JP2001248717A - Line pressure control device for continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a line pressure control device for a continuously variable transmission, assuring the optimum line pressure even in transient speed change. SOLUTION: The line pressure control device for the continuously variable transmission in which a target line pressure PLA of an working fluid to be supplied to the continuously variable transmission is set corresponding to the operated condition of a vehicle to control the line pressure of the continuously variable transmission mounted on the vehicle so that an actual line pressure PL of the working fluid is controlled to achieve the target line pressure PLA, comprises consumed flow correlation value deriving means 58A for deriving a consumed flow correlation value correlated to the consumed flow rate of the working fluid to be consumed in transient speed change and line pressure correction amount determining means 58B for determining a line pressure correction amount ΔDLSFT in accordance with the consumed flow correlation value derived from the consumed flow correlation value deriving means 58A, whereby the target line pressure PL is corrected corresponding to the line pressure correction amount ΔDLSFT determined by the line pressure correction amount determining means 58B.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a line pressure control device for a continuously variable transmission suitable for use in a belt-type continuously variable transmission mounted on a vehicle.

[0002]

2. Description of the Related Art In recent years, continuously variable transmissions have been focused on the point that they can avoid a shift shock by continuously controlling the gear ratio and that they have excellent fuel consumption efficiency. Have been. In such a continuously variable transmission, the gear ratio is generally controlled by hydraulic control.

For example, in the case of a belt-type continuously variable transmission, power generated by an engine is transmitted from a primary pulley to a secondary pulley via a belt. On this occasion,
Normally, hydraulic pressure (line pressure) set for the hydraulic piston of the secondary pulley according to basic characteristics such as transmission torque
To provide a clamping force to the belt and adjust the hydraulic pressure (primary pressure) applied to the hydraulic piston of the primary pulley to change the gear ratio (the speed ratio (the effective radius ratio between the primary pulley and the secondary pulley). Control].

[0004] In such a belt-type continuously variable transmission, in particular, if the line pressure is insufficient, the belt slips and hinders the transmission of power. Conversely, if the line pressure is excessive, the load on the hydraulic power source side is increased. Therefore, it is necessary to control the line pressure so that the line pressure does not excessively increase without causing the slip of the belt. The appropriate line pressure that does not cause the belt slip and does not become excessive corresponds to the magnitude of the torque transmitted by the belt, so that the torque input to the continuously variable transmission (transmission input torque) ) And the transmission ratio (transmission ratio) of the continuously variable transmission, a PID correction (proportional correction (P correction), integral correction) so that the actual line pressure becomes the target line pressure. (I
Correction), differential correction (D correction)].

[0005]

As a hydraulic circuit of such a belt-type continuously variable transmission, an oil chamber for driving a hydraulic piston (primary piston) of a primary pulley and a hydraulic piston (secondary piston) of a secondary pulley are used. Some are configured to communicate with a driving oil chamber via a flow control valve.

However, in the case of such a hydraulic circuit, at the time of gear shifting, in addition to hydraulic oil for securing a line pressure for applying a predetermined clamping force to the belt, an operation for moving both pistons for gear shifting control is performed. Since oil is also required, hydraulic oil discharged from the oil pump may be insufficient, and the line pressure may drop transiently, so that a predetermined clamping force may not be applied to the belt.

At this time, although the correction is performed by the feedback control as described above, when the shift speed is high, the moving speeds of the primary piston and the secondary piston are also increased, and the amount of hydraulic oil required for moving both pistons is reduced. Increases transiently. Further, since there is a delay in the rise response of the actual line pressure due to the shortage of the hydraulic oil, the difference between the actual line pressure and the target line pressure becomes excessive, and the correction by the feedback control cannot catch up. That is, the consumption flow rate of the hydraulic oil transiently increases, and the response for raising the line pressure again becomes slow. In the worst case, there is a possibility that a belt slip may occur.

SUMMARY OF THE INVENTION The present invention has been made in view of the above problems, and provides a line pressure control device for a continuously variable transmission which can secure an optimum line pressure even during a shift transition. Aim.

[0009]

Therefore, in the line pressure control device for a continuously variable transmission according to the present invention, a consumption flow rate correlation value correlated with a consumption flow rate of a working fluid consumed during a shift transition is provided. The consumed flow correlation value deriving means is derived, and the line pressure correction amount is determined by the line pressure correction amount determining means based on the consumed flow correlation value. Then, the target line pressure of the working fluid set according to the operation state is corrected by the line pressure correction amount, and the actual line pressure of the working fluid is controlled so as to be the corrected target line pressure.

