JP2007100593A - Controller of variable valve gear mechanism - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To suppress the effect of engine noise on feedback control by accurately estimating the displacement of valve characteristics with high response while estimating an offset current of an actuator in a variable valve gear mechanism variably controlling the characteristics of the engine valves of an internal combustion engine. <P>SOLUTION: An observer is formed based on a transfer function model using the spring constant of the variable valve gear mechanism and the offset current of the actuator. A value obtained by integrating a deviation between an actually measured value θcs and an estimated value θcso of a control shaft operating angle is converted by the observer and fed back while an offset current is estimated to accurately provide the estimated value θcso of a control shaft operating angle. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a control device for a variable valve mechanism that continuously and variably controls lift characteristics (working angle, lift amount) of intake and exhaust valves of an internal combustion engine, and in particular, control accuracy by avoiding the influence of engine vibration noise It relates to technology to improve.

As described in Patent Document 1, in order to ensure sufficient output by improving fuel efficiency at the time of engine low speed and low load, stable drivability, and improvement of intake charge efficiency at high speed and high load, etc. There has been proposed a variable valve mechanism that variably feedback-controls the valve lift characteristics in accordance with engine operating conditions.
JP 2001-3773 A

In this type of feedback control of the variable valve mechanism disclosed in Patent Document 1, vibration noise synchronized with engine rotation is input to a detection unit for lift characteristics (operating angle, lift amount), and detection with the added noise is detected. If a value is used, the feedback correction amount amplifies the noise component and the lift characteristic is controlled in a vibrational manner, so that the feedback gain cannot be set large, and there is a limit to improving the responsiveness.
Moreover, since the rotational speed of the drive motor for the variable valve mechanism cannot be accurately detected due to noise, it is difficult to estimate the back electromotive voltage, which also leads to a decrease in control accuracy.

For this reason, conventionally, noise is removed using a filter (elimination filter) that reduces the gain of the frequency band corresponding to the noise.
However, in this method, when a control response frequency band that is the same as the noise frequency band is requested, the detection of the lift characteristic is delayed, so that the control performance is significantly deteriorated. For example, when the engine rotational speed is 600 rpm, the vibration frequency of noise is 5 Hz, but the case where 1 Hz to 6 Hz is required as the frequency of the normative response at this time.

  The present invention has been made paying attention to such a conventional problem, and estimates a lift characteristic displacement from which engine vibration noise has been removed with high accuracy and high response, and further uses this estimated value to make a variable valve actuation. The purpose is to improve the feedback control accuracy of the mechanism.

  In order to solve the above-described problems, the present invention provides a link mechanism connected to a valve spring to drive an electric actuator against the reaction force of the valve spring to variably control the lift characteristics of the engine valve, and A control apparatus for a variable valve mechanism in an internal combustion engine having a non-linear relationship between a displacement of a lift characteristic and a driving force, the spring constant calculated as the driving force / displacement for each variable lift characteristic, The observer is configured by a variable valve mechanism model using the offset current of the electric actuator, and the displacement of the lift characteristic is estimated by the observer while estimating the offset current.

  With such a configuration, it is possible to estimate the displacement of the lift characteristic that is not affected by engine vibration noise while estimating the dynamic characteristic and static characteristic of the variable valve mechanism by the observer configured using the model of the variable valve mechanism. In addition, by performing feedback control using a lift characteristic displacement that is not affected by the engine vibration noise, stable and highly responsive control performance can be ensured.

  Further, since the offset current of the electric actuator can be estimated separately by the observer, the lift characteristic displacement can be estimated with higher response.

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. The other end 17b of the rod-shaped link 17 and the swing cam 18 are connected to each other via a third pin 29c that passes through both the ends 17b and 18 so as to be relatively rotatable.

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 rotation angle of the control shaft 14 (control shaft operating angle θ _CS ) detected at 23, that is, the detected value of the lift characteristic displacement (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 via 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 26 a of the electric motor 26 is driven to rotate according to the operating state of the engine, the control shaft 14 rotates through the worm gear 24 and the worm wheel 21, and the control cam becomes the rocking center of the rocker arm 16. 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 from the minimum operating angle to the intermediate position as the eccentric cam 12 rotates, but decreases thereafter. 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 θ _CS of the control shaft 14 and the drive current i _CS (proportional to the torque T1).
Here, the spring constant K (θ _CS ) for each control shaft operating angle θ _CS is defined and calculated for the nonlinear control shaft operating angle θ _CS −drive current i _CS characteristic, and the spring Positioning control of the control shaft operating angle θ _CS is performed while modeling the operating characteristics of the variable valve mechanism to be controlled using the constant K (θ _CS ).

