JP2001065373A - Hydraulic control device for variable valve timing mechanism - Google Patents

Abstract

A variable valve timing mechanism (VVT) ensures quick operation response and stabilizes position control when a ring gear is held at an intermediate position. A control valve (28) for controlling the supply of oil pressure to a ring gear (12) in a VVT (1) that varies valve timing by driving the ring gear (12) by oil pressure.
Is controlled based on a duty frequency synchronized with a phase inverted with respect to a cycle of torque fluctuation of the camshaft 2 generated when the intake valve is opened and closed.
Accordingly, the hydraulic pressure periodically applied to the ring gear 12 acts so as to cancel out the thrust force periodically applied to the ring gear 12 due to the torque fluctuation of the camshaft 2, and the fluctuation of the resultant force is reduced. Become smaller.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for a variable valve timing mechanism which is driven by hydraulic pressure in order to change the valve timing of an intake valve or an exhaust valve in an internal combustion engine.

[0002]

2. Description of the Related Art Conventionally, as a valve control device of this type of hydraulic drive type, for example, one disclosed in Japanese Patent Application Laid-Open No. 59-120707 is known. In this conventional valve control device, a sleeve (timing pulley) drivingly connected to a crankshaft of an engine is provided at one end of a camshaft, and a ring piston (ring gear) is provided between the timing pulley and the camshaft. Is interposed. At least one of the teeth provided on the inner and outer circumferences of the ring gear is a helical tooth, and the ring gear is rotatable relative to the camshaft by moving in the axial direction. Also, at one axial end of the ring gear,
Oil pressure from a hydraulic pump or the like is supplied through an oil passage. Further, a balancing spring is provided at the other end of the ring gear to oppose the oil pressure. When the oil pressure is applied to one end of the ring gear, the ring gear is moved in the axial direction against the urging force of the spring. As a result, the camshaft is twisted with respect to the timing pulley, the phase (rotational phase) of the camshaft and the timing pulley in the rotation direction is changed, and the opening / closing timing of the intake valve or the exhaust valve is changed, thereby causing a valve over. The lap is controlled.

In this conventional valve control device, an electromagnetic valve is provided in the middle of an oil passage for controlling the oil pressure acting on the ring gear. The duty of the solenoid valve is controlled by a computer based on various drive parameters of the engine. Accordingly, when the duty of the solenoid valve is controlled, a hydraulic pressure that fluctuates in synchronization with the duty frequency is applied to one axial end of the ring gear.

[0004]

However, in the conventional valve control device, when the camshaft is driven to rotate, the reaction force of the periodic torque fluctuation of the camshaft is input to the ring gear. The ring gear was subjected to a periodic axial load (thrust force). Further, the ring gear having the helical teeth has some friction with respect to the movement in the axial direction.

Therefore, when the friction of the ring gear is relatively small, there are the following problems. That is,
At both ends of the movement stroke of the ring gear, since the biasing force of the spring or the hydraulic pressure is the largest, the influence of the periodic thrust force applied to the ring gear is extremely small.
However, when the ring gear is held at the intermediate position by the duty control of the solenoid valve, the position is determined by the balance between the urging force of the spring and the hydraulic pressure. Therefore, a cyclic thrust force is further applied to the ring gear in the balanced state, whereby the position control of the ring gear becomes unstable. Therefore, the ring gear may be inadvertently moved, and the helical teeth may be worn, the oil seal may be worn, or the durability of the mechanism itself may be impaired.

On the other hand, when the friction of the ring gear is relatively large, there are the following problems. That is, the friction of the ring gear exists as a hysteresis for the movement of the ring gear. Therefore, when the hysteresis is large, especially when the ring gear is moved from the intermediate position, the response is affected. For example, when the oil pressure applied to the ring gear is slightly increased or decreased to slightly move the ring gear, the oil pressure change may be buried in the hysteresis. Therefore, the ring gear starts moving slowly,
The position control was likely to be non-linear.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned circumstances, and an object of the present invention is to provide a variable valve timing mechanism for varying a valve timing by hydraulic pressure, so that a quick operation response can be achieved when changing the valve timing. It is an object of the present invention to provide a hydraulic control device for a variable valve timing mechanism which can secure the valve timing and can stabilize the valve timing at a predetermined timing.