In the line pressure control device for a continuously variable transmission according to the present invention, the consumption flow rate correlation value deriving means derives the consumption flow rate correlation value based on the speed ratio change amount. Preferably, the consumption flow correlation value deriving means derives the consumption flow correlation value based on the actual change amount of the gear ratio. in this case,
The actual consumption flow rate of the working fluid can be derived, and based on the actual consumption flow rate, a target line pressure suitable for the operating state of the vehicle can be set.

Further, the consumption flow rate correlation value deriving means may derive the consumption flow rate correlation value based on the change amount of the target gear ratio. In this case, the responsiveness of the line pressure control can be improved as compared with deriving the consumption flow rate correlation value based on the actual change amount of the gear ratio.

[0012]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below with reference to the drawings. FIG. 1 to FIG. And will be described with reference to these figures. First, a power transmission mechanism of a vehicle equipped with the continuously variable transmission according to the present embodiment will be described with reference to FIGS. 2A and 2B.
As shown in FIG. 1, in the present power transmission mechanism, rotation output from an engine (internal combustion engine) 1 is transmitted through a torque converter (torque converter) 2 to a belt-type continuously variable transmission (CVT).
The transmission is transmitted to a front differential 31 from a counter shaft (not shown).

A forward / reverse switching mechanism 4 is provided between the output shaft 7 of the torque converter 2 and the input shaft 24 of the continuously variable transmission 20, and is input from the engine 1 via the torque converter 2. The rotation is input to the continuously variable transmission mechanism 20 via the forward / reverse switching mechanism 4. The continuously variable transmission 20 is a hydraulic continuously variable transmission that performs shift control and the like by hydraulic control described later.

The stepless speed change mechanism 20 will be described in more detail. The stepless speed change mechanism 20 includes a primary pulley 21, a secondary pulley 22, and a belt 23. The input rotation is input to the secondary pulley 22 via the belt 23 from the primary pulley 21 coaxial with the primary shaft 24.

The primary pulley 21 and the secondary pulley 22 are respectively two sheaves 21a, 21 which rotate integrally.
1b, 22a and 22b. One of the sheaves 21a, 22a is a fixed sheave fixed in the axial direction, and the other sheave 21b, 22b is a movable sheave movable in the axial direction by hydraulic pistons 21c, 22c.

The hydraulic pistons 21c and 22c are supplied with hydraulic oil (working fluid) in an oil tank 61 by an oil pump 62.
Is supplied to the movable sheaves 21b and 22b, and accordingly, the fixed sheaves 21a and 22a of the movable sheaves 21b and 22b are supplied.
The pressing force on the side is adjusted. The hydraulic piston 22c of the secondary pulley 22, a regulator valve (pressure regulating valve) 63 to a more regulated line pressure P _L is applied, the hydraulic piston 21c of the primary pulley 21
The shift control valve on whose pressure regulated by the regulator valve 63 (flow control valve) 64 flow-rate-controlled hydraulic oil is supplied by, acting as the working oil pressure adjustment speed ratio (primary pressure) P _P It is supposed to.

The regulator valve 63 is controlled by duty control of a line pressure control solenoid 63A by an electric signal.
Is controlled by duty control of the shift control solenoid 64A by an electric signal.
The hydraulic oil discharged from the discharge port 71 of the regulator valve 63 is sent to a lubrication valve (not shown) used for supplying the lubricating oil, and is adjusted to a predetermined pressure P _LUB . , And then returned to the oil tank 61. The oil discharged from the discharge port 72 of the flow control valve 64 is returned to the oil tank 61.

The line pressure P _L should be set as low as possible within a range where power transmission can be ensured by avoiding slippage of the belt 23 to reduce the energy loss by the oil pump 62 and the transmission itself. It is important for enhancing durability. Slip of the belt 23, the torque so generated according to the magnitude of the torque transmitted by the belt 23, which is input to be such a slip does not occur, the continuously variable transmission as described below the target line pressure P _LA ( The transmission pressure is set in accordance with the transmission input torque) and the speed ratio of the continuously variable transmission (transmission speed ratio), and the line pressure is controlled based on this.

