JP2006200398A - Controller of variable valve mechanism - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To stably control a drive motor without saturating a motor drive current by a variable valve mechanism of an internal combustion engine. <P>SOLUTION: When a deviation ¾θ<SB>CST</SB>-θ<SB>CSM</SB>¾ between the arrival operating angle θ<SB>CST</SB>of the final target of the variable valve mechanism varying the lift and the operating angle of an engine valve and a target control shaft operating angle θ<SB>CSM</SB>transiently set to a specified response characteristic is large, a time constant T treating by primary delay the arrival operating angle θ<SB>CST</SB>is increased to largely limit a change rate and a feedback compensation amount and a feedback control amount are calculated by using the limited arrival operating angle and a feedback control is performed. Thus, the drive motor can be controlled with stable response characteristic without saturating the motor drive current against the abrupt change of the arrival operating angle θ<SB>CST</SB>. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a variable valve mechanism that can continuously and variably control lift characteristics (operation angle and lift amount) of intake and exhaust valves of an internal combustion engine, and a control device for other actuators.

As is well known, the lift characteristics of intake and exhaust valves are used to improve fuel economy at low engine speed and low load, to ensure stable operation, and to ensure sufficient output by improving intake charging efficiency at high speed and high load. There has been proposed a variable valve mechanism that can change the engine according to the engine operating state.
As a control of an actuator such as this type of variable valve mechanism, the one disclosed in Patent Document 1 has a desired response characteristic (normative response) according to the ultimate displacement that is the final target (the ultimate operating angle in the variable valve mechanism). As can be obtained, two-degree-of-freedom control is performed in which feedback control is performed while setting a feedforward compensation amount corresponding to the reference response.
JP 2001-3773 A

In the control disclosed in Patent Document 1, the feedforward compensation is set from the reached displacement and the normative response that are input command values without considering the saturation current of the electric motor (DC motor) that is an actuator. When the change amount of the input command value is large and the normative response is fast, the control system outputs a current command value that cannot be realized.
At this time, the actual motor current value is significantly lower than the current command value, in other words, the actual motor torque is significantly lower than the required motor torque. The deviation of the operating angle is accumulated as an integral part of the feedback control, and the response is greatly deteriorated, resulting in overshoot (see FIG. 17). Even in the case of driving by a hydraulic actuator other than the electric motor, for example, the same problem arises because the actual change in hydraulic pressure is delayed with respect to the hydraulic pressure command value.

  The present invention has been made paying attention to such a conventional problem, and is stable while preventing the occurrence of overshoot even when the input command value of the actuator that drives the variable valve mechanism and other mechanisms greatly changes. The purpose is to be able to obtain the response.

  In order to solve the above-mentioned problems, the present invention limits the rate of change by applying a rate-of-change limiter that is variably set according to a parameter representing the change state of the displacement with respect to the final target arrival displacement such as lift characteristics. In addition, feedback control is performed so that a desired response characteristic is obtained based on the ultimate displacement with the change rate limited.

  With such a configuration, even when the ultimate displacement, which is the input command value, changes greatly, the feedforward compensation amount and the transient target value are set based on the ultimate displacement with the rate of change limiter applied and the rate of change limited. However, by performing feedback control, a desired response characteristic can be obtained with the maximum capacity of the actuator, a stable response can be obtained, and the occurrence of overshoot can be prevented.

1 to 3 show an embodiment in which a variable valve mechanism for an internal combustion engine according to the present invention is applied to the intake valve side. In FIG. 1, the configuration on the exhaust valve side (lower side in FIG. 1) is not shown.
A drive shaft 11 that is continuous over all the cylinders is provided on the upper portion of the cylinder head 10. The drive shaft 11 has a sprocket attached to one end (not shown) and rotates in conjunction with the crankshaft of the engine via a timing chain or the like.

