JP2653052B2 - Hydraulic control device for belt-type continuously variable transmission for vehicles - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a hydraulic control device for a belt-type continuously variable transmission for a vehicle, and in particular, controls a line hydraulic pressure based on a speed ratio deviation between an actual speed ratio and a target speed ratio. The present invention relates to the improvement of the hydraulic control device.

Prior artA pair of primary side variable pulleys and secondary side variable pulleys respectively provided on a primary rotation shaft and a secondary rotation shaft,
A transmission belt that is wound around the pair of variable pulleys to transmit power, and a pair of primary hydraulic cylinders and secondary hydraulic cylinders that change the effective diameter of the pair of variable pulleys, respectively. Continuously variable transmissions are known.

As a hydraulic control device for such a vehicle belt-type continuously variable transmission, (a) a first hydraulic fluid is supplied from a hydraulic source.
(B) supplying the hydraulic oil in the first line oil passage to one of the primary hydraulic cylinder and the secondary hydraulic cylinder; A control valve for simultaneously adjusting the speed ratio of the continuously variable transmission by causing the hydraulic oil in the other to flow to the second line oil passage; and (c) reducing the oil pressure in the second line oil passage to the first line oil passage. (D) adjusting the actual speed ratio of the continuously variable transmission to a target speed ratio determined in accordance with a driving state of the vehicle; A speed ratio control means for feedback-controlling the shift control valve in accordance with a control formula of a proportional operation has been considered. For example, Japanese Patent Application No. 61-37571 filed by the present applicant has been proposed. I have.

By the way, in such a hydraulic control device, it is inevitable that a deviation occurs between the actual speed ratio and the target speed ratio from the output hydraulic pressure characteristic of the shift control valve, but this speed ratio deviation always exceeds a certain value. If the control is to be made smaller, it is necessary to set the first line hydraulic pressure to a large value in advance in consideration of individual differences and aging of the continuously variable transmission, a pressure adjustment error of the pressure control valve, and the like. It was not always satisfactory in reducing power loss. For this reason, the applicant of the present application further discloses in Japanese Patent Application No. 61-172566, a first method for matching the speed ratio deviation between the actual speed ratio and the target speed ratio with a predetermined constant target deviation value.
A hydraulic control device has been proposed in which the first pressure regulating valve is feedback-controlled so as to obtain a line hydraulic pressure. In this way, the first line hydraulic pressure is the minimum necessary for realizing the speed ratio including the target deviation value regardless of individual differences and aging of the continuously variable transmission, pressure adjustment errors of the pressure control valve, and the like. Since the hydraulic pressure is controlled to the minimum, the power loss of the engine is reduced, the fuel efficiency of the vehicle is improved, and the accuracy required for the pressure regulating valve, the shift control valve, and the like is reduced.

Problems to be Solved by the Invention However, the feedback control of the first pressure regulating valve is for making the speed ratio deviation in the steady state where the speed ratio is substantially constant equal to the target deviation value. Although excellent effects are exhibited in the quasi-stationary state where the speed changes slightly, the difference between the actual speed ratio deviation and the target deviation value becomes much larger in the speed change in which the speed ratio changes than in the steady state. There has been a problem that the line oil pressure is set to an unnecessarily high pressure, so that the drive loss of the pump increases and the driving maneuverability is impaired due to a rapid change in the speed ratio. For this reason, conventionally, the feedback control of the first pressure regulating valve is performed only at the time of steady state, but the control form changes between the steady state and the shift state, so that it is not always satisfactory.

Means for Solving the Problems The present invention has been made in view of the above circumstances, and an object of the present invention is to obtain an appropriate first line oil pressure by feedback control of a first pressure regulating valve even during gear shifting. Is to be able to

In order to achieve the above object, the present invention provides a pair of primary variable pulleys and secondary variable pulleys provided on a primary rotary shaft and a secondary rotary shaft, respectively, and a pair of variable pulleys wound around the pair. In a vehicle belt-type continuously variable transmission including a transmission belt that is hung to transmit power, and a pair of a primary hydraulic cylinder and a secondary hydraulic cylinder that respectively change an effective diameter of the pair of variable pulleys,
(A) a first pressure regulating valve for regulating a hydraulic pressure in a first line oil passage to which hydraulic oil is supplied from a hydraulic pressure source to a first line hydraulic pressure;
(B) supplying the hydraulic oil in the first line oil passage to one of the primary hydraulic cylinder and the secondary hydraulic cylinder, and simultaneously causing the hydraulic oil in the other to flow to the second line oil passage, A shift control valve for adjusting the speed ratio of the stepped transmission, (c) a second pressure regulating valve for regulating the oil pressure in the second line oil passage to a second line oil pressure lower than the first line oil pressure, (d) Speed ratio control means for feedback-controlling the speed change control valve according to a proportional operation control formula such that an actual speed ratio of the continuously variable transmission matches a target speed ratio obtained according to a driving state of a vehicle; (E) using the difference between the speed ratio deviation and the target deviation value as an operation signal so that the speed ratio deviation between the actual speed ratio and the target speed ratio matches a predetermined target deviation value. The first adjustment for adjusting the first line hydraulic pressure; The feedback control of the valves 1
A hydraulic control device having line hydraulic control means,
A coefficient changing means for changing a coefficient of a control formula of feedback control by the first line hydraulic control means based on a speed ratio change rate of the continuously variable transmission is provided.