[0008]

The means for achieving the above object and the effects thereof will be described below. According to the present invention, in the hydraulic control device for a variable valve timing mechanism for changing at least one valve timing of an intake / exhaust valve driven to open and close by a camshaft of an internal combustion engine based on a hydraulic pressure supplied through an oil passage, A control valve provided in the middle of the road and driven to open and close to control the supply of the hydraulic pressure, and a duty frequency synchronized with a torque fluctuation cycle of the camshaft generated when opening and closing the intake and exhaust valves. And duty control means for duty-controlling the control valve based on the duty.

According to this configuration, when the opening / closing cycle of the control valve is synchronized in phase with the torque fluctuation cycle of the camshaft, the periodic thrust force caused by the torque fluctuation of the camshaft is adjusted to the duty frequency. In order to increase the hydraulic pressure that fluctuates periodically, it is possible to secure quick operation response when changing the valve timing. On the other hand, when the opening / closing cycle of the control valve is synchronized in the opposite phase to the torque fluctuation cycle of the camshaft, the hydraulic pressure that fluctuates periodically in accordance with the duty frequency becomes a periodic pressure caused by the torque fluctuation of the camshaft. Since it acts so as to cancel out the excessive thrust force, it is possible to stabilize the valve timing when maintaining it at a predetermined timing.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS (First Embodiment) A first embodiment of the present invention embodied as a hydraulic control device for a variable valve timing mechanism applied to a gasoline engine will be described below in detail with reference to FIGS. Will be described.

FIG. 1 shows a variable valve timing mechanism (hereinafter simply referred to as "VVT") 1 of this embodiment and its VVT1.
FIG. 2 is a schematic configuration diagram illustrating a hydraulic control device for driving the hydraulic control device. The VVT 1 has a camshaft 2 for opening and closing a valve for intake and exhaust, and the camshaft 2 is rotatably supported by a cylinder head 3 of an engine (not shown) by a cam journal 2a. In this embodiment, an intake valve and an exhaust valve (not shown) of the engine are opened and closed by separate camshafts, and the camshaft 2 corresponds to the intake valve.

A timing pulley 4 is mounted on the tip of the camshaft 2 so as to be relatively rotatable by a boss 4a. External teeth 4b are provided on the outer periphery of the timing pulley 4.
The pulley 4 has an accommodation recess 4c formed on one side in the axial direction (the left-right direction in the figure). In addition, a cap 5 is fastened and fixed to the tip of the camshaft 2 with a bolt 6 so as to cover the housing recess 4c.
The rotation is prevented by the knock pin 7. Further, between the opening end side of the accommodation recess 4 c and the outer periphery of the cap 5, an outer plate 8 press-fitted and fixed to the timing pulley 4 and an inner plate 9 integrally formed with the cap 5.
A well-known viscous coupling (a viscous coupling) 10 for buffering is provided. Furthermore, seal members 11 are interposed between the cap 5 and the outer plate 8 and between the cap 5 and the timing pulley 4, respectively.

A ring gear 12 is interposed between the timing pulley 4 and the camshaft 2, and the two are connected to each other. That is, the ring gear 12 is provided in the accommodation recess 4 c of the timing pulley 4 closed by the cap 5.
Is housed. In this ring gear 12, both teeth 12a and 12b provided on the inner and outer circumferences are helical teeth. Also, each tooth 12a, 12b of the ring gear 12
Are meshed with the internal teeth 4d formed on the boss 4a of the timing pulley 4 and the internal teeth 5a formed on the inner periphery of the cap 5, respectively. With this configuration, the ring gear 12 is moved in the axial direction,
Are rotatable relative to the camshaft 2. Furthermore,
The timing pulley 4 is drivingly connected to a crankshaft (not shown) via a timing belt 13 mounted on the external teeth 4b.

