JPH023749A - Oil pressure regulating device for v-belt type continuously variable speed change gear for vehicle - Google Patents

Abstract

PURPOSE:To improve durability of a V-belt, by a method wherein a line pressure responding to a torque ratio between an input and an output shaft is generated by a control means, and a necessary minimum belt nipping force in a given torque ratio is ensured. CONSTITUTION:Working oil from a hydraulic source 50 is regulated to the line pressure of an oil passage 1 to the hydraulic servo on the output side by means of a regulator valve 61. An oil pressure responding to a torque ratio between an input and an output shaft is generated by a torque ratio valve (control means) 66 having a spool 662 linked to the moving flange of a pulley on the output side, and feedback is made on the line pressure through the working of a throttle valve 65. Thus, the line pressure is changed by means of a torque ratio between two pulleys, serving as parameter, and a necessary minimum belt nipping force in a given torque ratio is ensured. An oil pressure regulated by a torque ratio control means 80 is fed through an oil passage 2 to the hydraulic servo of a pulley on the input side. This constitution prevents a nipping force more than necessary from being exerted on a V-belt, and enables improvement of durability.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention is applied to a vehicle that adjusts oil pressure (line pressure) supplied to a hydraulic servo mechanism that generates a clamping force for a drive belt of a belt-type continuously variable transmission; ■Relating to hydraulic adjustment devices for belt-type continuously variable transmissions.

[Conventional technology]

A belt-type continuously variable transmission uses an endless metal belt invented by Pan d'Orne wrapped around two pulleys. It is characterized in that the type blocks are arranged without interruption, and power is transmitted between the side surfaces of the pulley and the V-shaped block, and by the pressing force between the two-shaped blocks.

FIG. 33 shows a conventional continuously variable transmission for a vehicle disclosed in Japanese Patent Application Laid-Open No. 159730/1983. input shaft a
is provided with an input end pulley consisting of a fixed flange b and a movable flange C, and an output side pulley consisting of a fixed flange e and a movable flange f is provided on the output shaft d. A bell) g is tensioned and transmits power from the input shaft a to the output shaft d. This control device for a continuously variable transmission for a vehicle is provided with a regulator valve k that regulates the hydraulic pressure generated by a pump j to a predetermined line pressure, and the line pressure regulated by the regulator valve k is passed through an oil path i. is supplied to the movable flange r of the output pulley to give the output pulley a force to clamp the belt g, and the oil path connected to the movable flange C of the input pulley is selectively supplied by the torque ratio control valve l. Hydraulic pressure is supplied and discharged to move the movable flanges C and r.

Moreover, at one end of the spool m of the torque ratio control valve l,
Fluid pressure proportional to the rotational speed of the input shaft a acts on the pitot tube n, while on the other end of the spool m, pressure due to the rotation of the cam p linked to the movement of the throttle pedal is applied to the lever q1 spring. Therefore, the movable flange C of the input pulley by the torque ratio control valve l
Hydraulic pressure is supplied and discharged according to the rotational speed of the input shaft a and the movement of the throttle pedal. Furthermore, fluid pressure proportional to the rotational speed of the human power shaft a acts on one end of the spool S of the regulator valve through the pitot tube n, while the input pulley is affected by the axial movement of its movable flange C. A detection rod t is provided which is displaced in conjunction with the spool S, and the pressure of the detection rod t acts on the other end of the spool S via a lever U and a spring V.

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

[Problem to be solved by the invention]

■In a belt-type continuously variable transmission for vehicles, the minimum required belt clamping force to prevent slippage between the belt and pulley is determined by the line pressure and the hydraulic servo of the driven pulley in Figure 7. It is the product of the pressure-receiving area, and therefore,
The characteristics are similar to the curve shown by the solid line in FIG. In other words, the minimum increase rate of the belt clamping force is determined by the torque ratio (
(corresponding to the pulley position) becomes larger.