The gear ratio (primary pressure P _P ) is
It is controlled by feedback control based on the actual rotation speed of the primary pulley 21. Here, the primary pulley 21 is determined based on the rotation speed of the secondary pulley 22 (secondary rotation speed) corresponding to the vehicle speed and the load of the engine mounted on the vehicle (for example, accelerator opening).
, The deviation ΔN _P between the actual rotation speed N _P of the primary pulley 21 and the target rotation speed N _PT (= N _PT −
N _P) is calculated, based on the control amount subjected to PID correction on the deviation .DELTA.N _P (shift duty), the actual rotational speed N _P is the target rotational speed N becomes _PT way flow control valve of the primary pulley 21 64 Feedback control.

The hydraulic pulley of the secondary pulley 22
Line pressure P applied to the ton 22c _LAnd primary
Primary given to hydraulic piston 21c of tool 21
Pressure P_PThe controller (electronic control unit)
Are controlled by the command signal of the ECU 50.
It has become so. That is, as shown in FIG.
In addition, the ECU 50 includes an engine rotational speed sensor (clamp).
Angle sensor or cam angle sensor) 41, throttle opening
Sensor 46 and the rotational speed of primary pulley 21 are detected.
First rotation speed sensor 43, rotation of secondary pulley 22
Second rotational speed sensor 44 for detecting speed, line pressure P_L
Line pressure sensor 45 that detects the
Primary pressure) P_PPressure sensor 47 for detecting
Of the ECU 5 is input.
0, each of the pulleys 21, 22 based on these detection signals
Regulation provided in the hydraulic supply system (hydraulic circuit)
To control the data valve 63 and the flow control valve 64.
ing.

The ECU 50 has a function (shift control means or primary pressure control means) 52 for controlling the flow rate control valve 64 (speed ratio control) and a function (line pressure control) for controlling the regulator valve 63 (line pressure control). Pressure control means) 53, and the line pressure control device of the present embodiment, as shown in FIG.
And the above-mentioned regulator valve 63, a solenoid 63A for controlling line pressure, and a line pressure sensor 45.

The line pressure control means 53 includes a target line pressure setting means 54, an upper and lower limiter 55A, a primary filter 55
B, subtractor 55C, basic duty setting means 56A, rotation / oil temperature correction means 56B, adders 56C, 56D, 56
E, PID correction means 57, shift transient correction amount calculation means 5
8, and an upper / lower limiter 59 is provided. As a result, the target line pressure P _LA set by the target line pressure setting means 54 is first defined by the upper and lower limiters 55A, and then filtered by the primary filter 55B so as to suppress sudden fluctuations. Thereafter, the basic value (basic duty) D _LB of the line pressure control amount is set by the basic duty setting means 56A based on the filtered target line pressure P _LA and the engine speed (rotation speed) Ne. To the basic duty D _LB , an adder 56C adds a correction amount determined based on the engine rotation speed Ne and the oil temperature detected by an oil temperature sensor (not shown) based on the rotation / oil temperature correction means 56B, Further, a learning value determined according to the engine rotation speed Ne and the oil temperature is added by the adder 56D in order to correct variations in individual performances of the oil pump 62, and a line pressure control amount (control duty) D _L is calculated. It has become so.

On the other hand, the subtractor 55C calculates a deviation ΔP _L between the filtered target line pressure P _LA and the actual line pressure P _L detected by the line pressure sensor 45, and calculates PID.
The correction means 57 performs PID correction on the deviation ΔP _L to calculate a correction amount (a correction amount of the control duty) ΔDL of the line pressure control amount. In the present apparatus, the correction amount ΔDL and the correction amount ΔDL _SFT calculated by the shift transient correction amount calculation means 58 are added to the line pressure control amount DL by the adder 56E, and then the upper and lower limiters 59 set the upper and lower limits. After the regulation, it is output to the line pressure control solenoid 63A, whereby the regulator valve 63 is controlled to a predetermined opening.

Hereinafter, the target line pressure setting means 54 and the shift transient correction amount calculating means 58 will be further described. First, the target line pressure setting means 54 will be described. In the target line pressure setting means 54, the input torque to the continuously variable transmission 20 (transmission input torque), the gear ratio of the continuously variable transmission (transmission gear ratio) ratio, The target line pressure _PLA is set in accordance with.

The transmission input torque is
Engine output torque Te and torque converter torque ratio t (e)
And is calculated based on The engine output torque Te is calculated (or estimated) from the engine rotation speed Ne detected by the engine rotation speed sensor 41 and the throttle opening θth detected by the throttle opening sensor 46.
The torque converter torque ratio t (e) is calculated based on a map stored in advance, based on the speed ratio in the torque converter 2 [the output rotation speed of the torque converter (turbine rotation speed, that is, the primary pulley detected by the first rotation speed sensor 43). is calculated from 21 rotational speed) N _P input rotation speed of the torque converter 2 (the ratio between the engine rotation speed) Ne] N _P / Ne of.

