JPS63176750A - Hydraulic controller for v-belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To prevent slip between a belt and a pulley and to improve the durability by providing a means for boosting the line pressure of a first servo line for producing a gripping force when an engine is braked. CONSTITUTION:A hydraulic regulator 60 regulates the output line pressure of a regulator valve 61 according to variation of the shift position of a manual valve 62, an opening of a throttle valve 65 and a ratio of torque between input and output pulleys. Under L-range, oil paths 3, 5 are communicated with an oil path 1 in the manual valve 62. When the torque ratio is within a specific range and the throttle valve 65 has a specific % of opening theta, oil paths 5, 7 are communicated in a detent valve 64 to produce detent pressure which pushes up a throttle plunger and produces a high line pressure. Consequently, a high line pressure is fed to a first hydraulic servo when an engine is braked, and slip between pulley and belt is prevented.

Description

DETAILED DESCRIPTION OF THE INVENTION C. Industrial Application Field The present invention relates to a hydraulic control device for a V-belt continuously variable transmission for a vehicle.

[Conventional technology]

■Belt-type continuously variable transmissions can be used as automatic transmissions for automobiles and other vehicles in combination with planetary gear transmissions for forward and reverse switching. Figure 30 is from Japanese Patent Application Laid-Open No. 54-1579.
A conventional continuously variable transmission proposed in Publication No. 30 is shown. The input shaft a is provided with a fixed pulley b and a movable pulley C, and the output shaft d is provided with a fixed pulley e and a movable pulley r.
is provided, and a bell l-g is stretched between the input shaft a and the output shaft d, and the fluid from the pump j is supplied to the oil path and the oil path i to the valve. The movable pulley CSr is moved by supplying and discharging the fluid through the movable pulley CSr. On one end of the spool m of the valve l, a fluid pressure proportional to the rotation speed of the input shaft a is applied by a pitot tube n, and on the other end of the spool m, the rotation of a cam p that is linked to the movement of the throttle pedal is applied. The pressure due to the movement is acting through lever q and spring r. Furthermore, fluid pressure proportional to the rotational speed of the input shaft a is applied to one end of the spool S of the valve by a pitot tube n, and the other end of the spool S is connected to a movable pulley C of the input shaft a. The pressure of the detection rod t that is displaced by this movement acts via the lever U and the spring (2).

In the above configuration, the line pressure generated in the oil path i is caused by the fluid pressure generated by the pitot tube n depending on the rotation speed of the input shaft a and the spring V which changes due to the rotation of the lever U linked to the movement of the detection rod t. Controlled by the balance with the load, the oil pressure increases as the torque ratio increases and increases as the input shaft rotational speed decreases.

[Problem that the invention seeks to solve]

However, in the conventional continuously variable transmission mentioned above, the line pressure is set regardless of the output torque of the engine that is the drive source of this continuously variable transmission, so the line pressure is set at a relatively high line pressure that can always transmit the maximum output torque of the engine. pressure must be set. For this reason, more clamping force than necessary acts on the V-belt, impairing its durability, and generating a line pressure higher than necessary increases the driving torque of the oil pump, increasing torque loss. However, there was a problem in that the improvement in fuel efficiency was impaired. Therefore, it is conceivable to control the line pressure according to the output torque of the engine to always adjust it to the optimum oil pressure.

However, the transmission torque of the continuously variable transmission in the engine braking state is not the output torque of the engine, but the torque that drives the engine that is input from the tires due to the inertia of the vehicle, and this is considered in the above line pressure control. In this case, there is a problem that slipping occurs between the heald and the booster in the continuously variable transmission section, impairing the durability of the belt and the booster.

The present invention solves the above-mentioned problems, and is capable of improving the durability of belts and pulleys by increasing line pressure during engine braking and increasing torque transmission capacity. The purpose of the present invention is to provide a hydraulic control device for a transmission.

[Means for solving problems]

To this end, the hydraulic control device for the V-belt continuously variable transmission for vehicles of the present invention includes an input pulley and an output boolean that are respectively attached to the input shaft and the output shaft and have variable effective diameters.
A belt-type continuously variable transmission for vehicles consists of a drive band stretched between these pulleys, and the effective diameter of the pulley is hydraulically adjusted to control the torque ratio between the input and output shafts. , a pressure regulating device that generates line pressure by regulating the pressure of hydraulic oil from the hydraulic source according to the engine output torque, and a first pressure regulating device that is supplied with the line pressure and generates a clamping force for the drive band according to the line pressure. The engine is characterized by comprising a hydraulic servo and a pressure increasing means for increasing the line pressure during engine braking.

[Operation and Effects of the Invention] In the present invention, as shown in FIG. 2, for example, in the L range, the oil passage 3 and the oil passage 5 communicate with the oil passage 1 in the manual valve 62, and the torque ratio T is t1. ≦T≦1t
When the throttle opening degree θ is 050501%, the oil passage 5 and the oil passage 7 communicate with each other at the detent valve 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure. 65% 05100%
In this case, the pressure in the oil passage 7 is exhausted through the oil passage 6 and the drain boat 665 of the torque ratio valve as shown in Fig. 4 (B), and no detent pressure is generated, and the throttle pressure is in the D position. is the same as Therefore, the line pressure is as shown in Figure 11 (
The characteristics are as shown in (2).

Further, when the torque ratio T is tz <T S t 3, the oil passage 9 is communicated with the drain boat 666 in the torque ratio valve 66 to exhaust pressure, and the throttle valve 65 is connected to the oil passage 8
The throttle pressure outputted to the regulating valve 61 via the line pressure increases, and the line pressure is thereby expressed by a characteristic curve as shown in FIG. 11(h).