A cylindrical bearing portion 18a of a swing cam 18 that drives an intake valve (or exhaust valve) 19 is fitted on the outer periphery of the drive shaft 11 so as to be relatively rotatable. The swing cam 18 has a thin plate shape having a tip (cam nose) 18b, and a cam surface 18c that slides on an upper surface 19b of a valve lifter 19a as a transmission member provided at the upper end of the intake valve 19 is formed on the outer periphery thereof. Has been.
A ring-shaped eccentric cam 12 is fixed to the outer periphery of the drive shaft 11 by press fitting or the like. The center (axial center) C2 of the eccentric cam 12 is eccentric by a predetermined amount with respect to the center (axial center) C1 of the drive shaft 11. A base portion 13a of a ring-shaped link 13 is fitted on the outer periphery of the eccentric cam 12 so as to be relatively rotatable via a bearing or the like. Note that the swing center (axial center) of the swing cam 18 coincides with the center C1 of the drive shaft 11.

A control shaft 14 extends in the cylinder row direction substantially parallel to the drive shaft 11 obliquely above the drive shaft 11. The control shaft 14 is rotated and held in a predetermined rotation range according to the operating state of the engine by a drive unit 20 described later.
A ring-shaped control cam 15 is fixed to the outer periphery of the control shaft 14 by press fitting or the like. The center (axial center) C4 of the control cam 15 is eccentric by a predetermined amount with respect to the center (axial center) C3 of the control shaft 14. A cylindrical central base of the rocker arm 16 is fitted on the outer periphery of the control cam 15 so as to be relatively rotatable. One end portion 16a of the rocker arm 16 and the small-diameter tip portion 13b of the ring-shaped link 13 are coupled to each other via a first pin 29a that passes through both the 16a and 13b.

  The other end 16 b of the rocker arm 16 and the swing cam 18 are linked by a rod-shaped link 17. More specifically, the operating angle of the control shaft 14 is rotated to the large side, and the other end portion 16b of the rocker arm 16 and the one end portion 17a of the rod-shaped link 17 are the second pins that pass through both the portions 16b and 17a. It is connected via 29b so as to be capable of relative rotation. Further, the other end 17b of the rod-shaped link 17 and the swing cam 18 are coupled to each other via a third pin 29c through which both the ends 17b and 18 are inserted.

Next, the configuration of the drive unit 20 that rotates and holds the control shaft 14 will be described.
As shown in FIG. 1, the control shaft 14 extends into a case 22 fixed to the cylinder head 10, and a worm wheel 21 is fixed to one end thereof. An electric motor (DC motor) 26 driven by a control signal from an ECU (Engine Control Unit) 50 is attached to the case 22, and an output shaft 26 a of the electric motor 26 is connected to the case via a roller bearing 25. 22 extends rotatably. A worm gear 24 that meshes with the worm wheel 21 is fixed to the output shaft 26a. In order to increase the motor torque between the worm gear 24 and the worm wheel 21, the gear ratio is set appropriately large. Further, a rotation angle sensor 23 for detecting the rotation angle of the control shaft 14 (worm wheel 21) is attached to the case 22, and the output of the rotation angle sensor 23 is input to the ECU 50, and the rotation angle sensor The electric motor 26 is feedback-controlled based on the detected rotation angle of the control shaft 14, that is, the detected value of the operating angle (lift amount) of the intake valve 19.

  With such a configuration, when the drive shaft 11 rotates in conjunction with the rotation of the engine, the ring-shaped link 13 moves in translation via the eccentric cam 12, and the rocker arm 16 swings the center C4 of the control cam 15 accordingly. The swing cam 18 swings as a moving center, and the swing cam 18 swings through the rod-shaped link 17. At this time, the cam surface 18c of the swing cam 18 is in sliding contact with the upper surface of the valve lifter 19a as a transmission member provided at the upper end of the intake valve 19, and the valve lifter 19a is pressed against the reaction force of a valve spring (not shown). As a result, the intake valve 19 opens and closes in conjunction with the rotation of the engine.