In this way, the coefficient of the control formula of the feedback control by the first line hydraulic control means is changed according to the speed ratio change rate of the continuously variable transmission.
The first pressure regulating valve can be feedback-controlled so that an appropriate first line oil pressure can be obtained even during a shift in which the difference between the speed ratio deviation and the target deviation value increases. Therefore, it is possible to prevent an increase in the drive loss of the pump and a decrease in driving operability.

Embodiment Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.

In FIG. 1, an output of an engine 10 provided in a vehicle is transmitted to a primary rotation shaft 16 of a belt-type continuously variable transmission 14 via a clutch 12. The belt-type continuously variable transmission 14 includes a primary rotation shaft 16 and a secondary rotation shaft 18, and the primary rotation shaft
The primary variable pulley 20 and the secondary variable pulley 22 having variable effective diameters attached to the rotary shaft 16 and the secondary rotating shaft 18, and the power is wound around the primary variable pulley 20 and the secondary variable pulley 22. , A primary hydraulic cylinder 26 and a secondary hydraulic cylinder 28 for changing the effective diameter of the primary variable pulley 20 and the secondary variable pulley 22. The primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 are formed to have the same pressure receiving area, and the outer diameters of the primary variable pulley 20 and the secondary variable pulley 22 are made equal to each other, so that the belt type The continuously variable transmission 14 is small. The primary variable pulley 20 and the secondary variable pulley 22 are fixed rotating bodies 31 and 32 fixed to the primary rotating shaft 16 and the secondary rotating shaft 18, respectively.
A movable rotating body 34 which is provided on the primary rotating shaft 16 and the secondary rotating shaft 18 so as to be relatively non-rotatable and axially movable and forms a V groove between the fixed rotating bodies 31 and 32, respectively. And 36. The belt-type continuously variable transmission 14
The output from the secondary rotary shaft 18 is a sub-transmission (not shown)
The power is transmitted to driving wheels of the vehicle via a differential gear device or the like.

The hydraulic control circuit for operating the belt-type continuously variable transmission 14 configured as described above is configured as described below. That is, the oil tank passes through a return path (not shown).
The hydraulic oil returned to 38 is sucked into an oil pump 42 as a hydraulic pressure source via a strainer 40 and a suction oil passage 41, and is connected to an input port 46 of a shift control valve 44 and a first pressure regulating valve 48 by a first line oil. Pumped to road 50. Oil pump 42
The engine 10 is driven by the engine 10 via a drive shaft (not shown). The first pressure regulating valve 48 controls a part of the hydraulic oil in the first line oil passage 50 in accordance with a first drive signal VD1 described later. By discharging to the second line oil passage 52, the first line
The hydraulic line fluid passage 50 pressure the first line pressure Pl ₁ two tone. The second line oil passage 52 is connected to the first
Discharge port 54 and second discharge port 56 and second pressure regulating valve 58
And are respectively connected to. The second pressure regulating valve 58 causes a part of the hydraulic oil in the second line oil passage 52 to flow out to the drain oil passage 60 in accordance with a second drive signal VD2 described later, so that the hydraulic pressure in the second line oil passage 52 is increased. The above first line hydraulic pressure Pl ₁
Relatively pressure lower second line pressure Pl ₂ two tone than. The first pressure regulating valve 48 and the second pressure regulating valve 58 are constituted by so-called electromagnetic proportional relief valves.

The shift control valve 44 is a so-called proportional control solenoid valve, and is connected to the input port 46, the first discharge port 54 and the second discharge port 56, and is connected to the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28. A pair of first output port 62 and second output port respectively connected via paths 29 and 30
A cylinder bore 66 formed in the valve body 65 so as to communicate with each other, a single spool valve element 68 slidably fitted in the cylinder bore 66, and neutral from both ends of the spool valve element 68. Biasing the spool valve element 68 in the neutral position by biasing it toward the first position.
A spring 70 and a second spring 72, and a first electromagnetic solenoid 74 provided at each end of the spool valve 68 to move the spool valve 68 against the urging force of the second spring 72 or the first spring 70. And a second electromagnetic solenoid 76. The spool valve 68 has 4
The two lands 78, 80, 82, and 84 are sequentially formed from one end, and the pair of lands 80 and 82 located at the middle portion are arranged in the axial direction of the spool valve element 68 when the spool valve element 68 is in the neutral position. It is formed at the same position as the first output port 62 and the second output port 64. Further, on the inner peripheral surface of the cylinder bore 66, a position facing the pair of lands 80 and 82 when the spool valve element 68 is in the neutral position, that is, the first output port 62 and the second output port 64 correspond to the cylinder bore 66. At the position that opens to the inner peripheral surface, the land 80
And a pair of first annular grooves slightly wider than 82
86 and a second annular groove 88 are formed. The first annular groove 86 and the second annular groove 88 form a restrictor whose flow sectional area continuously changes between the lands 80 and 82 in order to control the flow of hydraulic oil.