Further, when the torque ratio T is t and <“l”≦t4, the oil passage 6 and the oil passage 1 are brought into communication by the torque ratio valve 66, and the pressure of the oil passage 9 is discharged from the drain boat 666. Since line pressure is supplied to both the oil passage 6 and the oil passage 5, the detent valve 64 outputs detent pressure regardless of the throttle opening, and the detent pressure and the above (2)
The regulating valve 61 to which the same throttle pressure is input is shown in Fig. 11 (
Outputs the line pressure shown in ).

That is, as shown in the characteristic curves (H) and (L) of Fig. 11, the line pressure is set high under operating conditions in which the torque ratio T is in the range of tl ≦T≦t2 and the throttle opening θ is 0.1% or less. This is due to the fact that relatively high oil pressure is required during engine braking even at low throttle intervals.

Therefore, according to the present invention, by increasing the line pressure and increasing the torque transmission capacity during engine braking, it is possible to prevent slippage between the belt and the pulley and improve the durability of the belt and the pulley.

〔Example〕

Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 is a schematic diagram of an automobile transmission using a continuously variable transmission.

100 is an engine, 102 is a carburetor, and 20 is a transmission device provided between the engine 100 and the drive side axle, and is connected to a hydraulic fluid coupling 21 and a differential gear 22 connected to the output side 101 of the engine. The present invention is constructed by a continuously variable transmission device consisting of a reduction gear mechanism 23, and (1) a belt type continuously variable transmission 30 and a planetary gear shift 8!40 for forward/reverse switching.

The fluid coupling 21 is well known and consists of a pump impeller 211 and a turbine runner 212 connected to a fluid coupling output shaft 214. Note that other hydraulic torque converters or AVI type clutches may be used instead of the fluid compression ring.

■The belt type continuously variable transmission 30 has a fixed flange 311 connected to the fluid coupling output shaft 214 which is the input shaft of the continuously variable transmission 30, and forms a V-shaped space facing the fixed flange 311. an input pulley 31 consisting of a movable flange 312 provided to
A fixed flange 321 connected to the intermediate shaft 26 which is the output shaft of the motor 30, a movable flange 322 provided to face the fixed flange 321 to form a V-shaped space, and a hydraulic servo that drives the movable flange 322. This is a well-known device consisting of an output pulley 32 made of 323, and a heald 33 which is a drive band connecting the input pulley 31 and the output pulley 32.

The movable flanges 312 and 322 of the input side boob IJ31 and the output side pulley 32 are displaced from 0 to At-4.
1, ~Ila (0< 121 < j!3 < 1
4), and as a result, the torque ratio T between the input shaft 214 and the output shaft 26 is t1~t2~t,~[4(tl<
Continuously variable transmission is performed within the range of tz < t3 < ta. In this embodiment, the pressure receiving area of the input side hydraulic servo 313 is approximately twice as large as the pressure receiving area of the output side hydraulic servo 323, and the hydraulic pressure applied to the hydraulic servo 313 is equal to or equal to the hydraulic pressure applied to the hydraulic servo 323. Even if the input side movable flange 312 is small, the movable flange 312 on the input side is formed to obtain a larger driving force than the movable flange 322 on the output side. This increase in the pressure receiving area of the hydraulic servo 313 is achieved by increasing the diameter of the hydraulic servo or by employing a piston having a double pressure receiving area in the hydraulic servo.

The planetary gear shift W140 for forward/reverse switching includes a sun gear 41 connected to the intermediate shaft 26 which is the output shaft of the continuously variable transmission 30, and a ring gear 43 engaged with the transmission case 400 via the multi-disc brake 42. , a double planetary gear 44 rotatably meshed between the sun gear 41 and the ring gear 43, which rotatably supports the double planetary gear 44 and is connected to the intermediate shaft 26 via a multi-plate clutch 45; It is composed of a planetary carrier 46 connected to a second intermediate shaft 47 that is the output shaft of the planetary gear transmission 40, a hydraulic servo 49 that operates the multi-disc brake 42, and a hydraulic servo 49 that operates the multi-disc clutch 45. . This planetary gear shift 840 for forward/reverse switching provides a forward gear with a reduction ratio of 1 when the multi-disc clutch 45 is engaged and the multi-disc brake 42 is released; When the brake 42 is engaged, it becomes a reverse gear with a reduction ratio of 1.02. This reduction ratio of 1.02 in reverse is smaller than the reduction ratio in reverse of a normal automobile transmission, but in this example, Since deceleration is performed in the deceleration gear mechanism 23, which will be described later, an appropriate deceleration ratio can be obtained as a whole.

The reduction gear mechanism 23 is intended to compensate for the fact that the speed change range obtained with the $130 belt valor gear shift is lower than the speed change range achieved by a normal vehicle transmission, and the reduction gear mechanism 23 has a reduction ratio of 1 between the input and output shafts. The torque is increased by performing a .45 speed change.

The differential gear 22 is connected to an axle (not shown) and provides final deceleration of 3.727:1.

FIG. 2 shows a hydraulic control circuit for controlling the continuously variable transmission in the transmission shown in FIG.

The hydraulic control circuit controls the timing of engagement of the multi-disc brake 42 and the multi-disc clutch 45 in the hydraulic source 50, the hydraulic adjustment device 60, and the planetary gear shift 8140, and
- It consists of a shift control mechanism 70 that cushions the impact during the R shift, and a torque ratio control device 80.