The transmission gear ratio
Is the effective radius (belt winding radius) ratio between the primary pulley 21 and the secondary pulley 22, for example, the rotation speed N _P of the primary pulley 21 detected by the first rotation speed sensor 43 and the second rotation speed. It can be calculated from the rotational speed N _S of the secondary pulley 22 detected by the sensor 44 (N _P / N _S ).

Next, the shift change correction amount calculating means 58 will be described. The shift change correction amount calculating means 58 consumes the hydraulic pistons 21c and 22c of the pulleys 21 and 22 during the shift control (during the shift shift). and Q consumption flow rate calculation means for calculating the _R (consumption rate correlation value derivation means) 58A (consumption flow rate of the working fluid) amount of hydraulic oil is consumed flow rate calculating unit 5
It is configured to include a line pressure correction amount determining means 58B for determining a correction amount dL _SFT line pressure control amount based on the consumed flow rate Q _R of the computed hydraulic oil by 8A.

The consumption flow rate calculating means 58A detects the hydraulic piston 21 during the shift control based on the speed ratio change amount of the continuously variable transmission.
c, it is intended for calculating the amount of hydraulic oil Q _R consumed by 22c. In the continuously variable transmission 20, when the upshift is performed, the hydraulic piston 21
By increasing the hydraulic pressure (primary pressure) acting on c, hydraulic oil is supplied to the oil chamber of the hydraulic piston 21c to move the movable sheave 21b closer to the fixed sheave 21a, thereby reducing the distance between the fixed sheave 21a and the movable sheave 21b. As a result, the belt 23 is placed in the V-shaped groove (V-groove) formed between the fixed sheave 21a and the movable sheave 21b.
1a, 21b to the outer peripheral side, and the belt winding radius r _P
Is to be increased.

On the other hand, in the secondary pulley 22,
Sheaves 22a, the belt 23 in V groove formed by the 22b sheave 22a, the belt entraining with a radius r _S retracted toward the axis of 22b decreases. Therefore, the hydraulic oil is forcibly discharged from the oil chamber of the hydraulic piston 22c, and the movable sheave 22b is separated from the fixed sheave 22a, so that the distance between the fixed sheave 22a and the movable sheave 22b is increased.

Here, the hydraulic oil forcibly discharged from the secondary pulley 22 is returned to the hydraulic circuit again and the line pressure P
Used to secure _L. Accordingly, at the time of upshift, the hydraulic oil amount that is consumed to separate the shift control from the hydraulic oil amount to ensure the line pressure P _L (flow consumption) Q _R, the operating oil supplied to the hydraulic piston 21c The quantity (primary oil flow rate) Q _P and the hydraulic oil quantity (secondary oil flow rate) Q _S discharged from the hydraulic piston 22c can be calculated by the following equation (1).

[0031] _{Q R = Q P -Q S ...} (1) On the other hand, when downshifting, the primary pulley 21, through the hydraulic (primary pressure) lowering the P _P (flow control valve 64 which acts on the hydraulic piston 21c The operating oil is discharged from the oil chamber of the hydraulic piston 21c to move the movable sheave 21b to the fixed sheave 21.
by spaced apart from a decrease the belt wrapping radius r _P. On the other hand, in the secondary pulley 22, the hydraulic piston 2
Movable sheave 22b is adapted to close the belt entraining with a radius r _S to the fixed sheave 22a is increased by the line pressure P _L supplied to the oil chamber of 2c. At this time, the hydraulic oil discharged from the primary pulley 21 is supplied to the flow control valve 64
Is not discharged from the discharge port 72, and does not contribute to securing the line pressure. Therefore, the hydraulic oil amount required for the shift control during downshifting Q _R is the hydraulic piston 2
Hydraulic oil amount (secondary oil flow rate) Q supplied to 2c
Only _{S is} available, and can be calculated by the following equation (2).

[0032] _{Q R = Q S ... (2} ) where the hydraulic oil amount Q _P of the primary pulley 21 side, as shown in the following equation (3), perpendicular to the piston movement direction of the hydraulic chambers of the hydraulic piston 21c It is calculated as the product of the cross-sectional area A _P and the movement amount X _P of the movable sheave 21b from the previous control cycle.