  Further, when the output shaft 26a of the electric motor 26 is rotationally driven in accordance with the operating state of the engine, the control shaft 14 is rotated via the worm gear 24 and the worm wheel 21, and the control cam serving as the rocking center of the rocker arm 16 is rotated. The position of the center C4 of 15 changes, and the lift characteristic of the intake valve 19 changes continuously. More specifically, the closer the distance between the center C4 of the control cam 15 and the center C1 of the drive shaft 11, the greater the valve lift amount and the operating angle, which are displacements of the lift characteristics.

Next, the operating characteristics of the variable valve mechanism will be discussed.
4 and 5, the reaction force F1 generated by the valve spring reaction force or the like of the intake valve 19 is applied to the other end portion 16b of the rocker arm 16 such as the swing cam 18, the rod-shaped link 17, the second pin 29b, and the like. Acting through. Further, a reaction force F2 generated as a reaction acts on the one end 16a of the rocker arm 16 via the eccentric cam 12, the ring-shaped link 13, the first pin 29a, and the like. Accordingly, the combined reaction force F3 of the reaction forces F1 and F2 substantially acts on the rocking center C4 of the rocker arm 16.

As a result, a torque T1 that is the product of the arm length r1 from the center C3 of the control shaft 14 to the direction line of the combined reaction force F3 and the combined reaction force F3 acts on the control shaft 14. Therefore, in order for the drive unit 20 to hold the control shaft 14 at a predetermined angle, at least a reverse torque that matches the torque T1 is required.
In a state where the control shaft 14 is held at a predetermined rotation angle, as shown in FIG. 4, when the swing cam 18 is pushed down to the highest lift side, that is, when it swings most counterclockwise in FIG. Further, the combined reaction force F3 is maximized. The direction of the resultant reaction force F3 at this time is substantially parallel to the first line L1 connecting the center C1 of the drive shaft 11 and the center C3 of the control shaft 14.

  Here, as shown in FIG. 6, the combined reaction force F3 increases with an increase in the lift amount (operating angle) [FIG. (A): minimum operating angle, (B): intermediate position, (C): The maximum operating angle] increases as the arm length r1 increases from the minimum operating angle to the intermediate position as the eccentric cam 12 rotates, but then decreases. Therefore, the torque T1, which is the product of the combined reaction force F3 and the arm length r1, has a non-linear characteristic with respect to the control shaft operating angle θ.

FIG. 7 shows the relationship between the operating angle θcs of the control shaft 14 and the drive current ics (proportional to the torque T1).
Here, the spring constant K (θcs) for each control shaft operating angle θcs is defined and calculated with respect to the nonlinear control shaft operating angle θcs-driving current ics characteristic, and the spring constant K (θcs) is calculated. ) Is used to control the positioning of the control shaft operating angle θcs while modeling the operating characteristics of the variable valve mechanism that is the control target.

A (θcs) = ics / θcs (1)
K (θcs) = A (θcs) · K _T (2)
However, K _T : Torque constant of the electric motor 26 The coefficient A (θcs) of the drive current ics with respect to the control shaft operating angle θcs is calculated from the characteristic shown in FIG. ) To calculate the spring constant K (θcs) (see FIG. 8). The calculated spring constant K (θcs) is assigned to each control shaft operating angle θcs to create an operating angle-spring constant map shown in FIG. As described above, by defining the spring constant, it is possible to set the operating angle and driving current having non-linear characteristics as continuous and unambiguous functions, and the variable valve mechanism can be modeled as described later. It becomes.

As shown in FIG. 10, this control system is roughly divided into an input limiting unit 101 according to the present invention, a dynamic characteristic compensation unit 102, a responsiveness compensation unit 103, and a variable valve mechanism 104 to be controlled. Composed.
The transmission characteristic G _{P (s)} of the variable valve mechanism 104 can be expressed as a product of the dynamic characteristic and the static characteristic in the 0th order / second order as shown in the following expression (the term on the left side of the right side of the following expression is the dynamic value). Characteristics, right term is static characteristics).