The hydraulic adjustment device 60 includes a manual valve 62 that is manually operated by a shift lever (not shown), a detent valve 64 and a throttle valve 65 that output detent pressure and throttle pressure according to the throttle opening θ of the carburetor 102, and an output side. The torque ratio valve 66 works in conjunction with the movable flange 321 of the pulley 32 and supplies line pressure to the detent valve 64 according to its displacement amount, and exhausts pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65, and is supplied from the hydraulic pressure a50. It is comprised of a regulator valve 61 that regulates the generated oil pressure and supplies it to various parts of the oil pressure adjustment device 60 as line pressure.

The hydraulic power source 50 supplies hydraulic oil pumped up from an oil strainer 51 by a pump 52 driven by an engine to a regulator valve 61 through an oil passage 11 to which a relief valve 53 is attached.

The manual valve 62 has a spool 621 set at positions P, R, N, D, and L, as shown in FIG. 3, corresponding to shift positions P, RSN, D, and L of a shift lever provided at the driver's seat. As shown in Table 1, the oil passage 1 to which line pressure is supplied is connected to the output oil passages 3 to 5.

Table I In Table I, ◯ indicates a state of communication with oil passage 1, and × indicates that oil passages 3 to 5 are in a discharged pressure state.

The regulator valve 61 includes a spool 611 and a regulator valve plunger 612 that controls the spool 611 by inputting detent pressure and throttle pressure. is adjusted, and line pressure is output from the output port 16 to the oil path 1. Oil is supplied from the boat 614 through the oil passage 12 to the fluid coupling, oil cooler, and parts requiring lubrication.

The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the butterfly valve of the capretor 102 as shown in FIG.
≦θ, as shown in FIG.
When it is 100%, the oil passage 7 and the oil passage 6 which connects the detent valve 64 to the torque ratio valve 66 are communicated as shown in FIG. 4(B). Note that the spool 641 may be linked to the amount of depression of the accelerator pedal.

The throttle valve 65 includes a spool 651 which is connected in series with the spool 641 of the detent valve via a spring 645 and has a spring 652 on the other side, and the throttle opening θ transmitted via the spool 641 and the spring 645. (spool 6 above)
51, the opening area of the boat 653 communicating with the oil passage 1 is adjusted, and throttle pressure is output to the oil passage 8 for throttle pressure output communicating with the input port 18 provided in the regulator valve 61. The spools 651 are branched from the oil passage 8 and have orifices 654 and 65, respectively.
The output hydraulic pressure is fed back to a land 656 and a land 657 having a larger pressure receiving area than the land 656 through the output hydraulic pressure feedback oil passages 9 and 10 provided with the output hydraulic pressure.

The torque ratio valve 66 has a spool 6 linked to the movable flange 322 of the output pulley 32 via a connecting rod.
62, and the amount of movement of the movable flange 322 is l, ≦L.
≦i, (torque ratio T is t2≧T≧1+), the spool 662 is located on the left side of the figure as shown in FIG. The input port 64 connected to the detent valve 64 is closed, and the output oil passage 6 to the detent valve 64 is communicated with the drain boat 665 to discharge pressure. The amount of movement of the movable flange 322 is 12≦L<J2 (ts≧T>t2
), the spool 662 is located at the intermediate portion as shown in FIG. 5(B), the boat 664 connected to the oil passage 9 and the drain boat 666 communicate with each other, and the oil passage 9 is evacuated. When the movement i1L is O≦L≦1t (t4≧7>t,), the spool 662 is located on the right side of the figure as shown in FIG. 5(C), and the boat 663 connected to the oil passage 1 and the oil passage 6, and line pressure is supplied to the oil passage 6.

In addition, the spool 662 is connected to the output pulley 3 which is in a rotating state.
The spool 662 slides in conjunction with the movable flange 322 of No. 2, but as shown in FIG. , the movement of the movable flange is not hindered, and it is possible to prevent wear and the like of sliding parts having large relative speeds.

The shift control mechanism 70 includes a shift control valve 71 including a spool 712 having a spring 711 placed behind it on one end and receiving line pressure from an oil chamber 713 provided at the other end, and an oil chamber 713.
Orifice 7 provided in oil passage 1 that supplies heline pressure
It consists of a pressure limiting valve 73 installed between the 2.8 orifice 72 and the oil chamber 713, and a solenoid valve 74 that is controlled by an electric control circuit to be described later and adjusts the oil pressure in the oil chamber 713. When the solenoid valve 74 operates to open the drain boat 741 and evacuate the oil chamber 713, the spool 712 of the shift control valve 71 is moved to the left in the figure by the action of the spring 711, and the spool 712 of the planetary gear transmission 40 is The oil passage 13 that communicates with the hydraulic servo 49 that operates the plate clutch 45 and the oil passage 14 that communicates with the hydraulic servo 48 that operates the multi-disc brake 42 are connected to the drain boats 714 and 715, respectively, to discharge pressure. The plate clutch 45 or the multi-plate brake 42 is released. When the solenoid valve 74 is not operating, the drain boat 741 is closed, and the spool 712 is located on the right side in the figure by the line pressure supplied to the oil chamber 713, and connects the oil passages 3 and 4 to the oil passage 13, respectively. and the oil passage 14, and engages the multi-disc brake 42 or the multi-disc clutch 45.

In this embodiment, the shift control valve 71 is provided with an oil chamber 717 and an oil chamber 716 that feed back the output oil pressure of the oil passage 13 and the oil passage 14, and this reduces the rise of the output oil pressure and controls the multi-disc clutch 45 and the multi-disc brake 42. This prevents shock when mating.

