JP2008202681A - Control system of belt-type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve performance of transmission-control in a belt-type continuously variable transmission. <P>SOLUTION: A transmission controller 21 controls in the following procedure. That stroke of a primary-side movable sheave which corresponds to a target transmission ratio tip is set as a target stroke tx<SB>p</SB>of the primary-side movable sheave (B3). By a non-linear controller B4 which is designed on the basis of an equation of motion of a continuously variable transmission 4 and a transmission control valve 8, a target position tx<SB>m</SB>of a step motor 12 which sets a real stroke x<SB>p</SB>of the primary-side movable sheave as a target stroke tx<SB>p</SB>is computed to drive the step motor 12 so that the real position x<SB>m</SB>of the step motor 12 becomes the target position tx<SB>m</SB>(B6). <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to shift control of a belt type continuously variable transmission.

  The belt type continuously variable transmission changes the primary pulley groove width by changing the primary pressure supplied to the primary pulley with the gear change control valve, and changes the diameter ratio between the primary pulley and the secondary pulley to reduce the gear ratio. It is a transmission that changes in stages.

  The speed change control valve is connected to the center of the speed change link, and the movable sheave of the primary pulley and the step motor are connected to both ends of the speed change link, thereby constituting a speed change ratio feedback control mechanism.

Conventionally, in the transmission control device described above, the entire transmission is approximated by a primary system (for example, a first-order lag transfer function + dead time, etc.), and the control system is designed so as to obtain a desired response. (Patent Document 1)
Japanese Patent No. 3358435

  It is the response of the shift control valve that has the greatest effect on the shift transient characteristics. However, in the conventional control device, since the entire transmission is approximated by a primary system, the response characteristics of the transmission control valve are not sufficiently reflected in the control, and there is still room for improvement in improving the transmission performance.

  Further, in the conventional control device, a target gear ratio is set from the driving state of the vehicle, a target position of the step motor for realizing this is obtained, and the step motor is driven so as to be the target position. However, the gear ratio and the step motor position have a non-linear relationship, the shift control valve has a non-linear characteristic, and the response characteristic of the transmission depends on the direction of the shift (upshift side shift, downshift side shift). Therefore, there is a problem that it is difficult to set a time constant of a transfer function that approximates the entire transmission.

  Even if feedback control is performed to improve steady-state characteristics, there is a problem that setting the feedback gain is difficult because the response characteristics of the transmission differ depending on the shift direction.

  The present invention has been made in view of such technical problems, and an object thereof is to further improve the performance of shift control of a belt-type continuously variable transmission.

  The present invention relates to a continuously variable transmission mechanism including a primary pulley composed of a primary movable sheave and a fixed sheave, a secondary pulley composed of a secondary movable sheave and a fixed sheave, and a belt wound around both pulleys; A speed change link connected to the primary side movable sheave, and a speed change control valve for adjusting a primary pressure supplied to a hydraulic cylinder of the primary side movable sheave by using a line pressure as a source pressure with a spool connected to the center of the speed change link. And a gear ratio feedback control mechanism comprising a gear shift actuator coupled to the other end of the gear shift link, the primary side movable gear corresponding to a target gear ratio. A target straw that sets the sheave stroke as the target stroke of the primary movable sheave A non-linear control means that is designed based on the equation of motion of the continuously variable transmission mechanism and the speed change control valve, and calculates a target position of the speed change actuator that sets the actual stroke of the primary side movable sheave to the target stroke; Actuator driving means for driving the speed change actuator so that the actual position of the speed change actuator becomes the target position.

  According to the present invention, the control system is designed in consideration of the response characteristic of the speed change control valve that has the most influence on the speed change transient characteristics, so that the speed change control is compared with the conventional control in which the whole transmission is linearly approximated. Performance (accuracy, followability) can be improved. Since the control system is designed based on the equation of motion, the control value does not become unstable due to an unrealizable indication value.

  Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 1 shows a schematic configuration of a belt type continuously variable transmission. The belt type continuously variable transmission includes a primary pulley 1 connected to an engine (not shown) via a torque converter or a starting clutch, and a secondary pulley 2 connected to driving wheels (not shown) via a reduction gear, a differential gear, and a drive shaft. And a continuously variable transmission mechanism 4 comprising a belt 3 wound around both pulleys.

