JP4595553B2 - Control device for variable valve mechanism - Google Patents

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 this type of variable valve mechanism, the one disclosed in Patent Document 1 performs control by estimating a non-linear characteristic reaction torque acting on the control shaft of the variable valve mechanism as a disturbance by the reaction force of the valve spring. Yes.
JP 2001-3773 A

In Patent Document 1, the relationship between the stability of the feedback control system and the reaction torque cannot be clarified, and variations in the feedback control system and the influence of secular change cannot be easily handled.
Further, in the stability analysis, a round transfer function obtained by multiplying the transfer function of the control object and the transfer function of the feedback compensator is required. However, in Patent Document 1, the reaction force torque is estimated as a disturbance, and the control object is controlled. Since the reaction force torque action is not incorporated in the (variable valve mechanism) model, even if the transfer function of the controlled object that is not correctly modeled is used to obtain the one-round transfer function as described above, the stability analysis is correctly performed. I can't do it.

  The present invention has been made paying attention to such a conventional problem, and an object of the present invention is to model a variable valve mechanism so that high-precision control can be performed.

In order to solve the above-described problems, the present invention provides an operation characteristic model of a variable valve mechanism having a nonlinear relationship between the displacement of the lift characteristic and the driving force, for each actuator operating angle. Force / estimated using the spring constant calculated as the operating angle of the control axis, and using the model as a target value of the operating angle of the control axis, the operating angle of the control axis is determined using the model. The actuator is controlled to follow the target value.

  With such a configuration, by modeling the variable valve mechanism using the spring constant calculated for each lift characteristic as described above, the reaction force from the valve spring over the entire variable range of the valve characteristic. Therefore, it is possible to design a model that accurately represents the characteristics corresponding to the above, so that it is difficult to be influenced by parameter fluctuations and disturbances, and a response of the lift characteristics desired by the designer can be obtained.

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 other end portion 16b of the rocker arm 16 and the one end portion 17a of the rod-like link 17 are coupled to each other via a second pin 29b that passes through both the portions 16b and 17a so as to be relatively rotatable. 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 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 rotates into the case 22 via a roller bearing 25. It extends as possible. 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 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, as the operating angle of the control shaft 14 is rotated to the larger side and the distance between the center C4 of the control cam 15 and the center C1 of the drive shaft 11 is closer, the valve lift amount, which is the displacement of the lift characteristics, and The operating angle increases.

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 nonlinear characteristic with respect to the operating angle θ.

FIG. 7 shows the relationship between the operating angle θcs of the control shaft (corresponding to the lift amount of the intake valve and the operating angle) and the drive current ics (proportional to the torque T1).
Here, in the present invention, the spring constant K (θcs) for each control shaft operating angle θcs is defined and calculated with respect to the non-linear control shaft operating angle θcs−driving current ics characteristic, and the spring is calculated. Positioning control of the control shaft operating angle θcs is performed while modeling the operation characteristics of the variable valve mechanism to be controlled using the constant K (θcs).

A (θcs) = ics / θcs (1)
K (θcs) = A (θcs) · K _T (2)
However, K _T : DC motor torque constant 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. 7 using the equation (1), and based on the A (θcs), (2) The spring constant K (θcs) is calculated from the equation (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. By defining the spring constant as described above, it is possible to set the control shaft operating angle and drive current having non-linear characteristics as a continuous and unique function, and to model the variable valve mechanism as described later Is possible.

As shown in FIG. 10, this control system is roughly composed of a dynamic characteristic compensator 101, a responsiveness compensator 102, and a variable valve mechanism 103 to be controlled.
The transmission characteristic G _{P (s)} of the variable valve mechanism 103 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.
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), the dynamic characteristic compensation output i _CSFF that is the feed forward amount of the drive current is calculated by the following equation (5). 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 output compensator 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 (6) are set according to the operating angle response desired by the designer.
In the dynamic characteristic output compensator 102b, a dynamic characteristic compensated output correction value i _CSFB is calculated by using a filter having an integral characteristic and whose stability is compensated with respect to a parameter change to be controlled. Calculated from θ _ERR .
θ _ERR = θ _CSM −θ _CS (7)
As an example of a filter having integral characteristics, equation (8) can be given. 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 equation (9).