The torque ratio control device 80 includes a torque ratio control valve 81, orifices 82 and 83, a downshift solenoid 84,
and an upshift solenoid 85. The torque ratio control valve 81 has a spool 812 with a spring 811 installed on one side, and oil chambers 81 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, respectively.
5 and 816, an input boat 817 that communicates with the oil passage 1 to which line pressure is supplied and whose opening area increases or decreases according to the movement of the spool 812, and ■ a belt-type continuously variable transmission 3;
an oil chamber 819 provided with an output boat 818 that communicates with the hydraulic servo 313 of the input pulley 31 of 0 through the oil passage 2;
, a drain boat 814 that evacuates the oil chamber 819 according to the movement of the spool 812, and a drain boat 813 that evacuates the oil chamber 815 according to the movement of the spool 812.

The downshift solenoid 84 and the upshift solenoid 85 are respectively attached to an oil chamber 815 and an oil chamber 816 of the torque ratio control valve 81, and both are operated by the output of an electric control circuit to be described later.
and the oil chamber 816 are evacuated.

FIG. 6 shows the solenoid valve 74 of the shift control mechanism 70 and the torque ratio control device 80 in the hydraulic control circuit shown in FIG.
The configuration of an electric control circuit 90 that controls the downshift solenoid 84 and the amplifier shift solenoid 85 is shown.

901 is a shift lever switch that detects whether the shift lever is shifted to P, R, N, D, or L; 902 is a rotation speed sensor that detects the rotation speed of the input end pulley 31; 903 is a vehicle speed sensor; 904 is a throttle sensor that detects the throttle opening of the carburetor or the amount of depression of the accelerator pedal; 905 is a rotation speed sensor 902
906 Speed detection processing circuit that converts the output of
is a vehicle speed detection circuit that converts the output of the vehicle speed sensor 903 into voltage, 907 is a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage, 908 to 911 are input interfaces of each sensor, and 912 is a central processing unit ( CPU), 913 is solenoid valve 74.84.8
A read-on memory (ROM) 914 stores a program to control the control program and data necessary for control; 914 a random access memory (RAM) 915 temporarily stores input data and parameters necessary for control;
is a clock, 916 is an output interface, 917 is a solenoid output driver, and output interface 9
16 into operating outputs of upshift solenoid 85, downshift solenoid 84, and shift control solenoid 74.

Input interfaces 908 to 911 and CPU 912,
ROM913, RAM914, output interface 9
16 through a data bus 918 and an address bus 919.

Next, the torque ratio valve 66, detent valve 64, throttle valve 65, manual valve 62, and regulator valve 6
1 will be described.

The hydraulic oil supplied to the hydraulic control circuit is supplied from a pump 52 driven by the engine, and if the line pressure is high, power consumption by the pump 52 increases accordingly. Therefore, in order to run the vehicle with low fuel consumption, it is necessary to bring the line pressure supplied to the hydraulic control circuit close to the necessary minimum. Each hydraulic servo is
It is defined by the hydraulic pressure that allows torque to be transmitted without causing the heald 33 to slip. When the engine is operated in a state that provides the best fuel efficiency, the minimum line pressure required for changes in the torque ratio T between the input and output shafts is shown by the solid line in Figure 7 using the throttle opening θ as a parameter. Within the range of torque ratios that can be realized by the pulley, it is impossible to operate the engine at the best fuel efficiency, so as shown by the dotted line, the best fuel efficiency characteristic curve shown by the solid line above is 2.
It is desirable to set the line pressure to about 0% as shown by the large broken line, and when the engine is braking, the throttle opening θ = 0.
Also, it is desirable to have a line pressure characteristic higher than that shown by the dashed line.

In this embodiment, the line pressure, which is the output of the regulator valve 61, is controlled by the hydraulic pressure adjusting device 60 to the manual valve 6.
2 shift positions (L, DSN, R.

P), the throttle opening θ, and the torque ratio of both pulleys (torque ratio between the input and output shafts) are adjusted as follows.

D position As shown in Table 1, in the manual valve 62, only the oil passage 3 communicates with the oil passage 1, and the oil passage 4 and the oil passage 5 are discharged. At this time, in the shift control mechanism 70, if the shift control solenoid 74 is in the OFF state and line pressure is supplied to the oil chamber 713, the spool 712 is positioned to the right, so that the oil passage 3 and the oil passage 1 The line pressure supplied to the oil passage 3 acts on the hydraulic servo 49 of the forward multi-disc clutch 45 through the oil passage 13, and the vehicle becomes ready to move forward.

(1) Torque ratio T is 1. ≦T:! When att.

As shown in FIG. 5(A), the torque ratio valve 66
The port 663 connected to the drain port 665 is closed, and the oil passage 6 is communicated with the drain port 665 to exhaust pressure.

As a result, regardless of the throttle opening θ, the oil passage 7
Detent pressure (equal to line pressure) is not generated in the throttle valve 65, and the port 664 of the torque ratio valve 66 communicating with the oil passage 9 is closed, and the spool 651
Since the land 657 receives feedback pressure in addition to the land 656, the throttle opening θ changes as shown in Fig. 8(c).
The throttle pressure having the characteristics shown in 7) is output to the regulator pulp plunger 612 of the regulating valve 61 via the oil passage 8.

As a result, the line pressure output from the regulating valve 61 is adjusted to the 9th (
(v) and as shown in (e) of Fig. 10.

(2) When the torque ratio T is t, <T≦t.

As shown in FIG. 5(B), the torque ratio valve 66 is connected to port 6.
63 is closed, allowing the oil passage 9 and the drain boat 666 to communicate with each other. In addition, the pressure in the oil passage 6 is exhausted through the port 665. Therefore, no detent pressure is generated, and the throttle pressure increases by the amount that the oil passage 9 is exhausted and feedback pressure is no longer applied to the land 657 of the spool 651.
It is expressed by the characteristic curve shown in (d) in the figure. The line pressure at this time has the characteristics shown in area (R) in FIG. 9 and (G) in FIG.

