JPH0359298B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to an improvement in a hydraulic control device for a belt-type continuously variable transmission for a vehicle.

Prior Art A pair of primary variable pulleys and a secondary variable pulley provided on the primary rotating shaft and the secondary rotating shaft, respectively, and a transmission belt that is wound around the pair of variable pulleys to transmit power; A belt-type continuously variable transmission for a vehicle is known that includes a pair of primary hydraulic cylinder and secondary hydraulic cylinder that respectively change the effective diameters of the pair of variable pulleys. The speed ratio and the tension of the transmission belt of such a continuously variable transmission are determined by the secondary hydraulic cylinder (hydraulic cylinder installed on the driven rotating shaft), as described in, for example, Japanese Patent Laid-Open No. 52-98861. The tension of the transmission belt is exclusively controlled by regulating the supplied hydraulic pressure, and the amount of hydraulic fluid supplied to the primary hydraulic cylinder (hydraulic cylinder installed on the drive side rotating shaft) or the hydraulic fluid discharged from it is controlled. The speed ratio is controlled exclusively by adjusting the speed ratio.

In such a hydraulic control device, one type of line hydraulic pressure regulated in relation to the speed ratio etc. is prepared, and this is supplied exclusively to the secondary side hydraulic cylinder that maintains the tension of the transmission belt, and also controls the speed ratio. It is also supplied to the primary hydraulic cylinder via a flow control valve. Therefore, the flow rate of the hydraulic oil supplied to the primary hydraulic cylinder or the hydraulic oil discharged from it changes depending on the line oil pressure, that is, the speed ratio of the continuously variable transmission, etc. It is unavoidable that it will be influenced by the speed ratio etc. Therefore, there have been cases in which sufficient transient response has not been obtained in controlling the gear ratio. Additionally, when the direction of power transmission is reversed during vehicle engine braking, the primary hydraulic cylinder essentially controls the tension of the transmission belt, and the secondary hydraulic cylinder exclusively controls the speed ratio. Therefore, there was a drawback that the control characteristics of the tension and speed ratio of the transmission belt could not be obtained properly.

On the other hand, as described in Japanese Patent Publication No. 58-29424, there is a control valve (four-way valve) that changes the speed ratio by simultaneously supplying hydraulic oil from a hydraulic source to one side of the hydraulic cylinder and letting it flow out from the other side. )
A belt-type continuously variable transmission is provided that includes a control valve and an electromagnetic relief valve that regulates the pressure of hydraulic fluid flowing out from the control valve.

In this type of continuously variable transmission, a relatively high hydraulic pressure from a hydraulic source is applied to the hydraulic cylinder located on the side (drive side) where the internal hydraulic pressure is high in the power transmission state among the two hydraulic cylinders, Since the hydraulic pressure regulated by the electromagnetic relief valve is applied to the opposite hydraulic cylinder, the tension and speed ratio of the power transmission belt can be suitably controlled even if the power transmission direction is reversed.

Problems to be Solved by the Invention However, in such conventional continuously variable transmissions, the pressure of the hydraulic pressure source is not controlled and is only maintained at a constant pressure by a normal relief valve, so the transmission If the oil pressure value in the hydraulic cylinder changes according to the torque or speed ratio, the speed ratio change speed, that is, the speed change responsiveness may not be constant. On the other hand, if the pressure of the hydraulic power source is set high in anticipation of a large excess oil pressure in order to obtain a sufficient gear ratio change speed over the entire operating condition, the power loss required to constantly maintain that pressure will be large. There was a drawback.

Means for Solving the Problems The present invention has been made against the background of the above-mentioned circumstances, and its gist is to provide a pair of primary rotating shafts each provided on a primary rotating shaft and a secondary rotating shaft. A variable pulley and a secondary variable pulley, a transmission belt that is wound around the pair of variable pulleys to transmit power, and a pair of primary hydraulic cylinders and secondary side that change the effective diameters of the pair of variable pulleys, respectively. A hydraulic control device for a vehicle belt-type continuously variable transmission equipped with a hydraulic cylinder, comprising: (1) adjusting the pressure of hydraulic oil supplied from a hydraulic source to a first line hydraulic pressure based on the output state of the engine; a first pressure regulating valve; and (2) a spool valve element that is movable in the axial direction, and as the spool valve element moves, hydraulic oil regulated to the first line hydraulic pressure is supplied to the primary side hydraulic cylinder. By simultaneously supplying hydraulic oil to one of the secondary hydraulic cylinders and flowing out the hydraulic oil in the other, the effective diameters of the primary variable pulley and the secondary variable pulley are changed to change the speed ratio of the continuously variable transmission. (3) regulating the pressure of the hydraulic fluid flowing out from the other of the primary hydraulic cylinder and the secondary hydraulic cylinder through the transmission control valve based on the output state of the engine; (4) a third pressure regulating valve that generates a constant control hydraulic pressure for driving the spool valve; (5) a second pressure regulating valve that generates a second line hydraulic pressure that is lower than the line hydraulic pressure;
and an electromagnetic control valve device that controls the moving position of the spool valve element by applying the control hydraulic pressure to the spool valve element.

In this way, the first line hydraulic pressure and the second line hydraulic pressure are prepared by the first pressure regulating valve and the second pressure regulating valve. The flow rate of hydraulic oil supplied to or discharged from one of the secondary hydraulic cylinders is determined. Therefore, the speed ratio change speed is the first regardless of the speed ratio of the continuously variable transmission.
Since it is determined according to the differential pressure between the line oil pressure and the second line oil pressure, sufficient transient response characteristics for speed ratio control can be obtained. In addition, by controlling the first pressure regulating valve in relation to the output state of the engine, the first line oil pressure is controlled to a necessary and sufficient value so that a sufficient gear ratio change speed is obtained and no power loss occurs. At the same time, by controlling the second pressure regulating valve in relation to the speed ratio and transmission torque, the second line oil pressure is controlled to a necessary and sufficient value within a range that does not cause slippage of the transmission belt, thereby reducing vehicle power loss. significantly reduced.