Accordingly, when the power of the crankshaft is transmitted to the timing pulley 4 via the timing belt 13, the timing pulley 4 and the cap 5 connected by the ring gear 12 are integrally rotated, and the camshaft 2 is rotated. It is driven to rotate. At this time, when the ring gear 12 is moved in the axial direction, the camshaft 2 is given a twist with respect to the timing pulley 4. In addition, when the camshaft 2 is torsioned, rattling due to backlash of the ring gear 12 is buffered by the action of the viscous coupling 10 and generation of abnormal noise is suppressed.

In order to drive the ring gear 12 by hydraulic pressure, one end in the axial direction of the ring gear 12 in the housing recess 4c is a pressurizing chamber 14 into which hydraulic pressure by hydraulic oil is introduced. Similarly, in the housing recess 4c, the other end side of the ring gear 12 is a spring chamber 16 in which a balancing spring 15 against the oil pressure is housed. Further, a head oil passage 17 and a shaft oil passage 18 communicating with each other are formed in the cylinder head 3 and the camshaft 2 in order to supply hydraulic pressure by hydraulic oil to the pressurizing chamber 14.

On the other hand, return oil passages 19 and 20 are formed in the timing pulley 4 and a part of the camshaft 2 to lead hydraulic oil leaked from the pressurizing chamber 14 to the spring chamber 16. At one end of the timing pulley 4, a seal member 21 is interposed between the pulley 4 and the cylinder head 3, and the timing pulley 4,
Cylinder head 3, camshaft 2, and seal member 21
The space surrounded by is defined as an oil recovery chamber 22. Then, return oil passages 19 and 20 are connected to the oil recovery chamber 22. Further, below the camshaft 2, an oil return hole 24 for returning the hydraulic oil collected in the oil recovery chamber 22 to the oil pan 23 of the engine is formed in a part of the cylinder head 3.

In this embodiment, the inner and outer teeth 12
The friction of the ring gear 12, which is determined by the inclination angles of a and 12b, is set to be relatively small. In this embodiment, the lubricating oil of the engine is used as the working oil. That is, the oil pump 25 constituting one hydraulic circuit for lubricating oil is driven in conjunction with the engine, whereby the lubricating oil stored in the oil pan 23 is sucked up through the oil filter 26. A main oil passage 27 is connected between the oil pump 25 and the head oil passage 17, and a solenoid type three control valve 28 is provided in the middle of the main oil passage 27. The duty of the control valve 28 is controlled, and its characteristics are shown in the graph of FIG. As can be seen from this graph, in the control valve 28, the hydraulic pressure as the output with respect to the duty ratio is proportional to the duty ratio within a certain range of the duty ratio as the input command value. That is, within a certain range of the duty ratio, the hydraulic pressure increases as the duty ratio increases. Then, by increasing the hydraulic pressure controlled by the control valve 28, the rotation phase of the camshaft 2 is advanced.

Therefore, during the operation of the oil pump 25, the on / off of the control valve 28 is duty-controlled to periodically open and close the main oil passage 27, so that the lubricating oil discharged from the oil pump 25 operates. The oil is supplied to the head oil passage 17 with a certain oil pressure as oil. The hydraulic oil supplied to the head oil passage 17 is further introduced into the pressurizing chamber 14 through the shaft oil passage 18. Then, the ring gear 12 is pressed in one of the axial directions (rightward in the drawing) against the urging force of the spring 15 by the hydraulic pressure of the hydraulic oil. Thereby, torsion is given to the camshaft 2 and the rotation phase of the camshaft 2 with respect to the timing pulley 4 is changed. As a result, the opening / closing timing of the intake valve is changed, and the valve overlap between the intake valve and the exhaust valve is changed.