Thus, when the spool valve 68 is in the neutral position, the first output port 62 and the second output port 64 are evenly communicated with the input port 46 and the discharge ports 54 and 56 with a small flow area, and Hydraulic fluid in an amount sufficient to replenish the primary hydraulic cylinder 26 and the secondary hydraulic cylinder
28 and a small amount of hydraulic oil
Spilled from 4,56.

However, as the spool valve 68 is moved from the neutral position in one axial direction thereof, for example, in a direction approaching the second electromagnetic solenoid 76, that is, in the right direction in the drawing, the first output port 62 and the first discharge port 54 Of the second output port 64 and the input port 46 while the flow cross-sectional area of the
The working hydraulic pressure output from the first output port 62 to the primary hydraulic cylinder 26 is reduced to the working hydraulic pressure output from the second output port 64 to the secondary hydraulic cylinder 28. It is lower than the comparison. As a result, the balance between the thrust of the primary hydraulic cylinder 26 and the thrust of the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is lost, so that the hydraulic oil flows into the secondary hydraulic cylinder 28, while the primary hydraulic cylinder The hydraulic oil in 26 leaks out and the belt-type continuously variable transmission 14
(The rotation speed N _out of the secondary rotation shaft 18 / the rotation speed N _in of the primary rotation shaft 16) becomes smaller.

Conversely, as the spool valve element 68 is moved from the neutral position toward the first electromagnetic solenoid 74, that is, to the left in the drawing, the flow cross-sectional area between the first output port 62 and the input port 46 becomes continuous. While the flow cross-sectional area between the second output port 64 and the second discharge port 56 is increased.
The working oil pressure output to 26 is higher than the working oil pressure output from second output port 64 to secondary hydraulic cylinder 28. As a result, the balance between the thrust of the primary hydraulic cylinder 26 and the thrust of the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is lost, so that the hydraulic oil flows into the primary hydraulic cylinder 26 while the secondary hydraulic cylinder 26 Hydraulic oil in 28 leaks out,
The speed ratio e of the belt-type continuously variable transmission 14 increases. As described above, the shift control valve 44 functions as a switching valve that supplies high-pressure hydraulic oil to one of the hydraulic cylinders 26 and 28 and supplies low-pressure hydraulic oil to the other, and continuously adjusts the flow rate of hydraulic oil. It has a flow control valve function.

On the other hand, between the second line oil passage 52 and the connection oil passages 29 and 30, throttle oil passages 112 and 116 having throttles 110 and 114 for restricting the flow of hydraulic oil are connected. As a result, the hydraulic oil cylinder 26 or 28 on the high pressure side (drive side) passes through the throttle oil passage 112 or 116 and the second line oil passage 52
Since hydraulic oil flows into the hydraulic pressure P _out of the output hydraulic pressure, i.e. the pressure P _in and secondary side hydraulic cylinder 28 of the primary side hydraulic cylinder 26 of the shift control valve 44, as shown in Figure 2 In the state where the spool valve element 68 is held at the neutral position (speed ratio control value V ₀ = 0), the spool valve 68 is brought close to the second line hydraulic pressure Pl ₂ and the low pressure side (driven side) hydraulic cylinder 26 or 28
Hydraulic P _in or P _out of the inner is allowed to substantially coincide with the second line pressure Pl _2.

The belt-type continuously variable transmission 14 of the vehicle has a first rotation sensor 9 for detecting the rotation speed N _in of the primary rotation shaft 16.
0, and the secondary-side second rotation sensor 92 for detecting the rotational speed N _out is provided in the rotary shaft 18, the rotational speed N _in from their first rotation sensor 90 and the second rotation sensor 92
Rotation signal S representative of the rotation signal SR1 and the rotational speed N _out represents the
R2 is output to the controller 94. The engine 10 has a throttle valve opening θ as an amount representing the required output of the vehicle.
a throttle sensor 96 for detecting the _th, and the engine rotation sensor 98 is provided for detecting the engine rotational speed N _e, representing the throttle valve opening theta _th from those throttle sensor 96 and the engine rotation sensor 98 rotation signal representing the throttle signal Sθ and the engine rotational speed N _e
SE is output to the controller 94.