A (θ _CS ) = i _CS / θ _CS (1)
K (θ _CS ) = A (θ _CS ) · K _T (2)
However, K _T: calculating the coefficients of the drive current i _CS for the control shaft operating angle theta _CS than the characteristics of torque constant 7 (1) of the electric motor 26 A (theta _CS), to the A (theta _CS) Based on the equation (2), the spring constant K (θ _CS ) is calculated (see FIG. 8). The calculated spring constant K (θ _CS ) is assigned to each control shaft operating angle θ _CS 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 a continuous and unambiguous function, and the variable valve mechanism can be modeled as described later. It becomes.

As shown in FIG. 10, the present control system is roughly divided into a dynamic characteristic compensator 101, a responsiveness compensator 102, a control shaft operating angle estimator 103 according to the present invention, and a variable valve mechanism that is a control target. 104.
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: Viscosity resistance of the same mechanism Assuming that the input of the variable valve mechanism 104 is the current command value i _CSC and the output is the control shaft operating angle θ _CS , the relationship between the current command value i _CSC and the control shaft operating angle θ _CS is It can be expressed as

Based on the above, each element of the present control system shown in FIG. 10 will be described.
First, the dynamic characteristic compensation unit 101 will be described. The dynamic characteristic compensation unit 101 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) / G so that the actual control shaft operating angle θ _CS follows the input characteristic and final target value θ _{CST as the} dynamic characteristic G _T (s). Based on the transfer function G _FF (s) of the dynamic characteristic compensator 101 obtained from the relationship of G _P (s), a dynamic characteristic compensation output i _CSFF that is a feed forward portion of the drive current is calculated according to the following equation (6). . That is, the dynamic characteristic compensator 101 is composed of a secondary / secondary filter.

Next, the response compensation unit 102 will be described. The response compensation unit 102 includes a target operating angle calculator 102a and a dynamic characteristic compensation output corrector 102b.
The target operation compensator 102a receives the final operation angle θ _CST as input, and calculates the target control shaft operation angle θ _CSM , which is a response of the operation angle desired by the designer, based on the 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 compensation output corrector 102b, a dynamic characteristic compensation output correction value i _CSFB is calculated using a filter having an integral characteristic and whose stability is compensated with respect to a parameter change to be controlled. Calculated from the quantity θ _ERR .
θ _ERR = θ _CSM −θ _CSO (8)
Here, instead of the detected value θ _CS at the sensor, the value θ _CSO estimated by the observer as described later is used as the control shaft operating angle.

Equation (9) is an example of a filter having integral characteristics. The proportional gain P and the integral gain I are gains whose stability is compensated. In this case, the dynamic characteristic compensation output correction value i _CSFB is calculated from the equation (10).
G _FB (s) = (Ps + I) / s (9)
i _CSFB = (Ps + I) / s · θ _ERR (10)
Based on the dynamic characteristic compensation output correction value i _CFSB , the current command value i _CSC is calculated from the equation (11) using the dynamic characteristic compensation output i _CSFF .
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.

Next, the control axis operating angle estimator (observer) 103 according to the present invention will be described. The control shaft operating angle estimator 103 is configured based on the variable valve mechanism model represented by the above formula (3), and corrects the model error based on the detected value θ _CS of the control shaft operating angle. The estimated value θ _CSO of the control shaft operating angle from which noise contained in the detected value θ _CS is removed is obtained.
By the way, as shown in FIG. 11, the offset current required for the operating angle to start to move due to the spring set load or friction of the variable valve mechanism 104 differs for each engine speed, and the control shaft operating angle increases. It has hysteresis characteristics in the direction and the decreasing direction. Even if the offset current changes, the control shaft operating angle can be estimated by the observer, but the estimation of the control shaft operating angle is delayed by the amount accompanied by the feedback correction according to the change of the offset current.

Therefore, the control axis operating angle estimator 103 according to the present invention has a function of separately estimating the offset current, thereby enabling the control axis operating angle to be estimated with higher response.
As shown in the block diagram of FIG. 12, the control axis operating angle estimator 103 _receives the current command value i _CSC and the control axis operating angle θ _CS detected by the rotation angle sensor 23, and _receives the control axis operating angle estimated value θ. Calculate _CSO .

A method of calculating the gain (observer gain) k of the control axis operating angle estimator 103 will be described below.
First, assuming that the offset current ofseti is present in the equation (4), the following relationship can be derived.

Here, if the state variables x ₁ , x ₂ , x ₃ are defined as follows,

From the equations (13) and (14), the following relationship is derived.