(3) When the torque ratio T is t, <'l'≦t4.

As shown in FIG. 5(C), the oil passage 9 is connected to the drain boat 666.
Therefore, the throttle pressure is expressed as (d) in FIG. 8, similar to (2) above. However, since the port 663 opens and the oil passage 1 and the oil passage 6 communicate with each other, the throttle opening degree θ is within the range of 0≦θ≦01%, and the detent valve 64
While the spool 641 is on the left side in the figure as shown in FIG. However, the throttle opening θ is 01%〈0510
When it is 0%, the spool 641 as shown in Fig. 4(B)
moves, the oil passage 6 and the oil passage 7 communicate with each other, and detent pressure is generated in the oil passage 7. As a result, the line pressure has a characteristic that changes in a stepwise manner at θ=01%, as shown in area (w) of FIG. 9 and (b) of FIG. 1O.

L position The oil passage 5 communicates with the oil passage 1 in the manual valve 62 .

Oil passage 3 and oil passage 4 are the same as the D position.

(1) When the torque ratio T is t, ≦T≦t2.

When the throttle opening degree θ is 050508%, the oil passage 5 and the oil passage 7 communicate with each other at the detent valve 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure. 01% to 05100%, the pressure in the oil passage 7 is exhausted through the oil passage 6 and the drain port 665 of the torque ratio valve as shown in FIG. 4(B), and no detent pressure is generated, and the throttle pressure is The same is true for the D position. Therefore, the line pressure has the characteristics shown in (R) in FIG.

(2) When the torque ratio T is t, <T≦t.

The difference from the above (1) is that in the torque ratio valve 66, the oil passage 9 communicates with the drain boat 666 to exhaust pressure, and the throttle pressure output from the throttle valve 65 to the adjustment valve 61 via the oil passage 8 increases. In particular, the line pressure is expressed by a characteristic curve as shown in FIG. 11(h).

(3) When the torque ratio T is t<T≦t4.

The oil passage 6 and the oil passage 1 are communicated with each other by the torque ratio valve 66, and the pressure of the oil passage 9 is discharged from the drain boat 666. Since line pressure is supplied to both the oil passage 6 and the oil passage 5, the detent valve 64 outputs detent pressure regardless of the throttle opening, and inputs the detent pressure and the same throttle pressure as in (2) above. The regulating valve 61 outputs the line pressure shown in FIG. 11 (N).

As shown in R position table (2), in the manual valve 62, the oil passage 4 and the oil passage 5 communicate with the oil passage 1, and the oil passage 3 is depressurized. At this time, in the shift control mechanism 70, if the shift control solenoid 74 is in the OFF state and line pressure is being supplied to the oil chamber 713, the spool 712 is positioned to the left, so that the oil passage 4 is connected to the oil passage 14. are connected, and the oil passage 4
The line pressure supplied to the hydraulic servo of the reverse multi-disc brake 42 is supplied through the oil passage 14 to the hydraulic servo 4 of the reverse multi-disc brake 42, and the vehicle enters the reverse state. Furthermore, since the line pressure is guided to the oil passage 5, the line pressure has the same characteristics as in the L position. At the R position, the torque ratio T in the V-belt continuously variable transmission 30 is set to the maximum, T=t4. Therefore, there is no need to perform a speed change (deceleration) within the planetary gear transmission 40, but according to this embodiment, even if the torque ratio T is changed at the R position, the line pressure remains the same as at the L position. control is possible.

In the P position and N position manual valve 62, oil passages 3.4 and 5 are both exhausted, and since oil passage 5 is exhausted, the line pressure that is the output of the regulator valve 61 is the same as in the D position. .

In this line pressure adjustment, the manual valve 62 is
When shifting to each shift position P, the line pressure when the torque ratio T is in the range of t, <T≦t, as shown in the characteristic curve (S) in Fig. 10, is set at a throttle opening of 01% or less. The reason why the line pressure is set low is when the line pressure is set high under operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low, and when the oil temperature is high and there is large oil leakage from various parts of the hydraulic circuit. This is because it becomes difficult to maintain line pressure, and furthermore, the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which is likely to cause trouble. In addition, when the manual valve 62 is shifted to the G and R shift positions, the torque ratio T is t1≦T≦t2 as shown in the characteristic curves (H) and (L) in FIG.
The reason why the line pressure is set high under the operating conditions where the throttle opening θ is 61% or less is that a relatively high oil pressure is required during engine braking even when the throttle opening is low. The required oil pressure at that time is 7th.
It is indicated by a dashed line in the figure. In this way, as shown in FIG. 9, by bringing the line pressure close to the minimum necessary oil pressure shown in FIG. 7, the power loss caused by the pump 52 can be reduced, thereby improving fuel efficiency and fuel consumption rate.

Next, the shift control mechanism 70 and torque ratio control device 80 controlled by the electric control circuit 90 explained in FIG.
The operation will be explained with reference to the program flowcharts shown in FIGS. 18 to 23.

In this embodiment, the electric control circuit 90 controls the input pulley rotation speed N to obtain the best fuel efficiency at each throttle opening θ.
An example of controlling is shown.