  The primary pulley 1 and the secondary pulley 2 are respectively composed of fixed sheaves 1a and 2a and movable sheaves 1b and 2b arranged so as to oppose them. The movable sheaves 1b and 2b are respectively provided with hydraulic cylinders on their back surfaces. When the hydraulic pressure supplied to the oil chambers (primary chamber 1c and secondary chamber 2c) of these hydraulic cylinders is increased or decreased, the movable sheaves 1b and 2b are fixed to the fixed sheave 1a. 2a. As a result, the width of the groove formed between the movable sheaves 1b, 2b and the fixed sheaves 1a, 2a is changed, the contact radius between the belt 3, the primary pulley 1, and the secondary pulley 2 is changed, and the transmission gear ratio is changed. Is changed.

  The hydraulic oil pumped from the oil pump 5 is adjusted to the line pressure Pl by the line pressure control valve 6, and this is used as a source pressure to supply hydraulic pressure (hereinafter referred to as primary pressure) Pp to the primary chamber 1c and to the secondary chamber 2c. Supply hydraulic pressure (hereinafter referred to as secondary pressure) Ps is adjusted by a shift control valve 8 and a secondary pressure control valve 9, respectively. The line pressure control valve 6 and the secondary pressure control valve 9 are solenoid valves (duty solenoids, linear solenoids) that can adjust the supply hydraulic pressure in accordance with the flowing current.

  The spool 8 s of the speed change control valve 8 is connected to the center of the speed change link 10. One end of the speed change link 10 is connected to the primary side movable sheave 1b, and the other end is connected to the step motor 12 (speed change actuator). Thus, a transmission ratio feedback mechanism is configured to bring the actual transmission ratio ip of the transmission closer to the target transmission ratio tip given by the position of the step motor 12.

  The speed change operation of the transmission is performed as follows.

  When changing the gear ratio ip to the Hi side (high speed side, small gear ratio side), the step motor 12 is displaced to the left in the figure, and the speed change link 10 is moved with the end connected to the primary side movable sheave 1b as a fulcrum. The spool 8s of the shift control valve 8 moves from the neutral position shown in the drawing to the left side in the drawing. As a result, the line pressure port 8l of the transmission control valve 8 and the primary pressure port 8p communicate with each other, and the line pressure Pl is introduced into the primary chamber 1c via the transmission control valve 8, thereby increasing the hydraulic pressure in the primary chamber 1c. Further, the secondary pressure control valve 9 is controlled to lower the hydraulic pressure in the secondary chamber 2c.

  Thereby, the primary side movable sheave 1b is displaced in a direction approaching the fixed sheave 1a, and the groove width of the primary pulley 1 is narrowed. Then, the belt 3 is pulled toward the primary pulley 1, the secondary movable sheave 2b is displaced away from the fixed sheave 2a, the groove width of the secondary pulley 2 is increased, and the actual transmission ratio ip of the transmission is shifted to the Hi side. Be changed.

  As the actual speed ratio ip approaches the target speed ratio tip, the speed change link 10 moves with the end connected to the step motor 12 as a fulcrum, and the spool 8s of the speed change control valve 8 moves to the right in the figure. When the actual speed ratio ip matches the target speed ratio tip, the spool 8s of the speed change control valve 8 returns to the neutral position in the figure where the primary pressure port 8p does not communicate with either the line pressure port 8l or the drain port 8d, and the speed change operation is completed. To do. At the same time, the secondary pressure control valve 9 is controlled to increase the secondary pressure Ps, and the belt 3 is clamped by the secondary pulley 2.

  Conversely, when changing the gear ratio ip to the Lo side (low speed side, large gear ratio side), the step motor 12 is displaced to the right side in the figure, and the speed change link with the end connected to the primary side movable sheave 1b as a fulcrum. 10 moves, and the spool 8s of the shift control valve 8 moves from the neutral position shown in the drawing to the right side in the drawing. As a result, the primary pressure port 8p of the transmission control valve 8 and the drain port 8d communicate with each other, and the hydraulic pressure in the primary chamber 1c decreases. Further, the secondary pressure control valve 9 is controlled to increase the hydraulic pressure in the secondary chamber 2c.

  Thereby, the secondary side movable sheave 2b is displaced in a direction approaching the fixed sheave 2a, and the groove width of the secondary pulley 2 is narrowed. Then, the belt 3 is pulled to the secondary pulley 2 side, the primary movable sheave 1b is displaced away from the fixed sheave 1a, the groove width of the primary pulley 1 is increased, and the actual transmission ratio ip of the transmission is set to the Lo side. Be changed.