However, in the conventional belt-type continuously variable transmission described above, the signal according to the torque ratio is received by the load of the spring V linked to the movement of the detection iron t, so the line pressure is linearly adjusted according to the torque ratio. It has characteristics that change to
If the line pressure is set low, the clamping force will be insufficient in the high torque ratio range, and the belt will slip during city driving where starting and running at low speeds are repeated, resulting in scuffing between the belt and the pulleys, which may reduce the durability of the belt. The problem is that the belt slips during a sudden start, making it impossible to start, and if the engine kicks down while running in overdrive, the belt slips during the transition to underdrive, causing the engine to overrun. However, there is a problem in that scuffing occurs between the belt and the pulley, resulting in damage to the belt.

On the other hand, if the line pressure is set high, the clamping force will be excessive in a small torque ratio range, and oil pump loss will increase during frequent high-speed driving, resulting in poor fuel efficiency. In addition, in high-speed conditions, the linear velocity of the belt is high and the contact radius between the output pulley and the belt is small, resulting in a large bending of the belt.If an excessive clamping force is applied to the belt in this condition, the steel that makes up the belt may There are problems in that the band breaks or stress is concentrated on the blocks, impairing the durability of the belt.

In order to solve the above-mentioned problems, the present invention aims to reduce the line pressure supplied to the hydraulic servo mechanism that generates the clamping force of the drive belt of the V-belt continuously variable transmission to the torque ratio of the V-belt continuously variable transmission. The belt clamping force is controlled in accordance with the specified torque ratio to ensure the minimum necessary belt clamping force at a predetermined torque ratio.

[Means for solving problems]

The hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission of the present invention includes a fixed flange, a movable flange, and a hydraulic servo provided on the movable flange, which are attached to an input shaft (214) and an output shaft (26), respectively. It consists of input and output pulleys (31, 32) with variable effective diameters, and a drive bell (33) stretched between these pulleys, and a hydraulic servo (33) for either the input or output pulleys. 323)
A clamping force for the drive belt is generated by supplying hydraulic pressure to the input and output pulleys, and by selectively supplying hydraulic pressure to the hydraulic servo (313) on the other of the input and output pulleys, the effective diameter of both pulleys is reduced. In a hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission that controls the torque ratio between input and output shafts by adjusting the hydraulic pressure, the hydraulic pressure l1 (50) and the hydraulic oil from the hydraulic source are connected to one of the hydraulic servos. A regulator valve (61) that regulates the line pressure supplied to The flat means 66) generates hydraulic pressure according to the torque ratio, and is characterized in that it includes means for ensuring the minimum necessary belt clamping force at a predetermined torque ratio.

Note that the numbers added to the above configurations are for comparison with the drawings, and the configurations of the present invention are not limited thereby.

[Action/effect of the invention]

According to the present invention, the line pressure is changed using the torque ratio between both pulleys as a parameter, and the hydraulic pressure supplied to the hydraulic servo that generates the clamping force of the drive belt of the V-belt type continuously variable transmission is adjusted as shown in FIG. By approximating the characteristic curve of
Only the minimum necessary belt clamping force is applied to prevent the belt from slipping, and it is possible to prevent the V-belt from receiving more clamping force than necessary and damage its durability. Since generation of line pressure higher than the oil pressure to be obtained can be prevented, it is possible to prevent a decrease in fuel efficiency due to an increase in loss torque of the oil pump.

〔Example〕

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

FIG. 1 is a schematic diagram of a V-belt type continuously variable transmission for a vehicle.

100 is an engine, 102 is a carburetor, 20 is a transmission device provided between the engine 100 and the drive side axle, and a fluid coupling 21 connected to the output side 101 of the engine, and a fluid coupling 21 connected to the fluid coupling 21. V-belt type continuously variable speed 4m30, the continuously variable speed Ia30
The planetary gear transmission 40 for forward/reverse switching is connected to the output shaft 2G of the planetary gear transmission 40, and the gear mechanism 23 is connected to the output shaft 47 of the planetary gear transmission 40.