Where J: Moment of inertia of variable valve mechanism
D: Viscous resistance of mechanism as above Based on the above, each element of the present control system shown in FIG. 10 will be described.
The input restriction unit 101 includes an input operation value θ _CST which is an input command value of the operating angle and a final target value, and a sequential target control axis operating angle from the target operating angle calculator 103 a constituting the response compensation unit 103. Enter θ _CSM and apply a rate of change limiter to the ultimate operating angle θ _CST . As a first embodiment of the velocity limiter, and is subjected to first-order lag processing to reach operating angle theta _CST, the deviation between the arrival operating angle theta _CST and the target control shaft operating angle θ _CSM | θ _CST -θ _CSM | Accordingly, the time constant T of the first-order lag processing is set with reference to the map shown in FIG.

In this map, the time constant T is set to a larger value as the deviation | θ _CST −θ _CSM | is larger.
Set with reference to the map shown in.
The limited reaching operating angle θ _CSTLMT limited by the first-order lag process is calculated by the following equation (4).

θ _CSTLMT = 1 / (Ts + 1) · θ _CST (4)
In this map, the time constant T is set to a larger value as the deviation | θ _CST −θ _CSM | is larger.
As a result, when the ultimate operating angle θ _CST changes suddenly and the deviation | θ _CST −θ _CSM | increases, the first-order lag processing of the above equation (4) is performed with a large time constant T, and the delay is increased. The rate of change of the ultimate operating angle is greatly limited.

Next, the dynamic characteristic compensation unit 102 will be described. The dynamic characteristic compensation unit 102 is a so-called feedforward compensator.
Here, it is assumed that the operating angle response (damping ratio ζ, frequency ω) desired by the designer is given by the transfer characteristic GT _(s) of the target operating angle calculator 102a shown in the following equation.

G _FF (s) = G _T (s) so that the actual control shaft operating angle θ _CS follows the dynamic characteristic G _T (s) with respect to the limited reaching operating angle θ _{CSTLMT subjected} to the change rate limiter. ) / G _P (s) The dynamic characteristic compensation output i _CSFF that is the feedforward portion of the drive current is calculated by the following equation (6) from the transfer function G _FF (s) of the dynamic characteristic compensation unit 101 obtained from the relationship To do. That is, the dynamic characteristic compensator 101 is composed of a secondary / secondary filter.

Next, the response compensation unit 103 will be described. The responsiveness compensator 103 includes a target operating angle calculator 103a and a dynamic characteristic output compensator 103b.
The target operation compensator 103a receives the limited reaching operation angle θ _CSTLMT and _calculates a target control shaft operation angle θ _CSM that is a response of the operation angle desired by the designer based on the following equation (7). The target control shaft operating angle θ _CSM is a transient operating angle until the control shaft operating angle θ _CS reaches the final reached operating angle θ _CST .

Ζ and ω in equation (7) are set according to the operating angle response desired by the designer.
In the dynamic characteristic output compensator 103b, the dynamic characteristic compensation output correction value i _CSFB is expressed by the following equation (8) using a filter having an integral characteristic and whose stability is compensated for the parameter change to be controlled. Calculated from the deviation θ _ERR .
θ _ERR = θ _CSM −θ _CS (8)
The following equation (9) is an example of a filter having an integral characteristic. The proportional gain P and the integral gain I are gains whose stability is compensated.

In this case, the dynamic characteristic output compensation value i _CSFB is calculated from the following equation (10).
G _FB (s) = (Ps + I) / s (9)
i _CSFB = (Ps + I) / s · θ _ERR (10)
Using the dynamic characteristic output compensation correction value i _CFSB and the dynamic characteristic compensation output i _CSFF , the current command value i _CSC is calculated from equation (11).
i _CSC = i _CSFB + i _CSFF (11)
By estimating the variable valve mechanism model using the spring constant calculated for each operating angle in this way, and calculating and controlling the current command value i _CSC using the model, the effects of parameter fluctuations and disturbances can be reduced. It is difficult to receive, and the operating angle response desired by the designer can be obtained.