The controller 94 is a so-called microcomputer including a CPU 102, a ROM 104, a RAM 106, and the like.
The input signal is processed according to a program previously stored in the ROM 104 while utilizing the storage function of the M106, and the first pressure regulating valve 48 is processed.
And a speed ratio for driving the first electromagnetic solenoid 74 and the second electromagnetic solenoid 76 for controlling the speed ratio e while supplying the first drive signal VD1 and the second drive signal VD2 to the second pressure regulating valve 58, respectively. Supply signals RA1 and RA2 to them.

Hereinafter, the operation of this embodiment will be described with reference to the flowchart of FIG.

First, by executing step S1, the rotation speed N _in of the primary rotation shaft 16 and the rotation speed of the secondary rotation shaft 18 are set.
N _out , throttle valve opening θ _th , and engine speed
N _e is the rotation signal SR1 and SR2, a throttle signal S.theta, are read into the RAM106 based on the rotation signal SE. Then, the determined actual output torque T _e of the engine 10 on the basis of well-known relationship stored in advance at step S2 ROM 104 to the throttle valve opening theta _th and the engine speed N _e, the further step S3 the throttle The target rotation speed N _in ^{* of the} primary rotation shaft 16 is determined based on the valve opening degree θ _{th and the} like. The relationship for determining the target rotation speed N _in ^* is, for example, that shown in FIG. 4, which is obtained in advance so that the engine 10 exclusively operates on the minimum fuel consumption rate curve shown in FIG. is there.

Subsequent step S4 rotational speed N _in and N the speed ratio e of the real differences from _out CVT 14 is calculated, from the rotational speed in step S5 N _out and the target rotational speed N _in ^* target speed ratio e ^* Is calculated. Then, in step S6,
Speed ratio control value V ₀ which is for operating the shift control valve 44 so that the speed ratio e is equal to the target speed ratio e ^* is, it like the speed ratio e, the target speed ratio e ^* and the control constant K ₀ It is calculated based on the following equation (1), which is a control equation for proportional operation, based on the following equation. In step S17 described later, the so rotational speed N _out of the spool 68 is moved to the left secondary side rotary shaft 18 is increased when the speed ratio control value V ₀ is positive The speed ratio signal RA2 is output, and when the speed ratio signal RA2 is negative, the spool valve element 68 is moved rightward to output the speed ratio signal RA1 such that the rotation speed N _in of the primary rotation shaft 16 increases. . The speed ratio control value V magnitude of _zero speed ratio signal RA1 or the speed ratio signal RA2 size, i.e. spool 68
Corresponding to the amount of movement.

V ₀ = K ₀ (e ^* −e) / e (1) Then, in step S7, the change rate de ^* / dt of the target speed ratio e ^* is calculated as the speed ratio change rate of the continuously variable transmission 14: It is calculated according to equation (2). In the equation (2), e ^* _(n) is the latest target speed ratio, e ^* _(n-1) is the target speed ratio obtained one time before, and h is the target speed ratio calculated. Is the interval time. Further, in the following steps S8 and S9, the rate of change de ^* / dt is determined from the relationship stored in the ROM 104 in advance.
Are determined based on the coefficients K _P and K _I , respectively. Note that other arithmetic expressions such as Expression (3) can be used instead of Expression (2).

de ^* / dt = (e ^* _(n) −e ^* _(n−1) ) / h ・ ・ ・ (2) de ^* / dt = (− 2e ^* _(n−4) −e ^* _(n−3) ) + E ^* _(n-1) + 2e ^* _(n) )
/ 10h ··· (3) followed by step S10 is executed, the actual output torque T whether _e is positive, or positive torque condition or the engine brake is outputted power from the engine 10 of the engine 10 It is determined whether it is in the state. The reason why such a judgment is necessary is that the power transmission direction differs between the positive torque state and the engine brake state, so that the oil pressure change characteristic with respect to the speed ratio e of the hydraulic cylinders 26 and 28 changes.
When the output torque T _e is determined to be positive, first, by the step S11 is executed, for generating a necessary and sufficiently squeezing force against the drive belt 24 secondary-side hydraulic cylinder 28 Hydraulic pressure (target hydraulic pressure) P _out ′
Is determined, the second line oil pressure Pl ₂ to be adjusted by the second pressure adjusting valve 58 is determined. That is, first, the ROM 10
Actual output torque T _e of the engine 10 from the relationship 4 to the stored equation (4), the thrust (calculated value) of the optimum secondary side hydraulic cylinder 28 based on the actual speed ratio e calculated W _out ' Then, from the following equation (5), this thrust W _out ′, the secondary hydraulic cylinder 28
The second line hydraulic pressure Pl ₂ is calculated based on the pressure receiving area A _out and the rotation speed N _out of the secondary rotation shaft 18. The hydraulic pressure determined by the equation (5) is originally the target hydraulic pressure P _out 'in the secondary hydraulic cylinder 28, but in this embodiment, the second line hydraulic passage 52 and the hydraulic cylinder 28
Is connected, the hydraulic pressure of equation (5) can be used as the second line hydraulic pressure Pl ₂ .