Furthermore, when a constant control oil pressure is applied to the spool valve element by the electromagnetic control valve device, the spool valve element is moved and the output of the speed change control valve is switched. For this purpose, a spool valve is formed by a link whose one end is connected to an operating member for changing the speed ratio, and whose other end is engaged with a movable rotating body that constitutes a part of a variable pulley and moves in the axial direction. Compared to a type of speed change control valve in which the speed change control valve is driven, the degree of freedom in control is greatly increased because the speed change pattern is not limited to the speed change pattern determined by the specifications of the link mechanism. Further, since there is no need to mechanically connect the speed change control valve to the variable pulley via a link mechanism, there is an advantage that the arrangement of both can be made freely. Furthermore, there is an advantage that a decrease in control accuracy caused by link rattling that is unavoidable in a link mechanism is eliminated.

Moreover, when the first line hydraulic pressure or the second line hydraulic pressure is applied to the spool valve element by the electromagnetic control valve device, the first line hydraulic pressure and the second line hydraulic pressure are controlled by the speed ratio or transmission of the belt type continuously variable transmission. Although it is inevitable that the shift response will vary due to changes related to torque,
According to the hydraulic control device of the present invention, the control hydraulic pressure that is constantly controlled by the third pressure regulating valve is applied to the spool valve element by the electromagnetic control valve device, so there is an advantage that stable shift response can be obtained. be.

Here, the speed change control valve is preferably provided in a cylinder bore into which the spool valve element is slidably fitted, and at both ends of the cylinder bore, and on a pressure receiving surface at the end of the spool valve element. The electromagnetic control valve device includes a pair of oil-tight first control oil chambers and a second control oil chamber for applying control oil pressure, and the electromagnetic control valve device includes a pair of first control oil chambers and a second control oil chamber that are oil-tight. It is composed of a pair of electromagnetic on-off valves that are duty-controlled to supply the control hydraulic pressure to the oil chamber.

Further, the third pressure regulating valve preferably includes the first pressure regulating valve.
The first line oil passage that guides the line oil pressure or the second line oil passage.
It is composed of a pressure reducing valve and a throttle connected in series between a second line oil passage that guides line oil pressure and a drain oil passage, and the control oil pressure is generated between the pressure reducing valve and the throttle. .

Further, the third pressure regulating valve preferably includes the first pressure regulating valve.
It is provided in an oil passage that guides the hydraulic oil flowing out from the pressure regulating valve or the second pressure regulating valve, and regulates the pressure of the hydraulic oil to a constant control oil pressure.

Embodiment Hereinafter, an embodiment of the present invention will be described in detail based on the drawings.

In FIG. 1, an engine 1 installed in a vehicle
The zero output is transmitted via the clutch 12 to the primary rotating shaft 16 of the belt type continuously variable transmission 14.

The belt type continuously variable transmission 14 has a primary rotating shaft 16
and a secondary rotating shaft 18, a variable primary pulley 20 and a variable secondary pulley 22 with variable effective diameters attached to the primary rotating shaft 16 and the secondary rotating shaft 18, and 20
and a transmission belt 24 that is wrapped around the secondary variable pulley 22 to transmit power, and a primary hydraulic cylinder 26 and a secondary hydraulic cylinder that change the effective diameters of the primary variable pulley 20 and the secondary variable pulley 22. It is equipped with 28. These primary side hydraulic cylinder 26 and secondary side hydraulic cylinder 28 are formed to have the same pressure receiving area, and the primary side variable pulley 20 and the secondary side variable pulley 2
Belt type continuously variable transmission 14
is becoming smaller. The primary variable pulley 20 and the secondary variable pulley 22 are connected to the fixed rotating bodies 31 and 32 fixed to the primary rotating shaft 16 and the secondary rotating shaft 18, respectively, and the primary rotating shaft 16 and the secondary rotating shaft 18, respectively. The movable rotary bodies 34 and 36 are provided on the next rotating shaft 18 so as to be non-rotatably movable in the axial direction and to form a V-groove with the fixed rotary bodies 31 and 32, respectively.

Secondary rotating shaft 1 of the belt type continuously variable transmission 14
The output from 8 is transmitted to the drive wheels of the vehicle via an auxiliary transmission, a differential gear device, etc. (not shown).

A hydraulic control circuit for operating the vehicle power transmission device configured as described above is configured as described below. That is, the hydraulic oil that has returned to the oil tank 38 via a return path (not shown) is sucked into the oil pump 42 via the strainer 40 and the suction oil path 41, and is connected to the input port 46 of the speed change control valve 44 and the first pressure regulating valve 48. The oil is fed under pressure to the first line oil passage 50. This oil pump 42
constitutes the hydraulic power source of this embodiment, and is driven by the engine 10 via a drive shaft (not shown).
The first pressure regulating valve 48 adjusts the first line oil pressure Pl 1 by causing a part of the hydraulic oil in the first line oil passage 50 to flow out to the second line oil passage 52 in accordance with a first drive signal VD ₁ to be described later. Control. This second line oil passage 52 is connected to the first discharge port 5 of the speed change control valve 44.
4, a second discharge port 56, and a second pressure regulating valve 58, respectively. This second pressure regulating valve 58
is the second drive signal VD2, which will be described later.
A part of the hydraulic oil in the line oil passage 52 is drained into the drain oil passage 6.
By causing the oil pressure to flow to 0, the second line oil pressure Pl ₂ is controlled to be relatively lower than the first line oil pressure Pl ₁ .
The first pressure regulating valve 48 and the second pressure regulating valve 58 are constructed from so-called electromagnetic proportional relief valves. Furthermore, in this embodiment, a control hydraulic pressure source consisting of a pressure reducing valve 110 and a throttle 112, which function as the third pressure regulating valve of this embodiment, is provided in series between the second line oil passage 52 and the drain oil passage 60. It is being As a result, the pressure reducing valve 110 reduces the second line oil pressure Pl ₂
A constant control oil pressure P _c generated by reducing the pressure is applied to the pressure reducing valve 11
0 and the throttle 112 through a control oil passage 114.