Q _P = A _P × X _P (3) Similarly, the hydraulic oil amount Q _S on the secondary pulley 22 side is, as shown in the following equation (4), the piston moving direction of the hydraulic chamber of the hydraulic piston 22c. Cross section A _S perpendicular to and movable sheave 2
And it calculates a product of the amount of movement X _S of 2b. Q _S = A _S × X _S (4) Each of the movement amounts X _P and X _S is a movement amount of the movable sheaves 21b and 22b with respect to the previous control cycle, and the following expression (5),
It can be calculated by (6).

X _P = (r _{P (n)} −r _{P (n−1)} ) × tan θ _P (5) X _S = (r _{S (n)} −r _{S (n−1)} ) × tan θ _S. (6) Note that θ _P and θ _S in the above equations (5) and (6) are as shown in FIG.
As shown in the figure, the inclination angle (cone angle) of the movable sheave 21b,
This is the inclination angle (cone angle) of the movable sheave 21b. Also,
r _{P (n)} is particularly the primary pulley 21 in the current control cycle.
Belt entraining indicates radius _{r P, r P (n-} 1) is a belt entraining radius r _P of the primary pulley 21 in particular prior control cycle
Is shown. Similarly, r _{S (n)} indicates the belt winding radius r _S of the secondary pulley 22 in the current control cycle, and r _{S (n−1)} indicates the belt winding radius r _S of the secondary pulley 22 in the previous control cycle. . The belt winding radii r _P and r _S are determined by the transmission gear ratio as shown in the following equations (7) and (8).
Each is calculated as a function of ratio.

R _P = f _P (ratio) (7) r _S = f _S (ratio) (8) Here, the above equations (3) and (4) are replaced with the above equations (5) and (6). By substituting and rearranging, the above equation (3) can be rearranged as shown in the following equation (9), and the upper equation (4) can be rearranged as shown in the following equation (10).

Q _P = (r _{P (n)} −r _{P (n−1)} ) × C _P (Note that C _P = A _P × tan θ _P ) (9) Q _S = (r _{S (n)} − r _{S (n-1)} ) × C _S (C _S = A _S × tan θ _S ) (10) C _P in the above equations (9) and (10) is the cross-sectional area A _P of the hydraulic piston 21c. Similarly, C _S is a constant obtained from the cross-sectional area A _S of the hydraulic piston 22c and the inclination angle θ _S of the movable sheave 22b, and C _P and C are constants obtained from the inclination angle θ _P of the movable sheave 21b. _S is stored in the ECU 50 in advance. Therefore, the hydraulic oil amount Q _P on the primary pulley 21 side is uniquely calculated from the change amount of the belt winding radius r _P , and the hydraulic oil amount Q _S on the secondary pulley 22 side is calculated from the change amount of the belt winding radius r _S. It is uniquely calculated. The belt entraining radius r _P, r _S of the formula (7) is a function of both the transmission gear ratio ratio as shown in (8), therefore, the hydraulic oil amount Q _P, the Q _S thus shift control the hydraulic oil amount Q _R required transmission gear ratio r
The calculation is performed as a function of the change amount of the atio (change ratio of the gear ratio).

Now, the line pressure correction amount determining means 58B
Based on the consumption quantity calculation means 58A hydraulic oil amount required on the calculated shift control by Q _R of the above, the pressure reduction amount of the line pressure P _L in the shift control (line pressure decrease amount)? PL _SFT calculated, this line order to compensate for pressure decrease amount? PL _SFT fraction, is what determines the correction amount dL _SFT line pressure control duty according to the line pressure decrease amount? PL _SFT.

The method of calculating the line pressure decrease amount ΔPL _SFT will be described. In a steady state where the shift control is not performed, the regulator valve passing flow rate Qv discharged from the discharge port 71 of the regulator valve 63 to the lubrication system is described.
Are the flow coefficient c, the regulator valve opening area A _R , the oil pressure at the inlet of the discharge port 71 (that is, the line pressure regulated by the regulator valve 63) P _L , and the oil pressure at the outlet of the discharge port 71 (adjusted by a lubrication valve not shown). Pressure) P _LUB and the density ρ of the hydraulic oil can be expressed by the following equation (11).

[0039]

(Equation 1)

On the other hand, the regulator valve passage flow rate Qv 'during the shift control can be similarly expressed by the following equation (12). P _L ′ indicates the line pressure during the shift control.