  As the actual speed ratio ip approaches the target speed ratio tip, the speed change link 10 moves with the end connected to the step motor 12 as a fulcrum, and the spool 8s of the speed change control valve 8 moves to the left in the figure. When the actual speed ratio ip matches the target speed ratio tip, the spool 8s of the speed change control valve 8 returns to the neutral position in the figure where the primary pressure port 8p does not communicate with either the line pressure port 8l or the drain port 8d, and the speed change operation is completed. To do.

  The step motor 12, the line pressure control valve 6, and the secondary pressure control valve 9 are controlled by the transmission controller 21. The transmission controller 21 detects a primary rotational speed sensor 25 that detects the rotational speed Np (angular speed ωp) of the primary pulley 1, a secondary rotational speed sensor 26 that detects the rotational speed Ns of the secondary pulley 2, and a secondary pressure Ps. A signal from the secondary pressure sensor 27 is input. The transmission controller 21 receives signals from the engine controller 22 regarding the throttle opening TVO and the transmission input torque Tin.

  The transmission controller 21 sets a target speed ratio tip based on these input signals, and controls the step motor 12 and the secondary pressure control valve 9 so that this is realized. Further, the line pressure control valve 6 is controlled so that the line pressure Pl is higher than both the primary pressure Pp and the secondary pressure Ps that are required at the time of shifting.

  FIG. 2 is a control block diagram of a portion related to the shift control of the transmission controller 21. The shift control performed by the transmission controller 21 will be further described with reference to this.

  The ultimate speed ratio setting unit B1 sets the ultimate speed ratio dip with reference to a predetermined speed map based on the vehicle speed VSP, the primary rotational speed Np, and the throttle opening TVO. The ultimate speed ratio dip is a speed ratio that should finally be reached, which is determined according to these three input parameters. The vehicle speed VSP is calculated from the secondary rotational speed Ns using the following formula (1):

Can be calculated. if is the final reduction ratio, and R is the radius of the drive wheel.

  In the target transmission ratio and target transmission speed calculation unit B2, the intermediate transmission ratio (hereinafter referred to as the target transmission ratio ip) set in order to bring the actual transmission ratio ip closer to the ultimate transmission ratio dip with a first-order delay of the time constant T. (Speed ratio) tip, and the target speed of change tip ′, which is the speed of change, is expressed by the following equations (2) and (3):

Calculated by s is a differential operator. The actual gear ratio ip can be calculated by dividing the primary rotational speed Np by the secondary rotational speed Ns.

In the gear ratio-stroke conversion unit B3 (target stroke setting means), a target speed ratio tip, target stroke tx _p of the target shift speed tip 'refers to a predetermined conversion table the primary movable sheave 1b, the target stroke speed tx _p Convert to '. At this time, the target stroke speed tx _p ′ that is too fast cannot be realized due to insufficient flow rate balance, and there is a concern that the control of the nonlinear controller B4 and the step motor 12 may be adversely affected, and the belt 3 is movable on the primary side. Since there is a possibility that the sheave 1b is suddenly pushed and slips in the radial direction, an upper limit value is set for the target stroke speed tx _p ′ so as not to exceed this value.

The nonlinear controller B4 (nonlinear control means), based on the actual stroke x _p and target stroke tx _p, primary rotation speed Np of the primary movable sheave 1b, the line pressure Pl (or drain pressure Pd of the shift control valve 8), the primary setting a target position tx _m of the step motor 12 for realizing the target stroke tx _p side movable sheave 1b. Nonlinear controller B4, based on the equation of motion of the continuously variable transmission mechanism 4 and the shift control valve 8 (model) is designed actual stroke x _p of the primary movable sheave 1b to match the target stroke tx _p. A design method of the non-linear controller B4 will be described later.

In stroke step conversion section B5, into a number of steps of the step motor 12 to the target stepper motor position tx _m with reference to the predetermined conversion table, in step motor driving unit B6 (shift actuator driving means), step based on this The motor 12 is driven.

  The target secondary pressure setting unit B7 sets the target secondary pressure tPs based on the target speed ratio tip, the target speed change speed tip ', and the input torque Tin. The target secondary pressure tPs is set so that the belt 3 does not slip because the force with which the secondary pulley 2 clamps the belt 3 is insufficient. At this time, if the belt 3 is clamped too much, the belt 3 may be damaged, and if the actual secondary pressure Ps cannot be increased to the target secondary pressure tPs, the speed change operation becomes unstable. An upper limit value is set for tPs so as not to exceed this value.