The speed change control valve 44 is a so-called proportional control solenoid valve, and is connected to the input port 46, the first discharge port 54, the second discharge port 56, the primary hydraulic cylinder 26, and the secondary hydraulic cylinder 28. The valve body 6 communicates with a pair of first output ports 62 and second output ports 64, which are connected via passages 29 and 30, respectively.
5, a spool valve element 68 that is slidably fitted into the cylinder bore 66, and a spool valve element 68 that is biased toward a neutral position from both ends of the spool valve element 68. a pair of first springs 70 and second springs 72 that hold the spool valve element 68 in a neutral position;
A pair of first control oil chambers 116 and a second control oil chamber are provided oil-tightly at both ends of the spool valve 68, respectively.
A control oil chamber 118 is provided. The spool valve 68 has four lands 78, 80, 82, 84.
A pair of lands 80 and 82 located in the middle are formed sequentially from one end to the other, and a pair of lands 80 and 82 located in the middle are formed so that when the spool valve 68 is in the neutral position, the spool valve 68
8 at the same position as the first output port 62 and the second output port 64. Further, the inner circumferential surface of the cylinder bore 66,
When the spool valve 68 is in the neutral position, the position facing the pair of lands 80 and 82, that is, the first output port 62 and the second output port 64
A pair of first annular grooves 86 and second annular grooves 88 having a width slightly larger than the lands 80 and 82 are formed at positions where the first annular grooves 86 and the second annular grooves 88 open into the inner circumferential surface of the cylinder bore 66 . The first annular groove 86 and the second annular groove 88 form a constriction whose flow cross-sectional area changes continuously in order to control the flow of hydraulic oil between the lands 80 and 82.

On the other hand, the control oil passage 114 is connected to a first electromagnetic on-off valve 74 and a second electromagnetic on-off valve via throttles 120 and 122.
An electromagnetic on-off valve 76 is provided respectively. This first electromagnetic on-off valve 74 and second electromagnetic on-off valve 76
The opening operation discharges the control hydraulic pressure in the downstream side of the throttles 120 and 122, that is, the first control oil chamber 116 and the second control oil chamber 118, respectively, to the drain, and the closing operation As a result, a control oil pressure P _c is applied to the first control oil chamber 116 and the second control oil chamber 118, respectively. As a result, the spool valve element 68 is moved from its neutral position to the first spring 70 or the second spring 72.
Move it against the urging force of. The first electromagnetic on-off valve 74 and the second electromagnetic on-off valve 76 are driven on and off, that is, duty-driven, in accordance with speed ratio signals RA1 and RA2, which will be described later, so that the first electromagnetic on-off valve 74 and the second electromagnetic on-off valve 76 have a 1 control oil chamber 11
The pressure within the control oil chamber 6 or the second control oil chamber 118 is continuously changed in relation to the duty ratio, and the position of the spool valve element 68 is accordingly changed continuously.

When the spool valve 68 is in the neutral position,
The first output port 62 and the second output port 6
4 is the input port 46 and the discharge port 54,
56 with a small circulation area, an amount of hydraulic oil sufficient to replenish leakage is supplied to the primary side hydraulic cylinder 26 and the secondary side hydraulic cylinder 28, and a small amount of hydraulic oil is discharged. port 54,
56.

However, as the spool valve element 68 is moved from the neutral position in one axial direction, for example, in a direction approaching the second spring 72 (i.e., rightward in the figure), the first output port 62 and the first discharge port 54 are moved. Since the flow cross-sectional area between the second output port 64 and the input port 46 is continuously increased, the flow is output from the first output port 62 to the primary hydraulic cylinder 26. The hydraulic pressure is lower than the hydraulic pressure output from the second output port 64 to the secondary hydraulic cylinder 28 . For this reason, the balance between the thrusts of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is disrupted, so that while hydraulic oil flows into the secondary hydraulic cylinder 28, the primary hydraulic cylinder 26 leaks out, and the speed ratio e of the belt type continuously variable transmission 14 (secondary side rotating shaft 1
8 rotational speed N _put / rotational speed of the primary rotating shaft 16
N _io ) becomes smaller.

Conversely, the spool valve 68 moves from the neutral position to the first position.
As the spring 70 is moved closer to the spring 70, that is, to the left in the figure, the flow cross-sectional area between the first output port 62 and the input port 46 increases continuously, while the flow cross-sectional area between the second output port 64 and the input port 46 increases continuously. 2
Since the flow cross-sectional area with the discharge port 56 is increased, the working pressure output from the first output port 62 to the primary hydraulic cylinder 26 is reduced to the second output port 6.
4 to the secondary side hydraulic cylinder 28. For this reason, the balance between the thrust forces of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is disrupted, so that while the hydraulic oil in the secondary hydraulic cylinder 28 flows out, the primary hydraulic cylinder Hydraulic oil flows into 26, and the speed ratio e of the belt type continuously variable transmission 14 increases. In this way, the speed change control valve 44
It has a switching valve function that supplies high-pressure hydraulic oil to one of the hydraulic cylinders 26 and 28 and low-pressure hydraulic oil to the other, and a flow control valve function that continuously adjusts the flow rate of the hydraulic oil. It is.