On the other hand, when the control valve 28 is turned off, the supply of the working oil to the head oil passage 17 is shut off. As a result, the oil pressure is released from the pressurizing chamber 14 and the ring gear 12
Is pressed back by the urging force of the spring 15 to the other axial direction (to the left in the figure). As a result, a reverse twist is applied to the camshaft 2, and the rotation phase of the camshaft 2 is changed back. As a result, the opening / closing timing of the intake valve is changed, and the valve overlap is changed. At this time, the operating oil leaked from the pressurizing chamber 14 to the spring chamber 16 passes through the return oil passages 19 and 20 to the oil collecting chamber 22.
To the oil pan 23 through the oil return hole 24.
Returned to.

In this embodiment, the control valve 28 is connected to an electronic control unit (hereinafter simply referred to as “EC
U ”) 31 is controlled according to the operating state of the engine. The ECU 31 has a central processing unit (CPU), various memories for storing a predetermined control program and the like in advance and temporarily storing the calculation results of the CPU, etc., and buses these units, an external input circuit and an external output circuit, etc. As a theoretical operation circuit connected by the

An air flow meter 32 for detecting an intake air amount Q is provided in an intake passage of the engine (not shown) as various sensors for detecting an operation state of the engine. Further, a water temperature sensor 33 for detecting the temperature (cooling water temperature) THW of the cooling water is provided in the cylinder block of the engine. Further, in the vicinity of a camshaft not forming the VVT 1 on the exhaust valve side, the rotation of the crankshaft is detected at predetermined angular intervals from the rotation of the camshaft, and is used as a crank angle signal CAS for obtaining the engine speed NE. An output crank angle sensor 34 is provided. Further, a cam angle sensor 35 for detecting the rotation of the camshaft 2 at predetermined angular intervals and outputting it as a cam angle reference signal CBS is provided at the base end of the camshaft 2 constituting the VVT 1 on the intake valve side. Have been. The cam angle sensor 35 is provided for calculating an advance angle value by comparison with the crank angle sensor 34 and for calculating a reference timing for turning the control valve 28 on and off. The cam angle sensor 35 includes a timing rotor 36 that rotates in conjunction with the rotation of the camshaft 2, and a pickup coil 37. A plurality of protrusions 36a are formed on the outer periphery of the timing rotor 36 so as to match the number of cam ridges (not shown) on the camshaft 2.
These projections 36a are arranged so as to match the cycle of the torque fluctuation generated in the camshaft 2 when the intake valve is opened and closed by each cam peak. In this embodiment, the projections 36a are arranged so as to match the position where the level of the torque fluctuation is the lowest. Further, the pickup coil 37 is arranged to face each projection 36a. Then, when the timing rotor 36 is rotated with the rotation of the crankshaft, the pickup coil 37 detects the passage of each projection 36a, and the cam angle signal CBS is a trigger pulse matching the position where the torque fluctuation level becomes the lowest. Is output periodically.

In this embodiment, the air flow meter 32 and the water temperature sensor 3
3. The crank angle sensor 34 and the cam angle sensor 35 are connected respectively. The control valve 28 is connected to an external output circuit of the ECU 31. Then, the ECU 31
In order to control the valve timing by the VT1, the opening / closing timing of the intake valve according to the current engine operating state is determined based on the output signals from the air flow meter 32 and the sensors 33 to 35, and the control valve 28 is suitably driven. Control.

Next, the processing operation for valve timing control executed by the ECU 31 will be described. 2 shows a “valve timing control routine” executed by the ECU 70.
It is executed by a periodic interruption every predetermined time during the operation of the engine.

When the process proceeds to this routine, first, in step 101, the air flow meter 32, the water temperature sensor 33, the crank angle sensor 34, and the cam angle sensor 35
Based on each output signal, the intake air amount Q, the cooling water temperature T
The HW, the crank angle signal CAS and the cam angle signal CBS are read.

Subsequently, at step 102, it is determined whether or not the currently read coolant temperature THW is higher than a predetermined reference temperature T0. Here, if the cooling water temperature THW is not higher than the reference temperature T0, it is assumed that the engine is not sufficiently warmed up and the subsequent processing is temporarily terminated.