  The secondary pressure feedback control unit B8 corrects the target secondary pressure tPs based on the deviation in order to bring the deviation between the target secondary pressure tPs and the actual secondary pressure Ps closer to zero.

  In the hydraulic pressure / current conversion unit B9, the corrected target secondary pressure tPs is converted into a solenoid target current value tis with reference to a predetermined conversion table, and in the solenoid drive unit B10, the current flowing through the solenoid of the secondary pressure control valve 9 Is controlled so that the applied voltage (duty ratio) to the solenoid is controlled to be the target current value tis.

Further, the gear ratio-stroke conversion section B11, determining an actual stroke x _p of the primary-side movable sheave 1b from the actual speed ratio ip by referring to a predetermined conversion table.

The linear control unit B12 (linear control means) is a feedback control for compensating for errors included in the equation of motion (model) of the continuously variable transmission mechanism 4 and the shift control valve 8 used to design the nonlinear controller B4. It is a vessel. Linear controller B12 is a so-called PI controller, based on the deviation e of the actual stroke x _p and the target stroke tx _p of the primary movable sheave 1b, the actual stroke x _p feedback correction amount for approximating the target stroke tx _p to x Calculate _fb . Computed feedback correction amount x _fb is added to the target stepper motor position tx _m, whereby the target step motor position tx _m is corrected.

  Then, the design method of nonlinear controller B4 is demonstrated.

  In designing the non-linear controller B4, first, equations of motion of the continuously variable transmission mechanism 4 and the transmission control valve 8 are established, that is, modeling is performed.

  The behavior of the continuously variable transmission mechanism 4 is expressed by the following equation (4):

Can be expressed as x _p ′ is the stroke speed of the primary movable sheave 1 b, k is the speed change per rotation, ω _p is the angular speed of the primary pulley 1 (= Np × 2π / 60), Pp ′ is the balance thrust equivalent pressure, and Pp is the primary The pressure, Ap, is the piston area of the hydraulic cylinder of the primary pulley 1. The balance thrust equivalent pressure Pp ′ is a primary pressure necessary to maintain the gear ratio at that time, and changes according to the secondary pressure Ps.

In addition, the amount of oil in and out of the primary chamber 1c _QIO is expressed by the following equation (5):

Can be calculated. C _d is a resistance coefficient, w is the circumference of the spool 8s of the transmission control valve 8 (= π × spool diameter), and ρ is the density of the hydraulic oil. ΔP is different between when hydraulic fluid flows into primary chamber 1c (during upshift) and when hydraulic fluid flows out from primary chamber 1c (during downshift), and during upshift, the following equation (6):

It becomes. When downshifting, the following formula (7):

It becomes. Pd is the hydraulic pressure (drain pressure) of the drain port 8d of the transmission control valve 8, and the following equation (8):

Can be calculated. A _ori is the opening area of the orifice 13 provided downstream of the drain port 8d, Q _O is the amount of oil flowing through the orifice 13, and can be calculated by Q _O = Ap × x _p ′. The hydraulic pressure generated downstream of the orifice 13 is α MPa.

Next, change in the flow rate of the primary chamber 1c will be considered. In the following description, the case of upshifting will be described in order to simplify the description. Since the output oil amount _QIO and the volume change of the primary chamber 1c are equal, the following equation (9):

Holds. Then, substituting equation (5) into equation (9), the following equation (10):

Is obtained.

Furthermore, when Expression (4) and Expression (10) are integrated and arranged for x _p ′, the following Expression (11):

Is obtained. Expression (11) is an equation of motion of the continuously variable transmission mechanism 4 and the transmission control valve 8, and the stroke speed x _p ′ of the primary movable sheave 1b, that is, the transmission speed of the transmission is the primary rotational speed Np (of the primary pulley 1). The angular velocity ω _p ), the stroke x _v of the spool 8 s of the speed change control valve 8, the line pressure Pl and the balance thrust equivalent pressure Pp ′.

Nonlinear controller B4, based on the equation of motion (11), is designed to approximate the actual stroke x _p of the primary-side movable sheave 1b with the target stroke tx _p.

  For this reason, first, the value σ is expressed by the following equation (12):

Defined in e is the actual stroke x _p and the target stroke tx _p deviation of the primary-side movable sheave _{1b (e = x p -tx p} ), λ is a positive value, c is the origin correction constant, the e of after infinite time equal.