Generally, when operating an engine at its best fuel efficiency,
The vehicle is operated according to the best fuel efficiency power line shown in the broken line in the graph of FIG. In this Figure 12, the horizontal axis is the engine rotation speed (rp
m), the vertical axis shows the torque (kg·m) of the engine output shaft, and the best fuel efficiency power line is obtained as follows. That is, the constant fuel consumption rate curve of the engine (shown by the solid line in FIG. 12)
From the equal horsepower curve (units are ps) shown by the two-point 'ut line (unit: g/ps-h), the fuel consumption rate at point A in the figure is Q < g/pS・h), and the horsepower is P( ps), then
At point A, fuel of S = Q x P (g/h) will be consumed every hour. By determining the fuel consumption per hour S at all points on each equal horsepower curve, the point where S is the minimum on each equal horsepower curve can be determined, and by connecting these points, the best value for each horsepower can be determined. The best fuel efficiency power line indicating the engine operating state resulting in fuel efficiency can be obtained. However, when the engine 100 and the fluid coupling 21, which is a fluid transmission mechanism, are combined as in this embodiment, the engine output performance curve at each throttle opening θ shown in FIG. , find the best fuel efficiency fluid coupling output line on the fluid coupling output performance curve as shown in FIG. 16 from the fluid cambling performance curve shown in FIG. 14 and the engine equivalent fuel consumption rate curve shown in FIG. 15. Can be done. FIG. 17 shows the best fuel efficiency fluid coupling output line shown in FIG. 16 replaced by the relationship between throttle opening and fluid coupling output rotation speed. This fluid coupling output rotation speed directly becomes the input pulley rotation speed in the continuously variable transmission of this embodiment.

In the continuously variable transmission device of this embodiment, the input pulley 31 and the output pulley 32 are Controls the gear ratio.

The torque ratio control device 80 controls the increase or decrease in the gear ratio between the input and output pulleys by comparing the rotational speed of the input pulley with the best fuel efficiency obtained in FIG. 17 and the actual rotational speed of the input pulley. This is done by operating two solenoid valves 84 and 85 provided in the control device 80, and the actual input side pulley rotation speed is made to match the input side pulley rotation speed for the best fuel efficiency. FIG. 18 shows an overall flowchart of input end pulley rotation speed control.

After reading the throttle opening θ using the throttle sensor 904 (921), the shift lever switch 90
1, the shift lever position is determined (922). As a result of the determination, if the shift lever is in the P position or N position, both the solenoid valves 84 and 85 are set to FFL by the P position or N position processing subroutine shown in FIG.
, (931), and stores the P or N state in the RAM 914 (932). As a result, the input pulley 31 is brought into a neutral state. Shift lever is in P position or N
When changing from position to R position and from N position to D position, N-R shift and N
- Perform shift shock control processing to alleviate shift shock associated with D shift (940, 950)
. Shift shock control is performed by applying pulses whose pulses gradually become smaller to the shift control solenoid valve 74 shown in FIG. 20 (hereinafter referred to as duty control). By duty-controlling the shift control solenoid valve 74 in this manner, a hydraulic pressure P adjusted in accordance with the duty is generated in the oil chamber 713 of the shift control valve 71.

The shift control mechanism 70 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49 of the free gear transmission 4o by the action of the solenoid valve 74 controlled by the output of the electric control circuit 90 described above, and reduces the impact during shifting. In addition, the pressure limiting valve 73 has the function of keeping the upper limit of the hydraulic pressure supplied to the hydraulic servos 48 and 49 below a set value, thereby limiting the engagement pressure of the clutch and brake.

To alleviate the locking problem during N-D shift and N-R shift, use the hydraulic servo 48 or the hydraulic servo 4.
9, the rise of the oil pressure P or PC is rolled as shown in the oil pressure characteristic curve shown in FIG.
complete the bond. FIG. 25 shows the relationship between the duty (%) of the solenoid valve 74 for controlling the hydraulic pressure supplied to the hydraulic servo 48 or 49 and the solenoid pressure P generated in the oil chamber 713 by the operation of the solenoid valve 74. , duty (%) is given by the following formula.

Duty (%) = The solenoid pressure shown in FIG. 25 is amplified by the shift control valve 71 and the hydraulic servo 48 or 49 shown in FIG.
A supply oil pressure Pb or P is obtained.

In this embodiment, as shown in FIG. 24(A), the pressure receiving area of the land provided on the spool 712 of the shift control valve 71 is S+ -S+, S+, St, and the elastic force of the spring 7110 is F, in order from the left side in the figure. Assuming that the oil pressure in the oil chamber 713 is P3, the oil pressures PC and P supplied to the hydraulic servo 49 of the multi-disc clutch 45 that is engaged during forward movement and the hydraulic servo 48 of the multi-disc brake 42 that is engaged during reverse movement are, respectively. The 0th formula which is the hydraulic balance formula of the shift control valve 71 and ■
From 2, it is given as follows.

When moving forward Ps XS+ =Pc XSz +F! I
■S t S x When moving backward Ps x3. = pb X (S+ St
) 10 FSI ■ S IS 2 S + S z Also, the valve body 731 inserted in the pressure limiting valve 73
The pressure-receiving area of is S1, and the elastic force of the spring 732 installed on the valve body 731 is F,2, then the pressure limiting valve 73 operates at the highest pressure P3 according to the hydraulic balance equation (2).

pj! 1m1tXs3=Fsz
■p llimit-F sz/ S s At this time, PC
and P, is the highest pressure p e according to the 0th equation and the 0th equation
It 1m1t, p, A 1m1t is limited and St s when moving forward.

When going backwards pb 11m1t = Q! imi tS
+ S z S + S z Returning to FIG. 18, following the N-D shift shock control processing 950, the input pulley rotation speed sensor 902 reads the actual input pulley rotation speed N (923). ), then it is determined whether the throttle opening degree θ is O or not (924), and if θ≠0, the data is stored in the ROM 913 in advance according to the subroutine shown in FIG.
Set the best fuel efficiency input pulley rotation speed N0 corresponding to the throttle opening θ shown in FIG. 17 (960).
), set the storage address of the input side boolean rotation speed N0 data corresponding to the throttle opening degree 961), and read the data of N'' from the set address 962
), the read data of N1 is stored in the data storage RAM 91.
4 (963).