In particular, the limited reaching operation angle θ _CSTLMT subjected to the first-order lag processing in which the time constant T is variably set according to the deviation | θ _CST −θ _CSM | between the reaching operation angle θ _CST and the target control shaft operation angle θ _CSM By calculating the dynamic characteristic compensation output i _CSFF and the target control shaft operating angle θ _{CSM and} thus the dynamic characteristic output compensation correction value i _CFSB , it is possible within the range that can be realized without saturating the electric motor 26 current. As long as possible, control with good responsiveness following the normative response can be performed.

FIG. 12 shows the response characteristic of the control shaft operating angle θ _CS when the ultimate operating angle θ _CST changes stepwise.
As a second embodiment, as shown in FIG. 13, the time constant T of the first-order lag processing is set according to the deviation | θ _CST −θ _CS | between the ultimate operating angle θ _CST and the actual control shaft operating angle θ _CS. It is good also as a structure to set, and the same effect is acquired. FIG. 14 shows a control block diagram of the second embodiment.

Further, as a third embodiment, the rate of change limiter with respect to the ultimate operating angle θ _CST may be variably set according to the operating angle-driving current characteristics. Specifically, the time constant T of the first-order lag process with respect to the ultimate operating angle θ _CST is set with reference to the map shown in FIG. 16 according to the operating angle-driving current characteristic shown in FIG.
Here, the map shows a time constant T in a region where the gain of the operating angle-driving current is large, that is, in a region where the amount of change in driving current with respect to the amount of change in operating angle is large (the slope of the driving current characteristic shown in FIG. 15 is large). Is set to a large value to slow the response, and in a region where the gain is small and the amount of change in the drive current with respect to the change in operating angle is small, the time constant T is set to a small value so as to speed up the response. The control block diagram of the third embodiment is as shown in FIG. 14 as in the second embodiment.

In this way, the drive current of the electric motor 26 is not saturated, and control with good responsiveness following the reference response can be performed while controlling at the maximum response speed that can be realized for each operating angle. This is particularly effective for a variable valve mechanism 104 having a nonlinear relationship between the displacement of the lift characteristic and the driving force.
Further, in the above embodiment, the change rate limiter is set by a time constant for first-order lag processing of the arrival operating angle, but may be simply set by a change rate or change amount of the reaching operation angle.
In addition to the variable valve mechanism described above, the present invention can be used as a variable valve mechanism as long as it is a control device that controls the displacement of a link mechanism linked to a spring by driving an actuator against the biasing force of the spring. It can be applied in the same manner as applied, and the same effect can be obtained.

The top view of the variable valve mechanism based on this invention. Sectional drawing of the principal part of a variable valve mechanism same as the above. The block diagram which shows the drive part of the said variable valve mechanism. The figure for demonstrating the effect | action of the said variable valve mechanism. The block diagram which shows the control shaft and control cam of the said variable valve mechanism. The figure for demonstrating the effect | action in each rotation angle of the said control shaft. The figure which shows the relationship between the operating angle of the said variable valve mechanism, and a drive current. The figure for demonstrating calculation of the spring constant of the said variable valve mechanism. Basic characteristic map of spring constant. The control block diagram of 1st Embodiment in the control apparatus of the said variable valve mechanism. The time constant setting map of the first-order lag processing in the first embodiment. The figure which shows the response characteristic in 1st Embodiment. The time constant setting map in the second embodiment. The control block diagram of 2nd Embodiment in the control apparatus of the said variable valve mechanism. The figure which shows the area | region where the time constant setting of a primary delay process is divided in the relationship between the operating angle of a variable valve mechanism and drive current in 3rd Embodiment. Map with different time constants for each area. The figure which shows the response characteristic when not having the structure of this invention.