The belt-type continuously variable transmission 14 of the vehicle includes a first rotation sensor 90 for detecting the rotation speed _Nio of the primary rotation shaft 16, and a first rotation sensor 90 for detecting the rotation speed N _put of the secondary rotation shaft 18. A second rotation sensor 92 is provided, and the first rotation sensor 90 and the second rotation sensor 92 output a rotation signal SR1 representing the rotation speed N _io and a rotation signal representing the rotation speed N _put .
SR2 is output to the controller 94. Also,
The engine 10 is provided with a throttle sensor 96 for detecting the throttle valve opening θ _th provided in the intake pipe, and an engine rotation sensor 98 for detecting the engine rotation speed _Ne . The throttle sensor 96 and engine rotation sensor 98 output a throttle signal Sθ representing the throttle valve opening θ _th and an engine rotation speed N _e
A rotation signal SE representing this is output to the controller 94.

The controller 94 includes a CPU 102, a ROM
It is a so-called microcomputer including 104, RAM 106, etc., and constitutes the control means of this embodiment. The CPU 102 processes input signals according to a program stored in advance in the ROM 104 while utilizing the memory function of the RAM 106, and controls the first pressure regulating valve 48 and the first pressure regulating valve 48 to control the first line oil pressure and the second line oil pressure. A speed ratio for driving the first electromagnetic on-off valve 74 and the second electromagnetic on-off valve 76 to control 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. Signals RA1 and RA2 are supplied to them.

The operation of this embodiment will be explained below with reference to the flowchart of FIG.

First, by executing step S1,
The rotation speed N _io of the primary rotation shaft 16, the rotation speed N _put of the secondary rotation shaft 18, the throttle valve opening θ _th , and the engine rotation speed _Ne are the rotation signals SR1 and SR2,
RAM based on throttle signal Sθ and rotation signal SE
106. Next, in step S2, the speed ratio e is calculated from the rotational speeds _Nio and _Nput according to the following equation (1) previously stored in the ROM 104.

e= _Nput / _Nio ...(1) Furthermore, in step S3, the target rotational speed _Nio ^* is determined based on the throttle valve opening _θth , etc. from the relationship stored in the ROM 104, and the above equation (1) is determined. The target speed ratio e ^* is calculated from the target rotational speed _Nio ^* and the actual rotational speed _Nput . Above target rotational speed N _io ^*
The relationship for determining is shown in FIG. 4, for example, and is determined in advance so that the engine 10 operates exclusively on the minimum fuel efficiency curve shown in FIG. In the following step S4, the ROM is
A speed ratio control value V ₀ is calculated according to the following equation (2) stored in 104. In step S14, which will be described later, if the speed ratio control value _V0 is positive, the spool valve 68 is moved to the left and the rotational speed _Nput of the secondary rotating shaft 18 is increased. A speed ratio signal RA2 is output, and if it is negative, the spool valve 68 is moved to the right and the speed ratio signal RA1 is output so that the rotation speed _Nio of the primary rotating shaft 16 increases. . The magnitude of the speed ratio control value V ₀ corresponds to the duty ratio of the speed ratio signal RA1 or RA2, that is, the amount of movement of the spool valve element 68. Therefore, the following equation
As is clear from (2), the speed ratio control value V ₀ is determined so as to match the actual speed ratio e and the target speed ratio e ^* . Note that k in equation (2) is a control constant.

V ₀ =k(e ^* -e)/e (2) Then, in step S5, the engine is adjusted based on the throttle valve opening θ _th and the engine rotational speed N _e from a well-known relationship stored in the ROM 104 in advance. The actual output torque T _e of the engine 10 is determined, and in step S6 it is determined whether the actual output torque T _e of the engine 10 is positive or not, that is, whether it is in a positive torque state where power is being output from the engine 10 or whether the engine 10 is in a positive torque state. It is determined whether the brake is on or not. The reason why such a determination is necessary is that the power transmission direction is different between the positive torque state and the engine brake state, so the oil pressure change characteristic with respect to the speed ratio e of the hydraulic cylinder changes.
For example, FIGS. 6 and 7 show the oil pressure change characteristics of the oil pressure P _io in the primary hydraulic cylinder 26 and the oil pressure P _put in the secondary hydraulic cylinder 28 in a positive torque state and an engine brake state, respectively, The magnitude relationship between the hydraulic pressure P _io and the hydraulic pressure P _put is opposite, and in both cases, the hydraulic pressure on the driving side is larger than the hydraulic pressure on the driven side. This phenomenon originally occurred in the primary hydraulic cylinder 26 and secondary hydraulic cylinder 28.
However, in this embodiment, the pressure receiving areas of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 are the same, so this is directly reflected in the magnitude relationship of the hydraulic pressure.

If it is determined in step S6 that the output torque T _e is positive, step S7 is executed to increase the pressure inside the secondary hydraulic cylinder 28 to generate the necessary and sufficient clamping force on the transmission belt 24. The second line oil pressure control value V ₂ is determined so that the oil pressure (target oil pressure) P _put ' can be obtained.
That is, first, the optimum thrust (calculated value) of the secondary hydraulic cylinder 28 is determined based on the actual output torque T _e of the engine 10 and the actual speed ratio e from the relationship of the following equation (3) stored in the ROM 104 in advance. Calculate W _put ′. In addition, from the following equation (4), the hydraulic pressure (calculated value) is calculated based on the thrust force W _put ', the pressure receiving area A _put of the secondary hydraulic cylinder 28, and the rotational speed N _put of the secondary rotating shaft 18.
In addition to calculating P _put ', the actual speed ratio e, the target speed ratio e ^* , and the actual output torque T e of the engine 10 are calculated from the relationship of the following equation (5) stored in advance in the ROM ₁₀₄ .
Calculate the corrected oil pressure ΔP ₂ based on. Then, the second line oil pressure Pl 2 is calculated from the following equation (6) based on the above oil pressure P _put ' and the corrected oil pressure ΔP ₂ , and the second line oil pressure Pl ₂ is calculated from equation (7) so that the calculated oil pressure Pl ₂ is obtained. Determine the line pressure control value _V2 .