[0041]

(Equation 2)

Here, during the shift control, the movable sheaves 21b, 2b are moved by the hydraulic pistons 21c, 22c as described above.
Since driving the 2b requires hydraulic oil amount Q _R for actuating the hydraulic piston 21c, a 22c, Therefore, passing the regulator valve flow Qv', the working oil amount of the regulator valve passing flow Qv of the steady state Q _R (Qv ′ = Qv−Q _R ), and the line pressure P _L ′ is reduced by ΔPL _SFT from the line pressure P _{L in the} steady state (P _L ′ = P _L −ΔPL). _SFT ). Therefore, the above equation (12) can be expressed as the following equation (13).

[0043]

(Equation 3)

Here, the following relational expression (14) can be derived from the above expressions (11) and (13).

[0045]

(Equation 4)

Then, the two sides of the above equation (14) are squared and rearranged to obtain the following equation (15).

[0047]

(Equation 5)

It should be noted that the regulator valve passing flow rate Qv
Is the discharge of the oil pump 62 as shown in the following equation (16).
Output Q_PUMPMore oil leakage from the transmission
Quantity Q_LIs subtracted. Qv = Q_PUMP−Q_L ... (16) The oil leakage amount Q_LIs the line pressure P_LFunction f_L(P_L)
And [Q_L= F_L(P _L)], Function f_L(P_L) Is E
It is stored in the CU 50. Also, the oil pump 62
Discharge volume Q_PUMPIs given by the following equations (17) and (18)
Pump theoretical discharge volume V_th, The function f of the engine speed Ne
_OPE(Ne), oil pump volumetric efficiency η_{O PV}, Line pressure P
_LAnd so on. Note that the function f_OPE
(Ne) is stored in the ECU 50 in advance.

Η _OPV = 1− [f _OPE (Ne) × P _L ] (17) Q _PUMP = V _th × (Ne / 1000) × η _OPV (18) Also calculated by the above equation (17) Oil pump volumetric efficiency η _OPV is smaller than the predetermined lower limit η _MIN ,
Oil pump volumetric efficiency eta _OPV is adapted to be set as a lower limit value _{η MI N (η OPV = η} MIN).

[0050] Then, the correction amount dL _SFT of the line pressure control duty, as shown in the following equation (19), by multiplying the line pressure decrease amount? PL _SFT function f _D of the line pressure P _L (P _L) Is calculated. The function f _D (P _L ) is stored in the ECU 50 in advance. ΔDL _SFT = f _D (P _L ) × ΔPL _SFT (19) Further, the correction amount ΔDL _SFT is equal to a predetermined upper limit value ΔDL.
_When it is larger than _SFTMAX , the correction amount ΔDL _SFT is set by this upper limit value ΔDL _SFTMAX (ΔDL _SFT = ΔDL
_SFTMAX ) On the other hand, if the correction amount ΔDL _SFT is smaller than the predetermined lower limit ΔDL _SFTMIN , the correction amount ΔDL _SFT
SFT is set to 0 (zero) (ΔDL
_SFT = 0).

Since the line pressure control device of the continuously variable transmission according to one embodiment of the present invention is configured as described above,
For example, as shown in the flowchart of FIG.
_SFT is determined. That is, in step S10, the primary pulley 2 detected by the first rotational speed sensor 43
1 of the rotation speed N _P and the transmission speed ratio Ratio from the rotational speed N _S of the secondary pulley 22 detected by the second rotational speed sensor 44 is calculated, in step S20, the primary pulley 21 as a function of the transmission gear ratio Ratio Side belt winding radius r _{P (n)}
Is calculated [r _{P ( n)} = f _P (ratio)].

In step S30, the belt winding radius r
_The movement of the movable sheave 21b from the previous control cycle is determined by the difference between _{P (n)} and the belt winding radius r _{P (n-1)} of the primary pulley 21 in the previous control cycle and the inclination angle θ _{P of} the movable sheave 21b. The quantity _XP is calculated [ _XP = (rP _{( n)} -rP _(n-1) ). Times.tan.theta.
_P], further, in step S40, the moving amount X _P and the hydraulic oil amount Q _P of the primary pulley 21 side as the product of the sectional area A _P of the hydraulic chambers of the hydraulic piston 21c is calculated (Q _P =
A _P × X _P ).

Then, similarly, in step S50, the secondary pulley 22 is determined from the transmission gear ratio.
Side belt winding radius r _{S (n)} is calculated [r _{S (n)} = f _s (r
atio)], in step S60, the variation of the belt winding radius r _{s from} the previous control cycle and the inclination angle θ of the movable sheave 22b.
_{From S} , the amount of movement X _{S of} the movable sheave 22b from the previous control cycle
[X _S = (r _{S (n)} −r _{S (n−1)} ) × tan
theta _S], further, in step S70, the amount of movement X _S hydraulic oil amount of the secondary pulley 22 side from the hydraulic chamber of the cross-sectional area A _S of the hydraulic piston 22c Q _S is calculated (Q _S = A _S ×
X _S ).