From the equations (14) and (15), the variable valve mechanism is expressed by the following equation of state.

x, A, B, and C are defined as in equation (17), and equation (16) is defined as equation (18).

In FIG. 13, when the feedback gain k of the deviation e between the estimated value θ _CSO of the control shaft operating angle and the detected value θ _CS is defined by the equation (19), the estimated value x _O of the state quantity x is (20 ) Expression.

The error between x _O and x is e,

By setting the gain k (k1, k2, k3) that stabilizes A-kC, the pole of the control axis operating angle estimator 103 is set to an appropriate value, and the control axis operating angle estimated value θ with high accuracy is set. You can get _CSO .
As for gains k1 and k2, in high speed regions where the frequency of noise generated in synchronization with engine rotation is high, increasing the gain and increasing the pole of the observer increases the response as much as possible while eliminating noise. If the response is increased in the low-speed region where the frequency is low, it becomes more susceptible to noise. Therefore, the gain is made smaller than the high-speed region so that the influence of noise can be removed. Thereby, only noise caused by engine rotation can be removed, and other modeling errors and the like can be followed promptly.
Further, since the peak value of the noise increases as the control shaft operating angle θ _CS increases, the influence of noise can be eliminated by reducing the gains k1 and k2 when the valve operating angle is large compared to when the valve operating angle is small.
However, the pole of the control shaft operating angle estimator 103 has a larger value than the pole of the feedback control system of the variable valve mechanism (increased by the decrease of the spring constant according to the increase of the control shaft operating angle). The minimum value of the pole arrangement is defined by regulating the lower limit values of the gains k1 and k2. Note that k1 and k2 may be substantially the same value.
On the other hand, the gain k3 for estimating the offset current ofseti may be simply a fixed value, but if the offset current is large according to the direction of increase / decrease of the engine rotation speed and the control shaft operating angle according to the characteristics of FIG. When predicted, the value may be set variably, for example, to be a large value.
According to the control axis operation angle estimator 103 configured as described above, it is possible to obtain the control axis operation angle estimated value θ _{CSO from} which the influence of noise is removed while correcting the modeling error. The estimated value θ _CSO can be used in place of the detected value θ _CS to obtain highly accurate feedback control performance.
Further, since the control axis operating angle estimator 103 estimates the control axis operating angle by separating and estimating the offset current ofseti at the portion indicated by the dotted line in the figure, the control axis operating due to the change in the offset current ofseti is performed. The angle estimation delay can be suppressed, the control shaft operating angle estimated value θ _CSO can be obtained with high response, and feedback control performance with high accuracy and high response can be obtained.
Next, a second embodiment of the present invention will be described.
In the present embodiment, as shown in the block diagram of FIG. 13, an offset current estimator 105 that estimates the offset current ofseti according to the increase / decrease direction of the engine rotation speed and the control shaft operating angle is provided. The offset current ofseti estimated in (1) is given to the control axis operating angle estimator 103 as a feedforward value ofseti _FF . Specifically, a map in which the feedforward value ofseti _FF shown in FIG. 14 is set based on the characteristics shown in FIG. 11 may be provided, and the search may be performed according to the increasing / decreasing direction of the engine speed and the control shaft operating angle. .
In this way, since the offset current ofseti in the control axis operating angle estimator 103 can be estimated only by correcting the feedforward value ofseti _FF , it can be estimated while correcting the offset current ofseti with higher response. As a result, the control shaft operating angle estimated value θ _CSO can be estimated with higher response and higher accuracy, and the feedback control performance can be further improved.
In the present embodiment, as shown in FIG. 15, the feedforward value ofseti _FF may be given to the feedback control system. Thereby, the current command value i _CSC is calculated from the equation (22).
i _CSC = i _CSFB + i _CSFF + ofseti _FF (22)
In this way, in the first embodiment, the offset current ofseti is included in the dynamic characteristic compensation output correction value i _CSFB that is the feedback correction amount, whereas only the correction of the feedforward value ofseti _FF is the dynamic characteristic compensation. Since it is included in the output correction value i _CSFB , the dynamic characteristic compensation output correction value i _CSFB can be corrected with high response to a change in the offset current.
FIG. 16 shows a first embodiment using an observer for separately estimating the offset current, FIG. 17 shows a second embodiment in which the offset current is further estimated while giving a feedforward value, and FIG. 18 shows an offset Using a normal observer that does not perform separate estimation of current, actual measurement of control shaft operating angle when stepping operating angle θ _CST is changed stepwise while there is no vibration noise and offset current is generated The value θ _CS and the estimated value θ _CSO are compared and shown. Although the original purpose of the observer is to estimate the control shaft operating angle from which noise has been removed, here, in order to compare only the influence of the offset current, the comparison is performed without vibration noise.
When there is no vibration noise, the estimated value θ _CSO should match the measured value θ _CS , but in FIG. 18, the estimated value θ _CSO and the detected value θ _CS are greatly different due to the influence of the offset current. .
In contrast, the first embodiment of FIG. 16 has a good estimation function although there is a slight error between the estimated value θ _CSO and the actually measured value θ _CS .
Further, in the second embodiment of FIG. 17, the estimated value θ _CSO and the actually measured value θ _CS substantially coincide with each other, and a better estimation function is provided.