On the other hand, if the cooling water temperature THW is higher than the reference temperature T0 in step 102, it is determined that the engine has been sufficiently warmed up and the process proceeds to step 103.
Then, a series of processes from Step 103 to Step 109 are executed.

That is, first, at step 103, the currently read crank angle signal CAS and cam angle signal C are read.
Based on BS, the engine speed NE and the camshaft 2
Is calculated.

Next, at step 104, an engine load equivalent value Q / N is calculated based on the currently read intake air amount Q and the currently calculated engine speed NE. In step 105, a target advance value θ of the camshaft 2 is calculated based on the engine speed NE and the load equivalent value Q / N calculated this time. The calculation of the target advance value θ is performed, as shown in FIG.
And a map in which the relationship between the target advance angle value θ and the load equivalent value Q / N is determined in advance.

Subsequently, in step 106, the advance value deviation (θ-θ0) between the target advance value θ and the current advance value θ0 is obtained.
, The current duty addition value R is calculated. The calculation of the duty addition value R is performed with reference to a map in which the relation between the duty addition value R and the advance angle deviation (θ−θ0) is predetermined as shown in FIG.

In step 107, the duty addition value R calculated this time is added to the duty ratio RB calculated last time, and the result is set as the current duty ratio RA.

Further, in step 108, based on the engine speed NE and the duty ratio RA obtained this time, the energizing time R for controlling the duty of the control valve 28 is determined.
Calculate T. The calculation of the energization time RT is performed with reference to a map (not shown) in which the relationship between the energization time RT and the engine speed NE and the duty ratio RA is predetermined.

In step 109, the duty of the control valve 28 is controlled based on the energization time RT obtained this time. In this embodiment, the energization time RT is output to the solenoid of the control valve 28 based on the duty frequency synchronized with the phase of the torque fluctuation generated in the camshaft 2 at an inverted phase, and the subsequent processing is temporarily terminated.

The processing operation of the valve timing control is executed as described above. Here, the operation of the valve timing control of this embodiment will be described with reference to the time chart of FIG.

This time chart shows the torque fluctuation generated in the camshaft 2, the thrust force in the ring gear 12 caused by the fluctuation, the trigger pulse from the cam angle sensor 35,
Energizing the solenoid of the control valve 28, the ring gear 12
2 shows the relationship between the change in the resultant oil pressure, the oil pressure, the spring force, and the thrust force.

As can be seen from this time chart,
The torque fluctuation of the camshaft 2 is generated periodically, and the resulting thrust force of the ring gear 12 is generated at a cycle inverted with respect to the cycle of the torque fluctuation. Then, in the solenoid of the control valve 28, energization for the energization time RT is started based on a trigger pulse synchronized with the timing at which the thrust force of the ring gear 12 becomes maximum. As a result, the hydraulic pressure applied to the ring gear 12
Of the thrust force with a period inverted with respect to the period of the thrust force.

Therefore, when the duty of the control valve 28 is controlled so as to advance the rotational phase of the camshaft 2 to the maximum, the maximum hydraulic pressure is supplied to the pressurizing chamber 14 and the ring gear 12 Energizing force (spring force)
Is moved to the right end position in FIG. Also, the control valve 28 is provided to retard the rotation phase of the camshaft 2 to the maximum.
Is turned off, no oil pressure acts on the pressurizing chamber 14, and the ring gear 12 is moved to the left end position in FIG. 1 by the spring force. When the ring gear 12 is moved to the right end position or the left end position as described above, the hydraulic pressure or the spring force applied to the ring gear 12 is sufficiently large. The position is held in a relatively stable state.