W _out '= f (T _e , e) (4) Here, the above equation (4) is obtained in advance in order to make the tension of the transmission belt 24, that is, the squeezing pressure on the transmission belt 24 a necessary and sufficient value, and the thrust W _out ′ is equal to the output torque _Te and It is varied in relation to the speed ratio e. In the relation of the expression (5), the second term is for correcting the second line oil pressure Pl ₂ (P _out ′) by subtracting the centrifugal oil pressure that increases with the rotation speed N _out from the first term.
C ₂ of the second term is the centrifugal force correction coefficient is predetermined from the specific gravity of the specifications and working oil in the secondary-side hydraulic cylinder 28.

In the following step S12, the first pressure regulating valve 48 is set so that the hydraulic pressure (target hydraulic pressure) Pin ′ _{in the} primary hydraulic cylinder 26 for generating necessary and sufficient thrust to achieve the target speed ratio is obtained. The pre-program value Pl _{1 (P) of} the first line oil pressure to be adjusted is determined. That is, first, the actual output torque T _e and the target speed ratio e ^* positive drive propulsive force ratio on the basis of the gamma ₊ (the secondary side of the engine 10 from the relationship shown in advance ROM104 the stored equation (6) Hydraulic cylinder 2
8 with thrust W _in) thrust W _out / primary-side hydraulic cylinder 26 is calculated in the following equation (7) from the thrust ratio gamma ₊ and secondary side hydraulic primary hydraulic cylinder from the thrust W _out 'of the cylinder 28 26 thrusts W _in 'are required. Then, the following equation (8)
From the primary-side hydraulic cylinder 26 thrust W _in ', the pressure receiving area A _in the primary side hydraulic cylinder 26, the rotational speed N _in the primary-side rotary shaft 16
The hydraulic pressure (calculated value) P _in ′ is calculated based on
It is from the following equation (9) to calculate the first line pressure Pl _{1 (P)} based on the hydraulic P _in 'and margin pressure [Delta] P.

γ ₊ = f (T _e , e ^* ) (6) Pl _{1 (P)} = P _in '+ ΔP (9) The above equation (6) shows a relationship obtained in advance so that the thrust ratio γ ₊ can be determined over a wide range of operating conditions. , so that the target speed ratio e ^* and the actual thrust ratio which are determined in relation with the output torque T _e gamma ₊ is obtained from this relationship, is for obtaining a first line pressure Pl _{1 (P).}
In the relation of the above equation (8), the second term is the rotational speed.
The centrifugal oil pressure that increases with N _in is corrected by subtracting it from the first term, and C _{1 in} the second term is the primary hydraulic cylinder.
It is determined in advance from the specifications of 26 and the specific gravity of the hydraulic oil. Further, the above equation (9) is equivalent to the hydraulic pressure P obtained by the equation (8).
_in 'by adding the margin oil pressure ΔP to the first line oil pressure.
Pl _{1 (P)} is determined.

Here, the surplus oil pressure ΔP is determined by the speed ratio e and the target speed ratio e ^*.
This is necessary in order to reduce the speed ratio deviation | e ^* −e | / e. That is, the output hydraulic pressure characteristic of this embodiment is the second hydraulic pressure characteristic.
The speed ratio control value V ₀ is shown in FIG.
From the expression, the actual speed ratio e and the target speed ratio e ^* can be completely matched when the speed ratio control value V ₀ is 0, but otherwise the actual speed ratio e ^* is the is / e occurs | e ^* -e | magnitude of the speed ratio deviation corresponding to the speed ratio control value V ₀ between the speed ratio e and the target speed ratio e ^*. The speed ratio deviation | e ^* −e | / e becomes smaller because the slope of the hydraulic characteristic becomes steeper when the first line oil pressure Pl ₁ is increased, and becomes smaller when the first line oil pressure Pl ₁ is decreased. Becomes large because it becomes moderate. However, the first line hydraulic pressure Pl ₁
Is increased, the driving loss of the pump 42 increases accordingly, and the margin oil pressure ΔP is determined at the equilibrium point between the driving loss and the steady-state deviation that are opposite to each other.