Explanation of symbols

11 Drive shaft,
DESCRIPTION OF SYMBOLS 12 Eccentric cam 13 Ring-shaped link 14 Control shaft 15 Control cam 16 Rocker arm 17 Rod-shaped link 18 Oscillation cam 19 Intake valve 20 Drive part 50 ECU
101 Input Limiting Unit 102 Dynamic Characteristic Compensation Unit 103 Responsiveness Correction Unit 104 Variable Valve Mechanism

Claims (12)

In a control device for a variable valve mechanism in an internal combustion engine that controls the lift characteristics of an engine valve by driving an actuator against a reaction force of the valve spring, and a link mechanism linked to the valve spring.
A change rate limiter that is variably set according to a parameter representing a change state of the displacement is applied to the ultimate displacement of the final target of the lift characteristic to limit the change rate, and the change rate is based on the limited ultimate displacement. A control apparatus for a variable valve mechanism, wherein feedback control is performed so as to obtain a desired response characteristic.   2. The control apparatus for a variable valve mechanism according to claim 1, wherein a feedforward compensation amount corresponding to the desired response characteristic is set.   The control device for a variable valve mechanism according to claim 1 or 2, wherein a transient target displacement corresponding to the desired response characteristic is set.   4. The control device for a variable valve mechanism according to claim 3, wherein the change rate limiter is variably set according to a deviation between the ultimate displacement and a transient target displacement.   The control device for a variable valve mechanism according to any one of claims 1 to 3, wherein the change rate limiter is variably set according to a deviation between the ultimate displacement and the actual displacement. .   4. The control device for a variable valve mechanism according to claim 1, wherein the change rate limiter is variably set according to the characteristics of the lift characteristic displacement and the driving force. .   The variable motion limiter according to claim 6, wherein the change rate limiter is set to a small value in a region where a gain of a driving force change amount with respect to the lift characteristic displacement change amount is large, and to a large value in a region where the gain is small. Control device for valve mechanism.   The control device for a variable valve mechanism according to any one of claims 1 to 7, wherein the change rate limiter is set by a time constant that performs first-order lag processing on the ultimate displacement of the lift characteristic. .   The variable valve mechanism control apparatus according to any one of claims 1 to 8, wherein the variable valve mechanism has a characteristic in which a relationship between a displacement of a lift characteristic and a driving force is nonlinear. .   The control device for a variable valve mechanism according to any one of claims 1 to 9, wherein the actuator is an electric motor, and the driving force is controlled by an energization current of the electric motor. The variable valve mechanism is
A drive shaft that rotates in conjunction with the rotation of the engine,
A swing cam that is fitted on the outer periphery of the drive shaft so as to be relatively rotatable, and drives the engine valve to open and close;
An eccentric cam fixed eccentrically on the outer periphery of the drive shaft;
A ring-shaped link that is fitted on the outer periphery of the eccentric cam so as to be relatively rotatable;
A control shaft extending substantially parallel to the drive shaft;
A control cam eccentrically fixed to the outer periphery of the control shaft;
A rocker arm that is fitted on the outer periphery of the control cam so as to be relatively rotatable, and is linked to the ring-shaped link at one end thereof;
A rod-shaped link that links the other end of the rocker arm and the swing cam;
12. The lift according to claim 1, wherein when the actuator rotates the control shaft, a rocking center position of the rocker arm is changed and the lift characteristic is changed. Control device for variable valve mechanism.   A control device that controls the displacement of a link mechanism linked to a spring by driving an actuator against the biasing force of the spring, and represents a change state of the displacement with respect to a final displacement of the link mechanism. A rate of change limiter that is variably set according to a parameter is applied to limit the rate of change, and the rate of change is feedback-controlled so that a desired response characteristic is obtained based on the limited ultimate displacement. Control device.

Images