W _put ′=f(T _e , e) …(3) P _put ′=W _put ′／A _put −C ₂ N _put ² …(4) ΔP ₂ = f(e, e ^* , T _e ) …( 5) Pl ₂ = P _put ′−ΔP ₂ …(6) V ₂ = f(Pl ₂ ) …(7) Here, the above equation (3) expresses the tension of the transmission belt 24, that is, the clamping force on the transmission belt 24. It is determined in advance to make it a necessary and sufficient value, and the thrust force W _put ' is increased proportionally with the quotient of the output torque T _e and the speed ratio e. In addition, in the relationship of equation (4), the second term is calculated by subtracting the centrifugal oil pressure, which increases with the rotational speed N _put , from the first term.
This is to correct P _put ′. Section 2 C ₂
is the centrifugal force correction coefficient, and the secondary hydraulic cylinder 2
8 and the specific gravity of the hydraulic oil.

Further, the above equation (5) is obtained in advance to calculate the corrected oil pressure ΔP ₂ . A and b in FIG. 8 show changes in the primary hydraulic cylinder 26 internal hydraulic pressure P _io and the secondary hydraulic cylinder 28 internal hydraulic pressure P _{put in} the speed change control valve 44 with respect to the speed ratio control value V ₀ (position of the spool valve 68). The characteristics are shown for different line hydraulic conditions. Assuming that the thrust is balanced at ΔV ₀ and the speed ratio at this time is e (= e ^* − Δe), the secondary hydraulic cylinder 28 at this time The internal oil pressure P _put has a value larger than the second line oil pressure Pl ₂ by ΔP ₂ . Therefore, the second line oil pressure Pl ₂ to be controlled can be obtained by subtracting the corrected oil pressure ΔP ₂ calculated using equation (5) from the oil pressure P _put ′ calculated using equation (4). This corrected oil pressure ΔP ₂ is calculated by the speed change control valve 44.
is determined by the output oil pressure change characteristic, the speed ratio control value V ₀ , and the line oil pressure difference (Pl ₁ - Pl ₂ ), but the speed ratio control value V ₀ is determined based on (e ^* - e) and the line oil pressure difference Since the difference (Pl ₁ - Pl ₂ ) is determined based on the output torque T _e and the speed ratio e, the corrected oil pressure ΔP ₂ is determined based on the speed ratio e, the target speed ratio e ^* , and the output torque T _e
The above equation (4) can be calculated in advance. Note that depending on the oil pressure change characteristics of the speed change control valve 44, the corrected oil pressure ΔP ₂ may be a small value over the entire range, but in such a case, the corrected oil pressure ΔP ₂ may be set to a predetermined constant value. .

In the above equation (7), the second line oil pressure Pl ₂ is calculated by using a data map stored in advance in consideration of the characteristics of the second pressure regulating valve 58 so that the calculated second line oil pressure Pl 2 can be obtained. Control value V ₂
is required.

In the following step S8, the first line pressure is controlled so as to obtain the oil pressure (target oil pressure) _Pio ' in the primary side hydraulic cylinder 26 to generate the necessary and sufficient thrust to achieve the target speed ratio. value
V ₁ is determined. That is, first, ROM1
The target speed ratio e ^* and the actual output torque T _e of the engine 10 are determined from the relationship shown in the following equation (8) stored in 04.
The thrust ratio γ ₊ (thrust force W _put of the secondary hydraulic cylinder 28 / thrust force W _io of the primary hydraulic cylinder 26 ) during forward drive is calculated based on the equation (9), and the above thrust ratio γ ₊ and the thrust force of the primary hydraulic cylinder 26 from the thrust force W _put ′ of the secondary hydraulic cylinder 28
W _io ′ is calculated. Then, from the following equation (10), the hydraulic pressure (calculated value) P io is calculated based on the thrust force W _io ' of the primary hydraulic cylinder 26, the pressure receiving area A _io of the primary hydraulic cylinder 26, and the rotational speed N _io of the primary rotating shaft 16 _. ′ is calculated, and the primary line oil pressure Pl ₁ is also calculated from the following equation (11) based on the above-mentioned oil pressure P _io ′ and the corrected oil pressure ΔP ₁ , and so that the calculated oil pressure Pl ₁ is obtained ( Determine the first line pressure control value V ₁ from equation 12).

γ ₊ = f (e ^* , T _e ) …(8) W _io ′=W _put ′／γ ₊ …(9) P _io ′=W _io ′／A _io −C ₁ N _io ² …(10) Pl ₁ = P _io ' + ΔP ₁ ... (11) V ₁ = f (Pl ₁ ) ... (12) Here, the above equation (8) is sufficient to obtain suitable shift response over a wide range of operating conditions. This indicates a relationship determined in advance so that the necessary and sufficient thrust ratio γ ₊ can be determined, and from this relationship, the thrust ratio γ ₊ determined in relation to the target speed ratio e ^* and the actual output torque T _e . The first line oil pressure is controlled so that the following is obtained. Furthermore, in the relationship of equation (10) above, the second term is corrected by subtracting the centrifugal oil pressure that increases with the rotational speed N _io from the first term, and the second term C ₁ is corrected by subtracting the centrifugal oil pressure that increases with the rotation speed N io.
It is determined in advance from the specifications of and the specific gravity of the hydraulic oil.
Furthermore, the above equation (11) is based on the hydraulic pressure determined by equation (10).
The first line oil pressure Pl ₁ is determined by adding the correction oil pressure ΔP ₁ to P _io ′, but this correction oil pressure ΔP ₁ is calculated by adding power loss and steady-state deviation Δe (ΔV ₀
) is determined at the equilibrium point. That is, Fig. 8 a and b show cases where the above-mentioned corrected oil pressure ΔP ₁ is changed. In case a, where the corrected oil pressure ΔP ₁ is small, the steady-state deviation becomes large, but in case b, where the corrected oil pressure ΔP ₁ is increased, In this case, the actual internal hydraulic pressure P _io of the primary side hydraulic cylinder 26 and the actual internal hydraulic pressure P _put of the secondary side hydraulic cylinder 28 change rapidly, so that the steady-state deviation becomes small. However, as the correction oil pressure ΔP ₁ is increased, an unnecessarily large first line oil pressure Pl ₁ is generated under more operating conditions.