On the other hand, when it is determined in step S10 that the vehicle is in the engine braking state, the power transmission directions in the belt-type continuously variable transmission 14 are reversed, so that steps S13 and S13 substantially similar to steps S11 and S12 are performed. By executing S14, the pre-program value Pl _{1 (P)} of the second line oil pressure Pl ₂ and the first line oil pressure is determined. That is, in step S13, the following equation (1
0) in the output torque from the relationship shown T _e, with thrust W _in the optimal primary hydraulic cylinder 26 based on the speed ratio e 'is calculated, to be supplied from the following equation (11) on the primary side hydraulic cylinder 26 hydraulic P _in ', that is, the second line pressure Pl ₂ is calculated. In Step S14, the output torque T _e from the following equation (12), the thrust ratio gamma based on the target speed ratio e ^* _- calculates a, the following equation (13) above the thrust ratio gamma _- to obtain The thrust W _out ′ of the secondary hydraulic cylinder 28 is changed to the thrust ratio γ ₋
And 'determined based on, further, (14) Hydraulic P _out needed in the secondary side hydraulic cylinder 28 from the expression' thrust W _in the primary side hydraulic cylinder 26 with Request, the pressure P from the following equation (15) _out 'and the first line oil pressure based on the margin oil pressure ΔP
Calculate Pl _{1 (P)} .

W _in '= f (T _e , e) (10) _{γ - = f (T e,} e *) ··· (12) W out '= W in' · γ - ··· (13) Pl _{1 (P)} = P _out '+ ΔP (15) When the second line pressure Pl ₂ and the first line pressure Pl _{1 (P)} are calculated in this way, the next step S15 is executed. According to the following equation (16), step S12 or S
Pre-program value Pl of the first line hydraulic pressure calculated in 14
_A feedback term is added to _{1 (P)} , and the speed ratio deviation | e ^* −
The first line pressure Pl ₁ to be regulated by the first pressure regulating valve 48 is determined so that e | / e matches a predetermined constant target deviation value ε.

Here, the target deviation value ε is determined by an experiment or a simulation as a value at which the best fuel efficiency is obtained in a comprehensive driving state of the vehicle such as, for example, 10-mode running, or the target fuel consumption is minimized in various driving states. For example, the minimum value of the steady-state deviation required for setting the target deviation value ε is set by various means. Then, the first line hydraulic pressure Pl ₁ is determined so that the speed ratio deviation | e ^* −e | / e matches this target deviation value ε.

For example, when the actual first line hydraulic pressure is smaller than the value that should be, the speed ratio deviation | e ^* −e | / e is large, and | e ^* −e | / e−ε is positive. Value. Therefore, an amount proportional to the positive operation signal is added to the first line oil pressure Pl _{1 (P)} , and the following steps S16 and S17 are executed based on the first line oil pressure Pl ₁ thus obtained. As a result, the actual first line hydraulic pressure is increased, the speed ratio deviation | e ^* -e | / e is reduced, and finally is made to coincide with the target deviation value ε. When the actual first line hydraulic pressure is larger than the value that should be, the steady-state error | e ^*
Since −e | / e is small, | e ^* −e | / e−ε has a negative value. Therefore, the amount proportional to the negative operation signal is subtracted from the first line oil pressure Pl _{1 (P),} and based on the thus obtained first line oil pressure Pl ₁ , a later-described step S16 is performed.
By executing steps S17 and S17, the actual first line hydraulic pressure is decreased, the speed ratio deviation | e ^* -e | / e is increased, and finally matched with the target deviation value ε. Then, when the speed ratio deviation | e ^* −e | / e is matched with the target deviation value ε, the actual first line hydraulic pressure becomes the minimum necessary to realize the speed ratio e including the target deviation value ε. Oil pressure value. That is, in this step S15, the first line oil pressure Pl _{1 (P)} calculated in step S12 or S14 is
By correcting based on the speed ratio deviation | e ^* −e | / e,
The control is performed so that the actual first line hydraulic pressure becomes the minimum necessary value. As a result, the drive loss of the pump 42 is reduced, and the fuel efficiency of the vehicle is improved.

On the other hand, the values determined in steps S8 and S9 are applied to the coefficient K _P in the proportional operation term and the coefficient K _I in the integral operation term in the above equation (16), and the rate of change de ^* of the target speed ratio e ^{* *} The first line oil pressure Pl ₁ is determined according to / dt. This is because equation (16) is used for feedback control so that the speed ratio deviation | e ^* −e | / e in the steady state where the speed ratio e is substantially constant is equal to the target deviation value ε. _I ,
K _{P is} also set to a constant value in advance based on the steady state. However, at the time of gear shifting in which the speed ratio e changes, the difference between the speed ratio deviation | e ^* −e | / e and the target deviation value ε is extremely small. Therefore, the first line hydraulic pressure Pl ₁ becomes excessively large, so that the coefficients K _P and K _I are changed based on the change rate de ^* / dt of the target speed ratio e. The relationship previously stored in the ROM 104 for determining the coefficients K _P and K _I in the steps S8 and S9 is, for example, as shown in FIG. Accordingly, the coefficients K _P and K _I are reduced as the rate of change de ^* / dt increases.
Incidentally, FIG. 6 is the rate of change de ^* / dt is positive, i.e. it is the case at acceleration speed of increasing the speed ratio e, even if during deceleration speed rate of change de ^* / dt is negative, As the rate of change de ^* / dt increases to the negative side, the coefficients K _P and K _I are reduced.

