JP4572794B2 - Control device for variable valve mechanism - Google Patents

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 conventional problems, and estimates the lift characteristics from which engine vibration noise has been removed, and further uses this estimated value to improve the feedback control accuracy of the variable valve mechanism. The purpose is to let you.

  In order to solve the above-described problems, the present invention provides a link mechanism connected to a valve spring to drive the 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 characteristic displacement and a driving force, wherein the variable valve mechanism model is defined as a driving force / displacement for each variable lift characteristic. It is estimated using the calculated spring constant, and the displacement of the lift characteristic is estimated by an observer configured using the model.

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. it can.
Further, 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.

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 is formed on the outer periphery of the swing cam 18 so as to be in sliding contact with an upper surface 19b of a valve lifter 19a as a transmission member provided at an upper end of the intake valve 19. 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 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 the 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 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, the response of the operating angle desired by the designer (damping ratio ζ, frequency ω) is the target operation shown in the following equation

It is assumed that it is given by the transfer characteristic GT _(S) of the angle calculator 102a.

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), the dynamic characteristic compensation output i _CSFF that is the feed forward amount of the drive current is calculated by 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 arrival operation angle θ _CST as an input, 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 equation (6). 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 basic current command value i _CSC0 is calculated from the equation (11) using the dynamic characteristic compensation output i _CSFF .
i _CSC0 = i _CSFB + i _CSFF (11)
Further, as shown in FIG. 11, since the current required for the operating angle of the variable valve mechanism to start to move varies depending on the engine rotational speed, the current shown in FIG. _Is added to the basic current command value i _CSC0 to calculate the final current command value i _CSC .

i _CSC = i _CSC0 + ofset i = i _CSFB + i _CSFF + ofset i (12)
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.

As shown in the block diagram of FIG. 13, 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 inputs 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, the following relationship is derived from Equation (3).

Here, if state variables x ₁ and x ₂ 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) 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 _CSO with high accuracy is set. Obtainable.
The gain k is set based on the engine speed and the control shaft operating angle θ _CS (the valve operating angle of the engine valve). FIG. 14 shows a characteristic of the gain k with respect to the engine rotational speed Ne. As shown in the figure, the value of the gain k is larger in the high speed region of the engine rotation speed Ne than in the low speed region.
FIG. 15 shows the characteristic of the gain k with respect to the control shaft operating angle θ _CS . As shown in the figure, the value of the gain k is decreased as the control shaft operating angle θ _CS is increased.
In order to set the gain k in consideration of both characteristics with respect to the engine speed and the control shaft operating angle θ _CS , for example, the search value k _N from the map having the characteristics shown in FIG. 14 and the characteristics shown in FIG. The gain set in the three-dimensional map using the engine rotation speed Ne and the control shaft operating angle θ _CS as parameters is searched for. In the present embodiment, k1 and k2 are set to substantially the same value.
That is, basically, the estimation accuracy of the control shaft operating angle is corrected while the modeling error of the control shaft operating angle estimator 103 (observer) configured by the variable valve mechanism model is quickly corrected based on the deviation e. Therefore, we want to increase the gain k and increase the observer pole to improve the response. However, if the responsiveness is too high, it is erroneously corrected so as to follow the fluctuation of the noise, and the noise cannot be removed.
Since noise is generated in synchronization with engine rotation, in a high speed range where the noise frequency is high, increasing the gain k and increasing the pole of the observer improves the responsiveness as much as possible while removing the noise, and the noise frequency is low If the response is increased in the low speed range, it becomes more susceptible to noise, so the gain is made smaller than in the high speed range 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 noise increases as the control shaft operating angle θ _CS increases, the influence of noise can be eliminated by reducing the gain k when the valve operating angle is large compared to when the valve operating angle is small.
However, as shown in FIG. 16, the pole of the control shaft operating angle estimator 103 is larger than the pole of the feedback control system of the variable valve mechanism (increasing due to 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 restricting the lower limit value of the gain k so as to be a large value (left half plane).
In this way, by making the pole of the control axis operating angle estimator 103 larger than the pole of the feedback control system, the control error of the control axis can be corrected faster than the feedback control response, and the influence of noise can be eliminated. By using the estimated value θ _CSO , highly accurate feedback control performance can be obtained.
FIG. 17 shows the present embodiment device, FIG. 18 shows a feedback control device that uses a detected control axis operating angle detection value from which noise is not removed, and FIG. 19 shows a conventional device that uses a control shaft operating angle detection value that has passed through a filter. The response of the control shaft operating angle θ _CS and the motor driving current i _CS when the ultimate operating angle θ _CST is changed stepwise in the operation state where the vibration noise frequency of the reference response is the same are shown.
In FIG. 18, the drive current is oscillating in synchronization with the noise, and the actual control shaft operating angle controlled accordingly is also oscillated. In FIG. 19, the vibration of the motor current can be suppressed by the filtering process, but the control shaft operating angle θ _CS overshoots the normative response. In the case of the present embodiment apparatus of FIG. 17, it is clear that the vibration of the motor current is suppressed, the vibration of the actual control shaft operating angle can be suppressed, and the control shaft operating angle θ _CS follows the normative response well. It is.