In equation (12), a data map or the like stored in advance in consideration of the characteristics of the first pressure regulating valve 48 is used so that the calculated first line oil pressure Pl ₁ can be obtained. The control value V ₁ is determined.

On the other hand, if it is determined in step S6 that the vehicle is in the engine braking state, the direction of power transmission in the belt type continuously variable transmission 14 is reversed, so steps S9 and S8, which are substantially the same as steps S7 and S8, are performed. By executing S10, the required oil pressure in the primary side hydraulic cylinder 26 is
The second line hydraulic pressure control value V ₂ is determined from P _io ', and the first line hydraulic pressure control value V ₁ is determined from the hydraulic pressure P _put ' required in the secondary side hydraulic cylinder 28. That is,
In step S9, the optimum thrust of the primary hydraulic cylinder 26 is determined based on the output torque T _e and the speed ratio e from the pre-stored relationship shown in the following equation (13).
W _io ′ is calculated, and the hydraulic pressure P _io ′ to be supplied to the primary side hydraulic cylinder 26 is calculated from the following equation (14), while the corrected oil pressure ΔP ₂ is calculated from the above equation (5). The second line oil pressure Pl ₂ is calculated from equation (15) based on the oil pressure P _io ' and the corrected oil pressure ΔP ₂ , and the second line oil pressure control value V ₂ is determined from equation (7). In addition, in step S10, the thrust force γ _- is calculated from the following equation (16) based on the target speed ratio e ^* and the output torque T _e , and
The thrust force W _put ′ of the secondary hydraulic cylinder 28 to obtain the above-mentioned thrust ratio γ ₋ is determined from the following equation (17) based on the thrust ratio γ ₋ and the thrust force W _io ′ of the primary hydraulic cylinder 26, and (18 ), calculate the hydraulic pressure P _put ′ in the secondary hydraulic cylinder 28, and then use the following equation (15).
The first line oil pressure Pl ₁ is determined based on the above oil pressure P _put and the corrected oil pressure ΔP ₁ , and the first line oil pressure control value V ₁ for obtaining the first line oil pressure Pl ₁ is determined from the above equation (19). do. Here, as described above, the second line oil pressure control value V ₂ and the first line oil pressure control value V ₁ are each a function of the output torque T _e of the engine 10, so the second line oil pressure Pl ₂
It can be said that the first line oil pressure Pl ₁ is regulated based on the output state of the engine.

W _io ′=f(T _e , e) …(13) P _io ′=W _io ′/A _io −C ₁ N _io ² …(14) Pl ₂ =P _io ′−ΔP ₂ …(15) γ ₋ = f (e ^* , T _e ) …(16) W _put ′=W _io ′・γ _- …(17) P _put ′=W _put ′／A _put −C ₂ N _put ² …(18) Pl ₁ = P _put ′+ΔP ₁ …(19) V ₁ =f(Pl ₁ ) …(20) In this way, the second line hydraulic pressure control value V ₂ ,
Once the first line hydraulic pressure control value _V1 is determined, the next step S11 is executed to determine whether the deviation Δe between the target speed ratio e ^* and the actual speed ratio e is positive. If so, in step S12 the following equation (21)
From (22), the above first line hydraulic pressure control value V ₁
and the second line oil pressure control value _V2 is corrected.
If it is negative, the first line hydraulic pressure control value V ₁ and the second line hydraulic pressure control value V ₂ are corrected from the following equations (23) and (24) in step S13.

V ₁ =V ₁ +k ₁ (e ^* -e)/e...(21) V ₂ =V ₂ -k ₂ (e ^* -e)/e...(22) V ₁ =V ₁ +k ₃ (e-e ^* )/e...(23) _V2 = _V2 - _k4 (e-e ^* )/e...(24) However, _k1 , _k2 , _k3 , _{and k4} are each proportionality constants.

As is clear from the above equation, steps S12 and
S13 is the first line oil pressure as the deviation |Δe| increases.
This is to increase the speed ratio change speed of the belt type continuously variable transmission 14 by increasing the difference between Pl ₁ and the second line oil pressure Pl ₂ . That is, for example, in a positive torque state, the first line oil pressure Pl ₁ is a correction oil pressure (margin oil pressure) with respect to the oil pressure P _io in the primary side hydraulic cylinder 26 (hydraulic pressure in the high pressure side hydraulic cylinder: P _put in the engine brake state). Although ΔP is increased by ₁ minute, it cannot be increased too much due to power loss and may not be sufficient in terms of speed ratio change speed. However, in this embodiment, the speed ratio change speed can be further increased by increasing the difference between Pl ₁ and Pl ₂ in a transient state where the deviation |Δe| becomes large, so that extremely suitable speed change response can be obtained. .