G _FB (s) = (Ps + I) / s (8)
i _CSFB = (Ps + I) / s · θ _ERR (9)
Based on the dynamic characteristic output compensation correction value i _CFSB , the current command value i _CSC is calculated from the equation (10) using the dynamic characteristic compensation output i _CSFF .
i _CSC = i _CSFB + i _CSFF (10)
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 addition, for example, without using the spring constant defined as in the present invention, a region having a linearity with a small operating angle and a region having a nonlinearity higher than the nonlinearity shown in FIG. Although it is possible to divide the drive current into multiple areas, such as a region where the drive current increases and decreases including the maximum point, and a region where the drive current decreases after that, design different models and control while switching models, but it is extremely complicated Obviously, it is difficult to achieve continuity at the switching point, and the accuracy does not increase.

Further, as shown in FIG. 11, the current required for the operating angle of the variable valve mechanism to start to move differs for each engine speed Ne. Therefore, in the second embodiment, the offset current ofset_i is calculated from the map shown in FIG. 12 based on the engine rotational speed Ne, and is added to the current command value i _CSC as shown in the following equation (11). Thereby, the control accuracy is further improved.
i _CSC = i _CSFB + i _CSFF + ofset_i (11)
Further, the relationship between the control shaft operating angle θ _CS and the drive current i _CS of the variable valve mechanism is the same when the control shaft operating angle θ _CS increases due to friction or the like, as shown in FIG. It has been experimentally confirmed that there is a hysteresis in which the drive current i _{CS of the} control shaft operating angle θ _CS increases, and that the hysteresis width increases as the engine speed Ne increases.

Therefore, in the third embodiment, the spring constant K (θ _CS ) is calculated for each of the control shaft operating angle θ _CS , the engine rotational speed Ne, and the operating angle change direction using the formula (2), as shown in FIG. An operating angle-spring constant map is created, and the spring constant K (θ _CS ) is calculated by referring to the map. In the present embodiment, the current command value i _CSC is calculated from Equation (10). In this way, highly accurate control can be performed regardless of the operating angle change direction.
15 and 16 show the response of the control shaft operating angle θ _CS when the ultimate operating angle θ _CST is changed stepwise using the control system of the present invention.
In FIG. 15, the normative response and the operating angle coincide with each other, which is a good response. FIG. 16 converges to the target value without overshooting the disturbance.
In addition to the variable valve mechanism, a link mechanism connected to a spring is driven against the biasing force of the spring, and the actuator control apparatus has a nonlinear characteristic between the displacement and the driving force. On the other hand, the same modeling as the variable valve mechanism can be performed and controlled. That is, an actuator operating characteristic model having nonlinear characteristics can be estimated using a spring constant calculated as driving force / displacement for each displacement, and the actuator can be controlled using the model. The same effect as when applied is 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 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 in 2nd Embodiment. The figure which shows the relationship between the operating angle of the said variable valve mechanism, an engine rotational speed, an operating angle change direction, and a drive current. The characteristic map of the spring constant used in 3rd Embodiment. The system block diagram of the variable valve mechanism in embodiment of this invention. The figure which shows the step response when the control apparatus which concerns on this invention is used. Similarly, the figure which shows the step response when the disturbance when the control device which relates to this invention is used occurs.

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 dynamic characteristic compensation unit 102 responsiveness correction unit 103 variable valve mechanism

Claims (4)

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;
When the actuator rotates the control shaft to change the operating angle, the rocking center position of the rocker arm changes, the lift characteristic changes, and the engine valve is opened and closed by driving the rocking cam. In the control device for the variable valve mechanism of the internal combustion engine, the reaction force from the valve spring has a non-linear characteristic with respect to the lift characteristic.
The operation characteristic model of the variable valve mechanism, and any time estimated in the device operates with a spring constant calculated as the working angle of the driving force / the control shaft of the actuator for each operating angle of control shaft,
Using the model, the target value of the operating angle of the control shaft is used as an input value, and the actuator is controlled using the model so that the operating angle of the control shaft follows the target value. Control device for valve mechanism. The control apparatus for a variable valve mechanism according to claim 1, wherein an operation characteristic model of the variable valve mechanism is estimated as the following equation.

Where, J: Moment of inertia of variable valve mechanism D: Viscosity resistance of the device as above K _T : Torque constant K (θcs) of actuator: Spring constant θcs: Typical operating angle of engine valve controlled by variable valve mechanism 3. The control device for a variable valve mechanism according to claim 1, wherein the spring constant differs depending on whether the valve lift amount, which is a lift characteristic, is increased or decreased. The control device for a variable valve mechanism according to any one of claims 1 to 3, wherein the actuator is an electric motor, and the driving force is controlled by an energization current of the electric motor.

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