Then, in step S80, the difference Δratio (=
ratio _{n -} ratio _{n -1} , ratio _n : transmission gear ratio in the current control cycle, ratio _n-1 : transmission gear ratio in the previous control cycle), it is determined whether an upshift or a downshift is in progress. And such a difference Δrati
If o is a positive number, it is determined that a downshift is being performed, and the process proceeds to step S90, where the hydraulic oil amount Q on the secondary pulley 22 side is set.
_S is the amount of hydraulic oil Q _R required shift control (Q _R =
Q _S).

On the other hand, the process proceeds to step S100 as such difference Δratio is in the upshift if negative, the hydraulic oil amount Q _R required for shift control, the secondary from the working oil amount Q _P of the primary pulley 21 side It is calculated as an amount obtained by subtracting the amount of hydraulic oil Q _S of the pulley 22 side (Q _R = Q _P -
Q _S). The oil pump volumetric efficiency eta _OPV is calculated [the above equation (17) refer] In step S110, the oil leak amount Q _L is estimated as a function of the line pressure P _L in step S120, the above equation in step S130 (16) , (18)
Regulator valve passing flow Qv is calculated from these oil pump volumetric efficiency eta _OPV and oil leak amount Q _L by.

Further, in step S140, the above equation (15) is obtained.
Accordingly, the hydraulic oil amount Q _R and regulator valve passing flow Qv line pressure decrease amount from such? PL _SFT required shift control is calculated by the equation (19) in step S150, the line pressure decrease amount? PL _SFT and the line pressure P The correction amount ΔDL _SFT of the line pressure control duty is calculated from _L. Then, the line pressure control duty is corrected by the correction amount ΔDL _SFT , and the regulator valve 63 is controlled with the corrected line pressure control duty.

[0057] Thus, during the shift control, whereas the required hydraulic oil amount Q _R for separate transmission control from the hydraulic oil amount to ensure the line pressure P _L, the present line pressure control device if the line pressure decrease amount ΔP by the hydraulic oil amount Q _R
Since the line pressure control duty is corrected to compensate for L _SFT , there is an advantage that the belt 23 can be clamped during shift control and belt slip can be prevented. This also has the advantage that the belt 23 can be prevented from deteriorating due to belt slip and the belt durability can be improved.

[0058] Further, since the hydraulic oil amount Q _R required shift control is calculated based on the amount of change in the transmission gear ratio ratio, it is not necessary to provide a special sensors, thus the above effects without increasing the cost There is also the advantage that it can be obtained. Further, the change amount of the transmission gear ratio is converted into the rotation speed N _P of the primary pulley 21 detected by the first rotation speed sensor 43 and the rotation speed N _S of the secondary pulley 22 detected by the second rotation speed sensor 44. can be detected as an actual change amount based, it is possible to calculate the amount of hydraulic oil Q _R consumed for actual shift control on the basis of this variation, and thus on the basis of the hydraulic oil amount Q _R, There is an advantage that the line pressure can be controlled in a manner similar to the actual operation state.

The line pressure control device for a continuously variable transmission according to the present invention is not limited to the above embodiment. For example, in the embodiment described above, as a flow consumption correlation value deriving means, although using a consumption quantity calculation means 58A for calculating the amount of hydraulic oil Q _R from the change amount of the transmission gear ratio ratio, instead of the consumption flow rate calculating unit 58A hydraulic piston 21c as consumption rate correlation value deriving means, 22c a flow meter provided, may be configured to detect the hydraulic oil amount Q _R directly by the flow meter. In any event, flow consumption correlation value deriving means, consumption rate (the amount of hydraulic oil) Q _R or as long as it derives the correlation values thereof, the derivation method, computation, calculation, either a detection or the like May be.

[0060] In the present invention, the hydraulic oil amount Q _R to be consumed for the shift control is calculated from the change amount of the transmission gear ratio ratio, in the embodiment described above, as such variation, the rotation of the primary pulley 21 Although so as to use the actual variation amount based on the rotational speed N _S of the speed N _P and the secondary pulley 22, such variation may be calculated from the target speed ratio. In this case, there is an advantage that control responsiveness can be improved as compared with the case where the actual change amount is used.