On the other hand, if the duty of the control valve 28 is controlled so as to advance the rotation phase of the camshaft 2 to a medium level, as shown in FIG. It is held at the middle position of the movement stroke. In this embodiment, the duty control of the control valve 28 is performed based on the duty frequency synchronized with the phase that is inverted with respect to the cycle of the torque fluctuation of the camshaft 2. That is, the ring gear 1
The duty of the control valve 28 is controlled so that the oil pressure applied to the pressure valve 2 fluctuates in a cycle inverted with respect to the cycle of the thrust force. Therefore, the hydraulic pressure periodically applied to the ring gear 12 acts so as to cancel out the thrust force periodically applied to the ring gear 12 due to the torque fluctuation of the camshaft 2.

Accordingly, when the friction of the ring gear 12 is relatively small as in the present embodiment, FIG.
As shown by the solid line in (f), the fluctuation range of the resultant force of the hydraulic pressure, the spring force and the thrust force acting on the ring gear 12 becomes very small, and does not exceed the friction width of the ring gear 12. The effect is also evident from a comparison with a case where a constant oil pressure is supplied to the ring gear 12 indicated by a broken line in FIG. Therefore, the ring gear 12 is not moved by the thrust force caused by the torque fluctuation of the camshaft 2.

As a result, when the ring gear 12 is held at the intermediate position, the position control can be stabilized. Further, as a result, inadvertent movement of the ring gear 12 can be prevented, and the ring gear 12
Wear on the helical teeth, and oil seal wear,
Alternatively, a decrease in the durability of the ring gear 12 itself can be prevented.

(Second Embodiment) Next, a second embodiment which embodies a hydraulic control device for a variable valve timing mechanism according to the present invention will be described with reference to FIGS. still,
As the configuration of this embodiment is the same as that of the first embodiment, the same components will be denoted by the same reference numerals and description thereof will be omitted, and different points will be mainly described below.

In this embodiment, the inner and outer teeth 12a, 12a,
The friction in the axial direction of the ring gear 12, which is determined by the inclination angle and the like of the ring gear 12b, is set to be relatively large. Further, the processing operation of the valve timing control executed by the ECU 31 is different from that of the first embodiment. That is, the flowchart of FIG. 7 shows a “valve timing control routine” executed by the ECU 70, which is executed by a periodic interruption every predetermined time during the operation of the engine.

In this flowchart, step 2
Since the processing of steps 01 to 208 is the same as the processing of the flowchart of FIG. 2 in the first embodiment, the description is omitted, and particularly a different step 209 is performed.
Will be described.

That is, in step 209, the duty of the control valve 28 is controlled based on the energization time RT obtained this time. In this embodiment, the energization time RT is output to the solenoid of the control valve 28 based on the duty frequency synchronized in phase with the cycle of the torque fluctuation of the camshaft 2, and the subsequent processing is temporarily terminated.

Here, the operation of the valve timing control of this embodiment will be described with reference to the time chart of FIG. 8 according to FIG. As you can see from this time chart,
The torque fluctuation of the camshaft 2 is generated periodically, and the resulting thrust force of the ring gear 12 is generated at a cycle inverted with respect to the cycle of the torque fluctuation. Then, in the solenoid of the control valve 28, the energization for the energization time RT is terminated based on the trigger pulse synchronized with the timing at which the thrust force of the ring gear 12 becomes maximum. As a result, the oil pressure applied to the ring gear 12
In other words, it fluctuates with the same phase as that of the thrust force.

Therefore, when the duty of the control valve 28 is controlled so as to advance the rotational phase of the camshaft 2, hydraulic pressure is supplied to the pressurizing chamber 14, and the ring gear 12 resists the spring force as shown in FIG. Moved to the right. In this embodiment, the duty control of the control valve 28 is performed based on the duty frequency synchronized in phase with the torque fluctuation cycle of the camshaft 2. That is, the duty of the control valve 28 is controlled so that the hydraulic pressure applied to the ring gear 12 fluctuates in the same phase as the cycle of the thrust force. For this reason, a thrust force caused by the torque fluctuation of the camshaft 2 is applied to the ring gear 12, and is added in addition to the hydraulic pressure periodically applied to the ring gear 12.