The fluid coupling 21 is connected to the output shaft 101 of the engine.
a pump impeller 211 connected to a turbine runner 21 connected to a fluid coupling output shaft 214;
This is a well-known method consisting of 2. Note that other fluid torque converters or mechanical clutches may be used instead of the fluid coupling.

■The belt type continuously variable transmission 30 includes an input pulley 31 connected to the output shaft 214 of the fluid coupling, and an output shaft 26 of the V-belt continuously variable transmission arranged parallel to the input pulley 31. The output pulley 32 is connected to the output pulley 32, and the V-belt 33 is stretched between the two pulleys.

The input pulley 31 has a fixed flange 311 connected to the output shaft 214, and a movable flange 312 provided opposite to the fixed flange 31+ to form a V-shaped space. It is provided so as to be movable in the axial direction by a hydraulic servo 313.

The output pulley 32 includes a fixed flange 321 connected to the output shaft 26 of the continuously variable transmission fi 30, and the fixed flange 32.
1 and a movable flange 322 provided so as to form a V-shaped space, and the movable flange 322 is provided so as to be movable in the axial direction by a hydraulic servo 323.

The forward/reverse switching 'tI star gear transmission 40 has a sun gear 4.
1, a ring gear 43, a sun gear 41, a double planetary gear 44 that meshes with the ring gear 43, and a carrier 46 that rotatably supports the double planetary gear 44. The sun gear 41 is connected to the output shaft 26 of the continuously variable transmission. , the carrier 46 is a J star gear shift 4i for forward/reverse switching! 40 output shaft 47. sun gear 41
and the carrier 46 are removably connected by a multi-disc clutch 45, and the ring gear 43 is removably connected to a transmission case 400 by a multi-disc brake 42.

In this planetary gear transmission 40 for forward/backward switching, when hydraulic pressure is supplied to the hydraulic servo 49, the multi-plate clutch 45 engages and the rotation of the output shaft 26 of the continuously variable transmission remains unchanged. The signal is transmitted to the output shaft 47 of the transmission 40 to enable forward travel. Further, when hydraulic pressure is supplied to the hydraulic servo 48, the multi-disc brake 42 is engaged and the ring gear is fixed, so that the output shaft 47 is connected to the output shaft 2 of the continuously variable transmission.
It rotates in the opposite direction to the rotation of No. 6 to enable a backward running state.

The reduction gear mechanism 23 is used to compensate for the fact that the speed change range achieved by the belt type continuously variable transmission 30 is lower than the speed change range achieved by a normal vehicle transmission. It decelerates and increases torque. The output shaft of the reduction gear mechanism 23 is connected to the differential gear 22, and performs final reduction at a reduction ratio of 3.727, for example.

FIG. 2 shows a hydraulic control circuit for the belt-type continuously variable transmission shown in FIG.

The oil pressure control circuit includes oil pressure 1so, oil pressure adjustment device 60, and N-
Shift control mechanism 7 that alleviates the impact during D and N-R shifts
0, and a torque ratio control device 80.

The hydraulic power source 50 supplies hydraulic oil pumped up from an oil reservoir by a pump 5z driven by the engine via an oil strainer 51 to an I/regulator valve 61 via an oil passage II to which a relief valve 53 is attached.

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. Torque ratio valves 6G, O, 1 which operate in conjunction with the movable flange 322 of the pulley 32, supply line pressure to the detent valve 64 according to its changing position, and discharge pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65.
, and a regiator valve 61 that regulates the hydraulic pressure supplied from the hydraulic pressure source 50 and supplies it to the hydraulic servo 323 as line pressure.

The manual valve 62 has a spool 621 in each position P, R, N, D, ■, as shown in FIG. and connects the oil passage 1 to which line pressure is supplied and the output oil passages 3 to 5 as shown in Table r.

table! In Table 3, ◯ 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. Adjust the line pressure and output the line pressure from the output port 16 to the oil path l. The oil discharged from the bow) 614 is supplied to the fluid coupling, oil cooler and lubrication points via the oil line 12.