When the first line oil pressure Pl ₁ and the second line oil pressure Pl ₂ are determined in this way, step S16 is executed next, and the first adjustment is performed so that the oil pressures Pl ₁ and Pl ₂ are obtained. valve 48, the first line pressure control value V ₁ and the second line pressure control value V ₂ for operating the second pressure regulating valve 58 are determined, respectively. Thereafter, by executing the last step S17, the first line hydraulic pressure control value V ₁ and the second
Based on the line hydraulic pressure control value V ₂ and the speed ratio control value V ₀ determined in step S6, drive signals VD1, VD2
And speed ratio signals RA1 and RA2 are output. This allows
Speed ratio e, first line oil pressure Pl ₁ and second line oil pressure Pl
₂ is controlled by the shift control valve 44, the first pressure regulating valve 48, and the second pressure regulating valve 58, and thereafter, step S1 and subsequent steps are repeatedly executed.

As described above, in the present embodiment, the first pressure regulating valve 48 performs the feedback control so that the speed ratio deviation | e ^* −e | / e between the speed ratio e and the target speed ratio e ^* matches the target deviation value ε. Therefore, the first line hydraulic pressure regulated by the first pressure regulating valve 48 is set to the minimum hydraulic pressure value necessary to realize the speed ratio including the target deviation value ε, and the drive loss of the pump 42 and the The power loss of the engine 10 is reduced, and the fuel efficiency of the vehicle is improved.

Further, since the first line oil pressure is adjusted based on the speed ratio deviation | e ^* −e | / e, the first line oil pressure in the pressure adjustment system including the first pressure adjustment valve 48 and the second pressure adjustment valve 58 is adjusted. If there is a pressure adjustment error of the hydraulic pressure or the second line hydraulic pressure,
The actual first line oil pressure is always controlled to the minimum necessary oil pressure value, even if the characteristics of 44 vary. That is, if a certain correlation is ensured between the actual first line pressure and the first line pressure control value V _1, the absolute value of the hydraulic pressure is not permissible even not very accurate. Therefore, it is not always necessary to employ a high-precision one as the pressure regulation system or the shift control valve 44, and it is possible to reduce the manufacturing cost of such a system.

On the other hand, in the present embodiment, the coefficients K _P and K _I of the control equation (16) for performing the feedback control are determined in consideration of the rate of change de ^* / dt of the target speed ratio e ^* , and are sufficient in a steady state. While the feedback effect is obtained, the feedback effect is reduced during gear shifting. Therefore, the speed ratio deviation | e ^* -e | without / e and the first line pressure Pl ₁ even during shifting the difference between the target deviation value ε is increased becomes excessive, the first line pressure Pl ₁ It is possible to prevent an increase in drive loss of the pump 42 and a decrease in driving maneuverability due to excessive pressure increase, and to apply the control formula (16) under all operating conditions. When the rate of change de ^* / dt increases, the coefficients K _P and K _I become 0 and the feedback effect is completely eliminated. In this case, however, it is not necessary to control the first line pressure Pl ₁ as strictly as in the steady state. Therefore, there is no problem even if the accuracy of the pressure regulation system or the like is reduced as described above.

Further, in the present embodiment shown in FIG. 6, the coefficient K _P, since adapted K _I is continuously changed according to the rate of change de ^* / dt, the coefficient K _P, K _I Of the first line hydraulic pressure Pl compared to a case where the control mode is changed stepwise or a control type is changed between a steady state and a shift as in the conventional case.
₁ can be controlled smoothly.

In this embodiment, of the series of signal processing by the controller 94, steps S6 and S17 correspond to speed ratio control means, steps S15, S16 and S17 correspond to first line hydraulic pressure control means, and steps S7 and S8. And S9 correspond to coefficient changing means.

As mentioned above, although one Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

For example, the control equation (16) of the embodiment includes a proportional operation term and an integral operation term, but it is also possible to use a control equation including only one of them or a differential operation term.

Further, in the above embodiment, de ^* / dt is used as the speed ratio change rate of the continuously variable transmission 14, but the actual change rate de / dt of the speed ratio e can be used. However, since the change in the speed ratio e has a delay with respect to the change in the target speed ratio e ^* , it is more effective to use de ^* / dt to prevent the first line oil pressure Pl ₁ from being excessively boosted at the start of the shift. On the other hand, de / dt is more desirable for preventing an excessive pressure increase during the shift and near the end of the shift. Therefore, ideal control can be performed by selectively using the larger one of de ^* / dt and de / dt.

Also, or using a ^{(1 / e) de * /} dt instead of de ^* / dt, d
It is also possible to use N _in ^* / dt, (1 / N _in ^* ) dN _in ^* / dt, dN _in / dt, etc. as the speed ratio change rate.