If σ becomes zero, e ′ = λ · tx _p −λ · x _p . Considering the steady state value after elapse of infinite time, since e ′ = 0, it is understood that λ · (tx _p −x _p ) = 0 and the actual stroke x _p converges to the target stroke tx _p. .

  Further, consider a sufficient condition for σ = 0. A candidate of the Lyapunov function V is expressed by the following formula (13):

Choose as follows. Obviously V ≧ 0, so if V ′ <0, V will monotonously decrease and V = 0 at some time. V = 0 represents σ = 0. In order for the Lyapunov function V to become zero after the infinite time has elapsed, the following equation (14):

Therefore, σ ′ = 0 is a condition. From this, as a sufficient condition for σ = 0, the following equation (15):

This condition can be derived. Substituting equation (11) into equation (15) and setting x _v as tx _v , the following equation (16):

Can be obtained. tx _v is the target stroke of the spool 8s realizing the target stroke tx _p of the primary-side movable sheave 1b. Further, when this equation (16) is arranged by the target stroke tx _v of the spool 8s of the transmission control valve 8, the following equation (17):

Is obtained. sgn () is a function that returns a sign in parentheses.

Further, the geometric relationship of the speed change link 10 is between the target stroke tx _p of the primary side movable sheave 1b, the target stroke tx _v of the spool 8s of the speed change control valve 8, and the target position tx _m of the step motor 12. From the following formula (18):

Therefore, if the target stroke tx _v of the spool 8 s of the speed change control valve 8 is deleted from the equations (17) and (18) and the target position tx _m of the step motor 12 is further arranged, an upshift is performed. sometimes a nonlinear controller B4 for calculating a target position tx _m of the step motor 12 to bring the actual stroke x _p of the primary-side movable sheave 1b with the target stroke tx _p, the following equation (19):

As you can design. x _p0 is the stroke of the primary side movable sheave 1b at the lowest Lo gear ratio.

Nonlinear controller B4 for calculating a target position tx _m of the step motor 12 at the time of downshifting can also be similarly designed by the above procedure, the following equation (20):

It becomes.

  Then, the effect by designing a control system as mentioned above is demonstrated.

The transmission controller 21 sets the stroke of the primary movable sheave 1b corresponding to the gear ratio tip to target as the target stroke tx _p of the primary-side movable sheave 1b (B3), the continuously variable transmission mechanism 4 and the transmission control valve 8 The target position tx _m of the step motor 12 (transmission actuator) that sets the actual stroke x _p of the primary side movable sheave 1b to the target stroke tx _p is determined by the nonlinear controller B4 designed based on the equation of motion (Equation (11)). Then, the step motor 12 is driven so that the actual position x _m of the step motor 12 becomes the target position tx _m (B6, inventions according to claims 1 and 3 to 6). According to the present invention, the control system is designed in consideration of the response characteristic of the shift control valve 8 that has the greatest influence on the shift transient characteristic. Control accuracy and followability can be improved. Since the control system is designed based on the equation of motion, an unrealizable instruction value is not output and the control is not unstable.

The transmission controller 21 is further by a linear controller B12, corrects the target position tx _m of the step motor 12 based on the deviation of the actual stroke x _p and the target stroke tx _p of the primary-side movable sheave 1b (claim 2 Invention). According to the present invention, the actual stroke x _p and the target stroke resulting from the error (modeling error) included in the equations of motion of the continuously variable transmission mechanism 4 and the shift control valve 8 used when designing the nonlinear controller B4. Since the deviation of tx _p can be corrected and the steady error between the actual stroke x _p and the target stroke tx _p can be eliminated, the performance of the shift control can be further improved.

  Further, since the linear controller B12 is designed based on the stroke of the movable sheave 1b that changes linearly with respect to the position of the step motor 12, the linear controller B12 is designed in comparison with the case where it is designed based on the non-linearly changing gear ratio. Easy. Conventionally, to design a feedback control system that eliminates modeling errors, the response characteristics of the transmission differ depending on the shift direction (upshift side shift, downshift side shift). However, according to the present invention, such a change is not necessary.