Next, the actual input-side pulley rotation speed N and the best fuel economy input-side pulley rotation speed N9 are compared (927), and when N<N'', an operation command for the downshift solenoid valve 84 is issued (928), >N", an operation command is issued for the upshift solenoid valve 85 (929), and when N=N", an OFF command is issued for both solenoid valves 84 and 85 (929).
920).

When θ=0 and the throttle is fully closed, it is determined whether the shift lever is set to the D position or the L position in order to determine the necessity of engine braking (
926), and engine braking processing 970 or 980 is performed as necessary. D position engine brake processing 97
0, as shown in FIG. 22, the vehicle speed V is read by the vehicle speed sensor 903 (971), the acceleration α at that point is calculated (972), and then the acceleration α is an appropriate acceleration A for the vehicle speed. It is determined whether or not (973). α〉
In the case of A, in order to perform downshift control 974, return after setting N1 to a value larger than N, and α≦
In the case of A, the best fuel consumption input side boolean rotation speed N0 corresponding to the throttle opening degree of 0 is set in N'' (975), and then the process returns.

The relationship between vehicle speed and appropriate acceleration A is determined by experiment or calculation for each vehicle, and is shown in (Fig. 22).
It is shown in the graph of B).

In the engine brake processing 980 for the L position, as shown in FIG. 23, after reading the vehicle speed (981), an operation is performed to calculate the torque ratio T from the vehicle speed V and the input pulley rotation speed N from the following formula. (982).

T=(N/V)Xk k is a constant determined from the reduction ratio of the reduction gear mechanism 23 inside the transmission, the final reduction ratio of the vehicle, the tire radius, etc. Next, the current torque ratio T is the vehicle speed■
It is determined whether the torque ratio TI is greater than the one that provides safe and appropriate engine braking (983).
, T<T'' so that a downshift is performed.
Set a value larger than N (984), and T≧T0
In this case, set N1 to a value equal to N (985).
Return. The torque ratio T0 that provides safe and appropriate engine braking for each vehicle speed is determined by experiment or calculation for each vehicle, and is shown in Figure 23 (
It is shown in the graph of B).

Next, the operation of the torque ratio control device 80 will be explained with reference to FIG. 27.

When the vehicle is traveling at a constant speed, the solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned off, as shown in FIG. 27(A). As a result, the oil pressure P in the oil chamber 816 becomes the line pressure, and the oil pressure P2 in the oil chamber 815 also becomes the line pressure when the spool 812 is on the right side in the figure. Spool 812
is moved to the left in the figure because of the pressing force P3 by the spring 811. When the spool 812 is moved to the left and the oil chamber 815 and the drain boat 813 communicate with each other, the pressure P2 is exhausted, so the spool 812 is moved to the left to connect the oil chamber 816 to the oil pressure P+.
is moved to the right in the diagram. When the spool 812 is moved to the right, the drain boat 813 is closed. Therefore, in this case, as shown in FIG. 27, the spool 812 is provided with a flat notch 812b at the land edge of the drain boat 813 and the spool 812, so that the spool 812 can be moved in a more stable state as shown in FIG. 27(A).
It is possible to maintain an equilibrium point at an intermediate position as shown in FIG.

In this state, the oil passage 2 is closed, and the oil pressure of the hydraulic servo 313 of the input pulley 31 is applied to the output pulley 3.
Due to the line pressure applied to the hydraulic servo 323 of No. 2, V
It becomes compressed via the belt 33 and is eventually balanced with the hydraulic pressure of the hydraulic servo 323. Actually, since there is oil leakage in the oil passage 2 as well, the input pulley 31 is gradually expanded and the torque ratio T changes in the direction of increasing. Therefore, as shown in No. 275 (A), at the position where the spool 812 is balanced, a flat notch 812a is formed on the land edge with the spool 812 so that the drain boat 814 is closed and the oil passage 1 is slightly open. established,
The oil leakage in the oil passage 2 is compensated for. Also the second
It is clear that the same function can be achieved even if the oil passage 822 having an orifice 821 is connected between the oil passage 1 and the oil passage 2 instead of the notch 812a as shown in FIG.

During an upshift, the solenoid valve 85 is turned on by the output of the electric control circuit as shown in FIG. 27(B). As a result, the pressure in the oil chamber 816 is evacuated, so the spool 812 is moved to the left in the figure. As the spool 812 moves, the pressure in the oil chamber 815 is also evacuated from the drain boat 813. However, due to the action of the spring 811, the spool 812 is set at the left end in the diagram.

In this state, the line pressure of oil passage 1 is supplied to oil passage 2 via port 818, so the oil pressure of hydraulic servo 313 increases, input pulley 31 operates in the direction of closing, and torque ratio T decreases. . Therefore, by controlling the ON time of the solenoid valve 85 as necessary, the torque ratio is reduced by a desired amount to perform an upshift.

During a downshift, the solenoid valve 84 is turned on by the output of the electric control circuit as shown in FIG. 27(C), and the pressure in the oil chamber 815 is evacuated.

The spool 812 is moved to the right in the figure by the line pressure of the oil chamber 816, the oil passage 2 is communicated with the drain boat 814 and the pressure is discharged, and the input pulley 31 is operated in the expanding direction to increase the torque ratio. By controlling the ON time of the solenoid valve 84 in this way, the torque ratio is increased and a downshift is performed.