Further, in the above-described embodiment, the transmission gear ratio ra
The line pressure correction amount ΔDL _SFT of the line pressure control duty is calculated from the change amount of tio.
0 stores the map, and according to this map,
The line pressure correction amount ΔDL _SFT may be directly determined from the change amount of the transmission gear ratio.

That is, in the above-described embodiment, in the process of calculating the line pressure correction amount ΔDL _SFT , for example, the above equation (1)
As shown in 7), although the oil pump volumetric efficiency eta _OPV is calculated as the variable of the engine speed Ne and the line pressure P _L, contrast, the variation of the transmission gear ratio Ratio using a map When determining the line pressure correction amount ΔDL _SFT directly, the engine speed N
Since e and the line pressure P _L is not taken into account in the calculation of the line pressure correction amount dL _SFT, although control accuracy decreases with a little, it is possible to greatly simplify the control system.

[0063]

As described above in detail, according to the line pressure control device of the continuously variable transmission according to the first aspect of the present invention, the consumption flow rate correlation value correlated with the consumption flow rate of the working fluid at the time of shift transition is calculated. It is derived by the consumption flow rate correlation value deriving means, the line pressure correction amount is determined by the line pressure correction amount determining means based on the consumption flow rate correlation value, and the target line pressure is corrected by the line pressure correction quantity to determine the actual line pressure. Since the pressure is controlled, the line pressure is controlled in consideration of the fluctuation of the line pressure due to the consumption flow rate of the working fluid, and there is an advantage that an optimum line pressure can be secured even during a shift transition. This also has the advantage that the belt-type continuously variable transmission can secure the belt clamping force and prevent belt slip even during shifting. Further, there is an advantage that belt durability can be improved.

According to the line pressure control device for a continuously variable transmission according to the second aspect of the present invention, the correlation value of the consumption flow rate of the working fluid during the shift transition is calculated based on the change ratio of the gear ratio, so that a special sensor is used. Therefore, there is an advantage that it is possible to obtain the same effect as the line pressure control device for a continuously variable transmission according to the first aspect without the necessity of providing any type.

[Brief description of the drawings]

FIG. 1 is a control block diagram showing a main configuration of a line pressure control device of a continuously variable transmission as one embodiment of the present invention.

FIG. 2 is a schematic diagram for explaining a power transmission system of a vehicle with a continuously variable transmission according to one embodiment of the present invention, and FIG.
Is a schematic configuration diagram of a power transmission system including the continuously variable transmission,
(B) is a block diagram of the continuously variable transmission.

FIG. 3 is a flowchart illustrating the contents of line pressure control by a line pressure control device of a continuously variable transmission as one embodiment of the present invention.

[Explanation of symbols]

 Reference Signs List 20 belt-type continuously variable transmission (CVT) 45 line pressure sensor 53 line pressure control means 54 target line pressure setting means 58 shift correction amount calculating means 58A consumption flow calculation means (consumption flow correlation value deriving means) 58B line pressure correction Volume determination means 63 Regulator valve (pressure regulating valve) 63A Solenoid for line pressure control

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) F16H 59:70 F16H 59:70 63:06 63:06 (72) Inventor Toru Hashimoto Shibago, Minato-ku, Tokyo No. 33-8, Mitsubishi Motors Corporation (72) Inventor Aju left Shimoda 5-33-8, Shiba, Minato-ku, Tokyo Mitsubishi Motors Corporation F-term (reference) 3J552 MA07 MA12 MA26 NA01 NB01 PA12 RA02 TA01 TA11 TB11 TB20 VA18Z VA32Y VA34Z VA37Z VA42Z VA48Z VA50W VA53W VA53Y VA74Z VB01Z VC01Z VC02Z VC03Z VD02Z

Claims (2)

[Claims] A target line pressure of a working fluid supplied to the continuously variable transmission is set according to an operation state of the vehicle to control a line pressure of the continuously variable transmission mounted on the vehicle. In a line pressure control device of a continuously variable transmission that controls an actual line pressure of the working fluid so as to reach a target line pressure, a consumption flow rate correlation value correlated with a consumption flow rate of the working fluid consumed during a shift shift is derived. A consumption flow correlation value deriving unit; and a line pressure correction amount determining unit that determines a line pressure correction amount based on the consumption flow correlation value derived by the consumption flow correlation value deriving unit. A line pressure control device for a continuously variable transmission, wherein the target line pressure is corrected in accordance with the line pressure correction amount determined by the means. 2. The continuously variable transmission according to claim 1, wherein the consumption flow rate correlation value deriving means is configured to derive the consumption flow rate correlation value based on a speed ratio change amount. Line pressure control device.