Therefore, when the friction of the ring gear 12 is relatively large as in this embodiment, FIG.
As shown by the solid line in (f), the resultant of the hydraulic pressure, the spring force, and the thrust force is applied for the axial movement of the ring gear 12. The variation width of the resultant force is substantially equal to the friction width of the ring gear 12. The effect is also evident from a comparison with a case where a constant oil pressure is supplied to the ring gear 12 indicated by a broken line in FIG.

As a result, even if the friction of the ring gear 12 is large, a resultant force sufficient to move the ring gear 12 in the axial direction is obtained, and the movement responsiveness of the ring gear 12 can be improved. For example, even when the hydraulic pressure applied to the ring gear 12 is slightly increased or decreased by the duty control of the control valve 28, the oil pressure change is not buried in the hysteresis caused by friction, and the ring gear 12 can be moved quickly. Can be. Therefore, the position control of the ring gear 12 can be performed in a more linear manner, and quick operation responsiveness when the valve timing is made variable by the VVT 1 can be secured.

It should be noted that the present invention is not limited to the above-described embodiments, and may be implemented as follows, with a part of the configuration appropriately changed without departing from the spirit of the invention. (1) In each of the above embodiments, the ring gear 12 is moved by supplying hydraulic pressure to one axial end of the ring gear 12 using a single hydraulic circuit. The ring gear may be moved by supplying hydraulic pressure using a hydraulic circuit.

(2) In each of the above-described embodiments, the VVT 1 that changes only the opening / closing timing of the intake valve is provided.
It is also possible to provide a VVT in which only the opening / closing timing of the exhaust valve is variable or a VVT in which the opening / closing timing of both the intake valve and the exhaust valve is variable.

(3) In each of the above embodiments, the VVT1 hydraulic control device is embodied in a gasoline engine.
Can be embodied in a diesel engine.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a variable valve timing mechanism and a hydraulic control device for driving the same in a first embodiment.

FIG. 2 is a flowchart showing a “valve timing control routine” executed by an ECU in the first embodiment.

FIG. 3 is a graph showing characteristics of a control valve that is duty-controlled in the first embodiment, and showing a relationship between a duty ratio and a hydraulic pressure.

FIG. 4 is a map showing a relationship between a target advance value and an engine speed and a load equivalent value in the first embodiment.

FIG. 5 is a map showing a relationship between an advance value deviation between a target advance value and a current advance value and a duty addition value in the first embodiment.

FIG. 6 is a graph showing torque fluctuations of a camshaft, a thrust force of a ring gear, a trigger pulse of a cam angle sensor, energization of a solenoid of a control valve, a hydraulic pressure and a hydraulic pressure on a ring gear, a resultant force of a spring force, and a thrust force in the first embodiment. 5 is a time chart showing the relationship between changes in the data.

FIG. 7 is a flowchart showing a “valve timing control routine” executed by the ECU in the second embodiment.

FIG. 8 is a diagram showing a variation in torque of a camshaft, a thrust force of a ring gear, a trigger pulse of a cam angle sensor, energization of a solenoid of a control valve, a hydraulic pressure, a hydraulic pressure, a hydraulic pressure, a spring force, and a thrust force of a ring gear in a second embodiment. 5 is a time chart showing the relationship between changes in the data.

[Explanation of symbols]

2 ... cam shaft, 3 ... cylinder head, 4 ... timing pulley, 12 ... ring gear, 12a, 12b ... teeth, 1
7: head oil passage, 18: shaft oil passage, 27: main oil passage, 28: control valve, 31: ECU constituting duty control means.

Claims (1)

[Claims] 1. A hydraulic control device for a variable valve timing mechanism for changing at least one valve timing of an intake / exhaust valve driven to open and close by a camshaft of an internal combustion engine based on a hydraulic pressure supplied through an oil passage, A control valve provided in the middle of the oil passage and driven to open and close to control the supply of the hydraulic pressure; and a duty frequency synchronized with a torque fluctuation cycle of the camshaft generated when opening and closing the intake and exhaust valves. And a duty control means for duty-controlling the control valve based on the hydraulic pressure control device.