As described above, the hydraulic servo 312 of the input side (drive side) pulley 31 is supplied with the output hydraulic pressure of the torque ratio control valve 81, and the line pressure is guided to the hydraulic servo 323 of the output side (driven side) pulley 32. If the oil pressure of the input side hydraulic servo 312 is Pl and the oil pressure of the output side hydraulic servo 322 is Po, then P, /P is the 28th
It has the characteristics as shown in the graph in the figure, for example, when the accelerator is loosened from a state where the throttle opening is θ -50% and the torque ratio T = 1.5 (point a in the figure), the accelerator is loosened. If it is 30%, P-/P! When the torque ratio is maintained as it is, it shifts to the point in the figure where the torque ratio T=0.87, and conversely, the torque ratio T=0.87.
1.5, the value of P-/Pr is increased by the output of the torque ratio control mechanism 80 that controls the input pulley and changed to the value of point C in the figure. In this way, P0/Pr is increased. Pi
By controlling the value of as necessary, it is possible to set any torque ratio corresponding to any load condition.

[Brief explanation of the drawing]

Fig. 1 is a schematic diagram of a belt-type continuously variable transmission for vehicles, Fig. 2 is a hydraulic control circuit diagram of a continuously variable transmission showing one embodiment of the present invention, and Fig. 3 explains the operation of a manual valve. Figure 4 is a diagram for explaining the operation of the detent valve and throttle valve. Figure 5 is a diagram for explaining the operation of the torque ratio valve. Figure 6 is a configuration diagram of the electric control circuit. Figure 7 is a diagram showing the required line pressure characteristics of the hydraulic control circuit, Figure 8 is a diagram showing the throttle pressure characteristics, Figures 9, 10, and 1.
1 is a diagram showing the line pressure characteristics obtained by the control device of the present invention, FIG. 12 is a diagram showing the best fuel efficiency power line of the engine, and FIG. 13 is a diagram showing the characteristics of the engine output performance.
Figure 4 shows the performance curve of the fluid transmission mechanism, Figure 15 shows the equal fuel consumption rate curve of the engine, Figure 16 shows the best fuel efficiency fluid coupling output curve, and Figure 17 shows the best fuel efficiency fluid cup. Diagram showing the characteristics of ring output rotation speed,
Figures 18, 19, 21, and 22 (A)
, FIG. 23(A) is a diagram for explaining the flow of processing in the electric control circuit, FIG. 20 is a diagram for explaining the action of the solenoid valve, FIG. 22(B) is a diagram showing the set acceleration, Figure 23 (B) is a diagram showing the set torque ratio, Figure 24 (
A) is a diagram for explaining the operation of the shift control mechanism;
Figure 4 (B) is a diagram showing the characteristics of the oil pressure supplied to the input and output side hydraulic servos, Figure 25 is a diagram showing the characteristics of the solenoid pressure, and Figure 26 is a diagram showing the characteristics of the output oil pressure of the shift control valve. , Fig. 27 is a diagram for explaining the operation of the torque ratio control device, Fig. 28 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo, and Fig. 29 is a diagram for explaining the operation of the torque ratio control device. FIG. 30 is a schematic diagram of a conventional V-belt type continuously variable transmission for a vehicle. 214... Input shaft, 26... Output shaft, 31... Input side pulley, 32... Output side pulley, 33... V belt, 30... Continuously variable transmission, 90... Electric control circuit, 61...Regulator valve, 64...Detent valve, 65...Throttle valve, 66...Torque register first
．． 313.323... Hydraulic servo, 71... Shift control valve, 81... Torque ratio control valve. Applicant Aisin Warner Co., Ltd. Figure 1 Figure 3 Figure 4 (A) (Snake) Figure 5 (Snake) Figure 6 Figure 7 (K9/Cm2) Figure 12 Figure 13 Figure 14 Figure 15 Figure 16 Figure 17 Figure 18 Figure 19 @ Figure 21 Figure 20 Figure 22 (A) (Snake) Figure 23<A) (B) Figure 24 (A) ! [Lra] Pc1A+tPb (kg/crn”) Ps Kawa-hachi\-k (C) Fig. 28 Fig. 29

Claims (4)

[Claims] (1) Consists of an input-side pulley and an output-side pulley that are attached to the input and output shafts and have variable effective diameters, and a drive band stretched between these pulleys, and the effective diameter of the pulleys is adjusted by hydraulic pressure. In a vehicle V-belt continuously variable transmission that controls the torque ratio between the input and output shafts, there is a hydraulic pressure source and an adjustment system that generates line pressure by regulating the pressure of hydraulic fluid from the hydraulic source according to the engine output torque. A first hydraulic servo that is supplied with the line pressure and generates a clamping force for the drive band according to the line pressure, and a pressure increasing means that increases the line pressure during engine braking. Hydraulic control device for vehicle V-belt continuously variable transmission. (2) The pressure increasing means includes an engine brake state determining means for determining an engine brake state, and a signal for supplying a pressure increasing signal oil pressure to the pressure regulating device when the engine brake state determining means determines that the engine brake state is present. A hydraulic control device for a V-belt type continuously variable transmission for a vehicle according to claim 1, further comprising a hydraulic pressure supply means. (3) The pressure regulating device includes a throttle valve that generates a throttle pressure according to a throttle opening degree that is approximately proportional to the engine output torque, and a throttle valve that is supplied with the throttle pressure and the signal oil pressure to generate the throttle pressure and the signal oil pressure. 3. The hydraulic control device for a V-belt type continuously variable transmission for a vehicle according to claim 2, further comprising a regulator valve that regulates line pressure according to the line pressure. (4) The engine brake state determining means is a means for determining that the engine brake state is present when the throttle opening is less than or equal to a predetermined value. Hydraulic control system for V-belt continuously variable transmission for vehicles.