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 figure which shows the difference in the offset electric current by the engine rotational speed change in the said variable valve mechanism, and the control shaft working angle increase / decrease direction. The block diagram which shows the structure of the control-axis working angle estimator of embodiment same as the above. The block diagram which shows the structure of the control-axis working angle estimator of 2nd Embodiment. The characteristic map of the feedforward value of the offset current used in the second embodiment. The block diagram which shows the structure of the control-axis working angle estimator of 2nd Embodiment. The figure which shows measured value (theta) _CS and estimated value (theta) _CSO of a control-axis operating angle when the reach _| attainment operating angle (theta) _CST is changed in steps in a state without vibration noise in 1st Embodiment. The figure which shows measured value (theta) _CS and estimated value (theta) _CSO of a control-axis operating angle when the reach _| attainment operating angle (theta) _CST is changed in steps in a state without vibration noise in 2nd Embodiment. An observer that does not have a function to separate and estimate the offset current, and the actual operating value θ _CS and the estimated value θ of the control shaft operating angle when the ultimate operating angle θ _CST is changed stepwise without vibration noise. _The figure which shows _CSO .

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
DESCRIPTION OF SYMBOLS 101 Dynamic characteristic compensation part 102 Responsibility compensation part 103 Control-axis working angle estimator 104 Variable valve mechanism 105 Offset current estimator

Claims (8)

The link mechanism linked to the valve spring drives the electric actuator against the reaction force of the valve spring to variably control the lift characteristics of the engine valve, and the relationship between the displacement of the lift characteristics and the driving force is nonlinear. A control device for a variable valve mechanism in an internal combustion engine having various characteristics,
An observer is configured by a variable valve mechanism model using a spring constant calculated as driving force / displacement for each variable lift characteristic and the offset current of the electric actuator, and the offset current is estimated by the observer. A control device for a variable valve mechanism, wherein the displacement of the lift characteristic is estimated. The link mechanism linked to the valve spring drives the electric actuator against the reaction force of the valve spring to variably control the lift characteristics of the engine valve, and the relationship between the displacement of the lift characteristics and the driving force is nonlinear. A control device for a variable valve mechanism in an internal combustion engine having various characteristics,
A variable valve mechanism model using a spring constant calculated as a driving force / displacement for each variable lift characteristic and an offset current of the electric actuator, and estimating the offset current A control apparatus for a variable valve mechanism, wherein an observer for estimating displacement is provided.   The control apparatus for a variable valve mechanism according to claim 1 or 2, wherein a lift characteristic displacement obtained by removing noise by the observer is estimated while inputting a lift displacement detection value detected by a sensor.   The control apparatus for a variable valve mechanism according to any one of claims 1 to 3, wherein feedback control is performed using a displacement of a lift characteristic estimated by the observer.   The observer estimates an offset current based on a value obtained by integrating a deviation between an estimated value of a lift characteristic displacement and a detected value, and estimates the lift characteristic displacement while feeding back the estimated value of the offset current. The control apparatus of the variable valve mechanism as described in any one of Claims 1-3.   6. The control device for a variable valve mechanism according to claim 5, wherein a feedback gain of the estimated value of the offset current is variably set according to a change direction of an engine speed and a lift characteristic displacement.   The variable motion according to any one of claims 1 to 6, wherein a feedforward value of an offset current estimated based on a change direction of an engine rotational speed and a lift characteristic displacement is input to the observer. Control device for valve mechanism.   The variable valve mechanism is rotationally driven by a crankshaft of an engine, a drive shaft having a drive cam fixed to the outer periphery thereof, a rocker arm that swings by rotation of the drive cam linked to one end, and the other end of the rocker arm A rocking cam that opens the engine valve in linkage with a portion is provided, and a rocking fulcrum of the rocker arm is changed by drive control of the motor. A drive control device for a variable valve mechanism according to claim 1.

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