53 ... Spool valve body 56 ... 1st port 57 ... 2nd port 58 ... 3rd port 59 ... 4th port 60 ... 1st valve part

Claims (4)

By switching the hydraulic path to supply the hydraulic pressure from the hydraulic source to the hydraulic drive mechanism of the vehicle equipment by operating the hydraulic control valve, the operation direction of the vehicle equipment is switched, and the hydraulic supply state to the hydraulic drive mechanism is controlled. A control device for hydraulically driven vehicle equipment that holds the operating position of the vehicle equipment,
A hydraulic drive vehicle characterized in that an operation amount-oil flow characteristic of the hydraulic control valve is substantially linear, and an operation signal of the hydraulic control valve is vibrated to maintain a hydraulic pressure supply state to the hydraulic drive mechanism. Equipment control equipment. Feedback control means for detecting the actual operation position and performing feedback control so that the operation position of the vehicle equipment is the target position;
2. The control device for a hydraulically driven vehicle device according to claim 1, wherein an operation signal of the hydraulic control valve is vibrated when a deviation between the target position and an actual operation position is a predetermined value or less.   The level of vibration given to the operation signal is set so that the amount of change in the operation position of the hydraulic drive mechanism due to the vibration is within an allowable control deviation. The control apparatus of the apparatus for hydraulic drive vehicles of description.   The control device for a hydraulically driven vehicle device according to any one of claims 1 to 3, wherein the vehicle device is a variable valve mechanism of an internal combustion engine mounted on the vehicle.

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