Here, in the above equations (21), (22), (23), and (24), the proportional constants are used to change the speed change response, and generally speaking, deceleration speed is less than speed change. The faster the speed, the better the driving sensation, so k ₁ < k ₃ , k ₂ <
It has been determined that k ₄ . Figure 9 shows the hydraulic pressure values when the speed ratio changes in a positive torque state (P _io > P _put ) when the above equations (21), (22), (23), and (24) are applied to control. It shows temporal change characteristics. As is clear from the figure, the speed change control valve 4
Due to the operation of the spool valve 68 of No. 4, the primary side hydraulic cylinder 2 is
At the same time, the hydraulic pressure P _io in the primary hydraulic cylinder 26 is increased and the hydraulic pressure P _put in the secondary hydraulic cylinder 28 is lowered, while the hydraulic pressure P _io in the primary hydraulic cylinder 26 is transiently lower during deceleration shifting (speed ratio decrease). At the same time, the hydraulic pressure P _put in the secondary hydraulic cylinder 28 is increased. As a result, a large thrust force difference is generated in both hydraulic cylinders 26 and 28 in a transient state, so that suitable speed change responsiveness in speed ratio control can be obtained.

In the last step S14 of the series of steps, the speed ratio control value V ₀ , the first line hydraulic pressure control value V ₁ , and the second line hydraulic pressure control value V ₂ determined in the previous steps are output. This results in
As shown in FIGS. 5, 6, 7, and 8, the speed ratio e, the first line oil pressure Pl ₁ , and the second line oil pressure Pl ₂ are controlled.

In this way, according to this embodiment, the first pressure regulating valve 4
8 and the second pressure regulating valve 58 to control the first line oil pressure.
Since Pl ₁ and second line hydraulic pressure Pl ₂ are prepared, the pressure difference between them causes the hydraulic oil to be supplied to one of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28, or the hydraulic oil discharged from it. The oil flow rate is determined. Therefore, regardless of the actual speed ratio and transmission torque (output torque T _e ) of the belt type continuously variable transmission 14, the speed ratio change speed depends on the differential pressure between the first line oil pressure Pl ₁ and the second line oil pressure Pl ₂ . Therefore, sufficient transient response characteristics for speed ratio control can be obtained.

Furthermore, by controlling the first pressure regulating valve 48 in relation to the output torque T _e of the engine 10 and the actual speed ratio e, the first line oil pressure Pl ₁ can be adjusted to a sufficient speed ratio change speed and reduce power loss. It is controlled to a necessary and sufficient value so that it does not occur, and
By controlling the second pressure regulating valve in relation to the speed ratio and transmission torque, the second line oil pressure Pl ₂ is controlled to a necessary and sufficient value within a range that does not cause slippage of the transmission belt, reducing vehicle power loss. This has the advantage that it is significantly reduced.

Further, according to this embodiment, the first pressure regulating valve 48 and the second pressure regulating valve 58 are connected, so that the pressure of the hydraulic oil flowing out from the first pressure regulating valve 48 is regulated by the second pressure regulating valve 58. This has the advantage that the amount of oil necessary for regulating the pressure of the second pressure regulating valve 58 can be saved, the discharge capacity of the oil pump 42 can be reduced, and power loss can be further reduced.

Further, according to this embodiment, when a constant control oil pressure P _c is applied to the spool valve element 68 by both the electromagnetic on-off valves 74 or 76, the spool valve element 68 is moved, and the output of the speed change control valve 44 is A switching operation is performed. For this purpose, a spool valve is formed by a link whose one end is connected to an operating member for changing the speed ratio, and whose other end is engaged with a movable rotating body that constitutes a part of a variable pulley and moves in the axial direction. Compared to a type of speed change control valve in which the speed change control valve is driven, the degree of freedom in control is greatly increased because the speed change pattern is not limited to the speed change pattern determined by the specifications of the link mechanism. Further, since there is no need to mechanically connect the speed change control valve to the variable pulley via a link mechanism, there is an advantage that the arrangement of both can be made freely. Furthermore, there is an advantage that a decrease in control accuracy caused by link rattling that is unavoidable in a link mechanism is eliminated.

Moreover, when the first line hydraulic pressure Pl ₁ or the second line hydraulic pressure Pl ₂ is applied to the spool valve element by the electromagnetic on-off valves 74 and 76, the first line hydraulic pressure Pl ₁ and the second line hydraulic pressure Pl ₂ are applied to the belt. Although it is unavoidable that the speed change responsiveness varies due to changes related to the speed ratio e and transmission torque of the continuously variable transmission 14, according to the hydraulic control device of this embodiment, the pressure reducing valve The control oil pressure P _c controlled constant by 110 is applied to both electromagnetic on-off valves 74 and 76
Since it is made to act on the spool valve 68 by
This has the advantage of providing stable shift response.

Further, according to this embodiment, the spool valve 6
Compared to electromagnetic proportional control valves 74 and 76, which are driven directly by proportional electromagnetic solenoids, it is possible to use electromagnetic on-off valves 74 and 76, which are simpler in structure and smaller in size due to their on-off operation. There is an advantage that it does not have to be disposed integrally with the control valve 44. In addition, in the speed change control valve 44, the spool valve element 68 is moved based on pressure-pressure equilibrium, so the spool valve element 68 can be driven with a large driving force, and the influence of foreign substances in the hydraulic oil can be significantly reduced. It has the advantage of increasing contamination.
Furthermore, the output of the controller 94 made up of a computer can be supplied to the electromagnetic on-off valve without going through an A/D converter, and a constant current amplifier to eliminate the influence of temperature changes in the winding resistance of the linear solenoid is no longer required. There are advantages.

Next, another embodiment of the present invention will be described. In addition,
In the following description, parts having the same functions are given the same reference numerals and the description thereof will be omitted.

In FIG. 10, a pressure reducing valve 110 and a throttle 112 are provided in series between the first line oil passage 50 and the drain oil passage 60. In this way, the first line oil pressure Pl ₁ is reduced to a constant control oil pressure P _c by the pressure reducing valve 110, and is led out through the control oil path 114.