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 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 speed change in the said variable valve mechanism. The characteristic map of the offset current used by embodiment same as the above. The block diagram which shows the structure of the control-axis working angle estimator of embodiment same as the above. The figure which shows the characteristic of the observer gain with respect to the engine speed of embodiment same as the above. The figure which shows the characteristic of the observer gain with respect to the control-axis operating angle of embodiment same as the above. The figure which shows pole arrangement | positioning of the observer with respect to the feedback control system of embodiment same as the above. In according to the exemplary embodiment, it shows the response of the control shaft operating angle theta _CS in case of stepwise varying the reach operating angle theta _CST, the driving current i _CS of the motor. The control device that does not noise removal, shows the response of the control shaft operating angle theta _CS in case of stepwise varying the reach operating angle theta _CST, the driving current i _CS of the motor. In the control apparatus for noise removal filter, it shows the response of the control shaft operating angle theta _CS in case of stepwise varying the reach operating angle theta _CST, the driving current i _CS of the motor.

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 Responsiveness compensation part 103 Control-shaft operating angle estimator 104 Variable valve mechanism

Claims (9)

  The link mechanism linked to the valve spring drives the actuator against the reaction force of the valve spring to variably control the lift characteristic of the engine valve, and the relationship between the displacement of the lift characteristic and the driving force is nonlinear. An apparatus for controlling a variable valve mechanism in an internal combustion engine having characteristics, wherein the model of the variable valve mechanism is estimated using a spring constant calculated as a driving force / displacement for each variable lift characteristic, and the model A control apparatus for a variable valve mechanism, wherein the displacement of the lift characteristic is estimated by an observer configured using the above-mentioned.   The link mechanism linked to the valve spring drives the actuator against the reaction force of the valve spring to variably control the lift characteristic of the engine valve, and the relationship between the displacement of the lift characteristic and the driving force is nonlinear. An apparatus for controlling a variable valve mechanism in an internal combustion engine having characteristics, wherein the model of the variable valve mechanism is estimated using a spring constant calculated as a driving force / displacement for each variable lift characteristic, and the model A control device for a variable valve mechanism, characterized in that an observer is provided for estimating the displacement of the lift characteristic by using.   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 control device for a variable valve mechanism according to any one of claims 1 to 4, wherein the pole of the observer is set based on at least one of engine rotational speed and displacement of lift characteristics.   6. The control of the variable valve mechanism according to claim 5, wherein the pole of the observer is made larger as the engine rotational speed increases and made smaller as the lift characteristic displacement (valve operating angle) increases. apparatus.   7. The control device for a variable valve mechanism according to claim 5, wherein the pole of the observer is set to a larger value than the pole of the feedback control system determined for each lift characteristic displacement.   The observer converts the drive current of the actuator into a lift characteristic displacement by a model transfer function of a variable valve mechanism, and corrects it according to a deviation from a detected value of the lift characteristic displacement to estimate the lift characteristic displacement. The control device for a variable valve mechanism according to any one of claims 1 to 7, wherein the control device is configured.   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 9. A rocking cam for opening the engine valve in linkage with the portion, 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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