Further, as a result of designing the non-linear controller B4 based on the equation of motion (the equation (11)) of the continuously variable transmission mechanism 4 and the transmission control valve 8, the stroke speed x _p ′ of the primary side movable sheave 1b, that is, the transmission speed change. It can be seen that the speed depends on the rotation speed Np (angular speed ω _p ) of the primary pulley (Equation (11)). In the nonlinear controller B4, the step motor 12 is based on the rotation speed Np (angular speed ω _p ) of the primary pulley. The target position tx _m is calculated (invention according to claims 3 to 6). According to this invention, the performance of the shift control can be further improved.

  The embodiment of the present invention has been described above, but the above embodiment is merely an example of application of the present invention, and the technical scope of the present invention is not intended to be limited to the specific configuration of the above embodiment.

It is a schematic block diagram of a belt type continuously variable transmission. It is a control block diagram of the part regarding the shift control of a transmission controller.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Primary pulley 2 Secondary pulley 1a, 2a Fixed sheave 1b, 2b Movable sheave 3 Belt 4 Continuously variable transmission mechanism 6 Line pressure control valve 8 Shift control valve 8s Spool 9 Secondary pressure control valve 10 Shift link 12 Shift control valve (shift actuator)
21 Transmission controller 25 Primary rotational speed sensor 26 Secondary rotational speed sensor 27 Secondary pressure sensor

Claims (6)

A continuously variable transmission mechanism comprising a primary pulley composed of a primary movable sheave and a fixed sheave, a secondary pulley composed of a secondary movable sheave and a fixed sheave, and a belt wound around both pulleys;
Shift control, one end of which is connected to the primary movable sheave, and a spool that is connected to the center of the transmission link and adjusts the primary pressure supplied to the hydraulic cylinder of the primary movable sheave using the line pressure as a source pressure. A transmission ratio feedback control mechanism comprising a valve and a transmission actuator coupled to the other end of the transmission link;
In a shift control device for a belt-type continuously variable transmission comprising:
Target stroke setting means for setting a stroke of the primary side movable sheave corresponding to a target gear ratio as a target stroke of the primary side movable sheave;
Nonlinear control means that is designed based on the equation of motion of the continuously variable transmission mechanism and the transmission control valve, and that calculates the target position of the transmission actuator that makes the actual stroke of the primary movable sheave the target stroke;
Transmission actuator driving means for driving the transmission actuator so that the actual position of the transmission actuator becomes the target position;
A shift control apparatus for a belt-type continuously variable transmission, comprising:   The belt-type continuously variable transmission according to claim 1, further comprising linear control means for correcting the target position of the speed change actuator based on a deviation between the actual stroke and the target stroke of the primary movable sheave. Shift control device.   The non-linear control means calculates the target position of the speed change actuator based on the actual stroke and target stroke of the primary side movable sheave, the rotation speed of the primary pulley, and the line pressure during upshifting. The shift control device for a belt type continuously variable transmission according to claim 1 or 2. The non-linear control means calculates the target position of the speed change actuator at the time of upshift by the following formula:

Where x _p0 : stroke of primary movable sheave at maximum Lo gear ratio, x _p : actual stroke of primary movable sheave, tx _p : target stroke of primary movable sheave, tx _p ': target of primary movable sheave Stroke speed, Ap: Piston area of hydraulic cylinder of primary pulley, ρ: Hydraulic oil density, k: Shift speed per rotation, w: Circumference of spool, ω _p : Angular speed of primary pulley, C _d : Resistance coefficient , Pl: line pressure, λ: positive coefficient, Pp ′: balance pressure equivalent value,
The shift control device for a belt-type continuously variable transmission according to claim 3, wherein   The non-linear control means calculates the target position of the shift actuator based on the actual stroke and target stroke of the primary movable sheave, the rotation speed of the primary pulley, and the drain pressure of the shift control valve during downshift. The shift control device for a belt-type continuously variable transmission according to any one of claims 1 to 4, wherein: The non-linear control means represents the target position of the speed change actuator at the time of downshift:

Where x _p0 : stroke of primary movable sheave at maximum Lo gear ratio, x _p : actual stroke of primary movable sheave, tx _p : target stroke of primary movable sheave, tx _p ': target of primary movable sheave Stroke speed, Ap: Piston area of hydraulic cylinder of primary pulley, ρ: Hydraulic oil density, k: Shift speed per rotation, w: Circumference of spool, ω _p : Angular speed of primary pulley, C _d : Resistance coefficient , Pd: drain pressure, λ: positive coefficient, Pp ′: balance pressure equivalent value,
6. The shift control apparatus for a belt-type continuously variable transmission according to claim 5, wherein

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