Further, in the above-described embodiment, the pre-program value Pl _{1 (P)} of the first line oil pressure is obtained in advance, and the final first line oil pressure Pl ₁ is determined by adding a feedback term thereto. It is also possible to determine the first line oil pressure Pl ₁ only by the feedback term without obtaining the preprogram value Pl _{1 (P)} .

Further, in the above-described embodiment, | e ^* −e | / e is used as the speed ratio deviation, but feedback control can also be performed based on | e ^* −e |. If the rotation speed N _out of the secondary rotation shaft 18 is determined, the speed ratio e and the rotation speed of the primary rotation shaft 16 are determined.
N _in (or the engine speed N _e ) has a fixed relationship, so feedback is performed based on the deviation | N _in −N _in ^* | / N _in ^* between the speed N _in and the target speed N _in ^*. By controlling, the speed ratio deviation | e ^* −e | / e can be made to coincide with the target deviation value ε.

Although not specifically exemplified, the present invention can be embodied in various modified and improved forms based on the knowledge of those skilled in the art without departing from the spirit thereof.

[Brief description of the drawings]

FIG. 1 is a configuration diagram of a hydraulic control device of a vehicle belt-type continuously variable transmission according to an embodiment of the present invention. FIG. 2 is a diagram showing output hydraulic pressure characteristics of the transmission control valve of FIG. FIG. 3 is a flowchart for explaining the operation of the apparatus shown in FIG. FIG. 4 is a diagram showing the relationship between the target rotation speed of the primary rotation shaft and the throttle valve opening in the apparatus shown in FIG.
FIG. 5 is a diagram showing a minimum fuel consumption rate curve of the engine of FIG. FIG. 6 shows the coefficients in the flowchart of FIG.
FIG. 9 is a diagram showing a relationship between a predetermined change rate de ^* / dt for determining K _P and K _I. 14: Belt-type continuously variable transmission 16: Primary rotating shaft, 18: Secondary rotating shaft 20: Primary variable pulley, 22: Secondary variable pulley 24: Power transmission belt 26: Primary hydraulic cylinder 28: Secondary Side hydraulic cylinder 44: Transmission control valve 48: First pressure regulating valve, 58: Second pressure regulating valve 94: Controller Pl ₁ : First line hydraulic pressure, Pl ₂ : Second line hydraulic pressure K _P , K _I : Coefficient Steps S7, S8 , S9: coefficient changing means, steps S6, S17: speed ratio control means, steps S15, S16, S17: first line hydraulic pressure control means

Claims (3)

(57) [Claims] 1. A pair of primary variable pulleys and a secondary variable pulley provided on a primary rotary shaft and a secondary rotary shaft, respectively, and a transmission belt wound around the pair of variable pulleys to transmit power. And a pair of primary hydraulic cylinders and secondary hydraulic cylinders for respectively changing the effective diameters of the pair of variable pulleys, wherein the hydraulic oil is supplied from a hydraulic source. A first pressure regulating valve for regulating the oil pressure in the one-line oil passage to the first line oil pressure, and supplying the hydraulic oil in the first line oil passage to one of the primary hydraulic cylinder and the secondary hydraulic cylinder while simultaneously supplying the other. A shift control valve for adjusting a speed ratio of the continuously variable transmission by causing hydraulic oil in the second line to flow out to the second line oil passage; and a hydraulic pressure in the second line oil passage lower than the first line oil pressure. 2nd lie A second pressure regulating valve that regulates hydraulic pressure; and the shift control valve according to a proportional operation control formula such that an actual speed ratio of the continuously variable transmission matches a target speed ratio obtained according to a driving state of a vehicle. Speed ratio control means for feedback-controlling the first line hydraulic pressure so that a speed ratio deviation between the actual speed ratio and the target speed ratio matches a predetermined target deviation value. A first line hydraulic pressure control unit that performs feedback control of the pressure regulating valve, wherein a coefficient of a control formula of feedback control by the first line hydraulic pressure control unit based on a speed ratio change rate of the continuously variable transmission. A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising: a coefficient changing means for changing the coefficient. 2. The control expression of feedback control by said first line pressure control means is: a first line pressure to be regulated by said first pressure regulating valve is Pl ₁ , a pre-programmed first line pressure pre-program value. Where Pl _{1 (P)} , the actual speed ratio is e, the target speed ratio is e ^* , the target deviation value is ε, the coefficient of the proportional action term is K _P , and the coefficient of the integral action term is K ₁ , formula 2. The hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 1, wherein the hydraulic control device is represented by: 3. The coefficient changing means decreases the coefficients K _P and K _I of the control formula as the rate of change de ^* / dt increases when the target speed ratio e ^* increases. The hydraulic control device for a vehicle belt-type continuously variable transmission according to claim 2.