In FIG. 11, a pressure control valve 130 is provided in the drain oil passage 60 for regulating the pressure of the hydraulic oil flowing out from the second pressure regulating valve 58 to a constant control oil pressure P _c .
This pressure control valve 130 corresponds to the third pressure regulation valve of this embodiment, and thereby the control oil pressure P _c is led out from between the second pressure regulation valve 58 and the pressure control valve 130 through the control oil passage 114. Ru.

In FIG. 12, the first pressure regulating valve 48 that regulates the first line hydraulic pressure Pl ₁ and the second line hydraulic pressure
A second pressure regulating valve 58 that regulates the pressure of Pl ₂ is provided in parallel, and a pressure control valve 130 that regulates the pressure of the hydraulic oil flowing out from the first pressure regulating valve 48 to a constant control pressure P _c is provided in the drain oil path 60. ing. As a result, the control oil pressure P _c is led out from between the first pressure regulating valve 48 and the pressure control valve 130 through the control oil passage 114 .

Note that the above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of drawings]

FIG. 1 is a diagram showing the configuration of an embodiment of the present invention. FIG. 2 is a diagram showing the relationship between the duty ratio of the electromagnetic on-off valve in FIG. 1 and the oil pressure in the control oil chamber of the speed change control valve. FIG. 3 is a flowchart for explaining the operation of the embodiment shown in FIG. FIG. 4 is a diagram showing relationships used to explain the operation of the flowchart of FIG. 3. FIG. 5 is a diagram showing a minimum fuel consumption rate curve of the engine of FIG. 1. 6 and 7 are diagrams showing the change characteristics of the oil pressure of each part with respect to the speed ratio in the embodiment shown in FIG. 1, respectively. FIG. 6 shows the positive torque state, and FIG. 7 shows the engine braking state. There is. FIG. 8 is a diagram showing the output characteristics of the speed change control valve shown in FIG.
b indicates a state where the differential pressure between the line hydraulic pressure and the second line hydraulic pressure is small, and b indicates a state where the differential pressure between the first line hydraulic pressure and the second line hydraulic pressure is large. FIG. 9 is a diagram showing oil pressure change characteristics of various parts in a transient state of the embodiment shown in FIG. FIG. 10, FIG. 11, and FIG. 12 are diagrams showing main parts of other embodiments of the present invention, respectively. 14: Belt type continuously variable transmission, 16: Primary side rotating shaft, 18: Secondary side rotating shaft, 20: Primary side variable pulley, 22: Secondary side variable pulley, 24: Transmission belt, 26: Primary side hydraulic cylinder , 28: Secondary side hydraulic cylinder, 44: Speed change control valve, 48, 130:
1st pressure regulating valve, 58, 132: 2nd pressure regulating valve, 11
0: Pressure reducing valve (third pressure regulating valve), 112: Throttle, 11
4: Control oil path, 116, 118: Control oil chamber, 13
0: Pressure control valve (third pressure regulating valve).

Claims (1)

[Scope of Claims] 1. A pair of primary variable pulleys and variable secondary pulleys provided on the primary rotating shaft and secondary rotating shaft, respectively, and a device that is wound around the pair of variable pulleys to transmit power. A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising a power transmission belt, and a pair of primary and secondary hydraulic cylinders that respectively change the effective diameters of the pair of variable pulleys, the hydraulic control device comprising: a hydraulic power source; A first pressure regulating valve that regulates the pressure of the hydraulic oil supplied from the engine based on the output state of the engine and makes it the first line oil pressure, and a spool valve element that is movable in the axial direction, and as the spool valve element moves. By supplying the hydraulic oil whose pressure has been regulated to the first line hydraulic pressure to one of the primary hydraulic cylinder and the secondary hydraulic cylinder, and simultaneously extracting the hydraulic oil from the other, the primary variable pulley and a speed change control valve that adjusts the speed ratio of the continuously variable transmission by changing the effective diameter of a secondary variable pulley; and an operation that flows out from the other of the primary hydraulic cylinder and secondary hydraulic cylinder through the speed change control valve. a second pressure regulating valve that regulates the oil pressure based on the output state of the engine and sets a second line hydraulic pressure lower than the first line hydraulic pressure; and generating a constant control hydraulic pressure for driving the spool valve element. A belt-type continuously variable transmission for a vehicle, comprising: a third pressure regulating valve; and an electromagnetic control valve device that controls the movement position of the spool valve by applying the control hydraulic pressure to the spool valve. Hydraulic control device. 2. The speed change control valve is provided at a cylinder bore into which the spool valve element is slidably fitted, and at both ends of the cylinder bore, and for applying control hydraulic pressure to pressure receiving surfaces at both ends of the spool valve element. The electromagnetic control valve device includes a pair of oil-tight first control oil chambers and a second control oil chamber, and the electromagnetic control valve device includes:
The vehicle belt-type valve according to claim 1, which is a pair of electromagnetic on-off valves that are duty-controlled to supply the control oil pressure to the pair of first control oil chamber and second control oil chamber, respectively. Hydraulic control device for gear transmission. 3 The third pressure regulating valve is a pressure reducing valve and a throttle connected in series between a first line oil passage for guiding the first line oil pressure or a second line oil passage for guiding the second line oil pressure and a drain oil passage. The belt-type continuously variable transmission for a vehicle according to claim 1 or 2, wherein the control hydraulic pressure is generated between the pressure reducing valve and the throttle. Hydraulic control device. 4. The third pressure regulating valve is provided in an oil passage that guides the hydraulic oil flowing out from the first pressure regulating valve or the second pressure regulating valve, and regulates the pressure of the hydraulic oil to a constant control oil pressure. A hydraulic control device for a vehicle belt-type continuously variable transmission according to item 1 or 2.