The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the butterfly valve of the carburetor 102 as shown in FIG.
≦θ, as shown in FIG. ,ku0
When the torque ratio is 5100%, the oil passage 6 connects the oil passage 7 and the detent valve 64 to the torque ratio valve 66 as shown in FIG. 4(B).
communicate with.

The throttle valve 65 is arranged in series with the spool 641 of the detent valve via a spring 645, and includes a spool 651 with a spring 652 installed in front of the other side, and throttle opening transmitted via the spool 641 and the spring 645. Due to the action of the spool 651 that moves according to the fluctuation of the degree θ, the boat 65 connected to the oil passage 1
3 and outputs throttle pressure to the throttle pressure output oil passage 8 which communicates with the input port 18 provided in the regulator valve 61. The spool 651 branches from the oil path 8 and outputs to a land 656 and a land 657 having a larger pressure receiving area than the land 656 via output oil pressure feedback oil paths 9 and 10 provided with orifices 654 and 655, respectively. Receiving hydraulic feedback.

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 movement i1L of the movable flange 322! ,≦
When L≦l, (torque ratio T is t2≧T≧tl), the fifth
As shown in Figure (A), the spool 662 is located on the left side of the figure, closes the input port 64 connected to the feedback oil passage 9 of the output oil pressure provided in the throttle valve 65, and supplies the output oil to the detent valve 64. The passage 6 is communicated with the drain boat 665 to exhaust pressure.The amount of movement of the movable flange 322 is smaller than the first set value l, and 12≦1.<
it, (ts ≧T>t), the spool 662 is located in the middle as shown in FIG. is depressurized.

When the movement IL of the movable flange 322 is smaller than the second set value 12 and 0≦L≦lz (L4≧T>t,), the spool 662 is positioned on the right side as shown in FIG. 5(C). However, the boat 663 connected to the oil passage 1 and the oil passage 6 communicate with each other, and line pressure is supplied to the oil passage 6.

The shift control mechanism 70 includes a shift control valve 71 including a spool 712 having a spring 711 installed in front of one end and receiving line pressure from an oil chamber 713 provided at the other end;
Orifice 7 provided in oil passage 1 that supplies heline pressure
2. It consists of a pressure limiting valve 73 installed between the orifice 72 and the oil chamber 713, and a solenoid valve 74 that adjusts the oil pressure of the oil chamber 713 under the control of an electric control circuit to be described later.

When the solenoid valve 74 is turned on to open the drain boat 741 and evacuate the oil chamber 713, the shift control valve 71
The spool 712 is moved to the left in the figure by the action of the spring 7it, and the multi-plate clutch 45 of the planetary gear shift 1!40
The oil passage 13 that communicates with the hydraulic servo 49 that operates the multi-disc clutch 42 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. Alternatively, the multi-disc brake 42 is released.

Drain port when solenoid valve 74 is off)
741 is closed, and the spool 712 is located on the right side in the figure with line pressure supplied to the oil chamber 713, and connects the oil passages 3 and 4 to the oil passages 13 and 14, respectively, and connects the multi-disc brake 42. Alternatively, the multi-disc clutch 45 is engaged.

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 engaged.

The torque ratio control device 80 includes a torque ratio control valve 81, orifices 82 and 83, and a downshift solenoid valve 84.
, and an upshift solenoid valve 85.

The torque ratio control valve 81 includes a spool 812 with a spring 811 installed in front of it on one side, oil chambers 815 and 816 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, respectively, and line pressure is supplied. Oil road 1
The input side of the human-powered boat 817 and the V-belt continuously variable transmission 30 whose opening area increases and decreases according to the movement of the spool 812. The oil chamber 819 is equipped with an output boat 818 that communicates with the hydraulic servo 313 of J 31 via the oil passage 2.
drain boat 814 and spool 812
The train boat 8 evacuates the oil chamber 815 according to the movement of the train boat 8.
13. The downshift solenoid valve 84 and the amplifier shift solenoid valve 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.
The pressure in the oil chamber 815 and the oil chamber 816 is 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 valve 84 and the upshift solenoid valve 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 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, and 905 is a rotation speed sensor 902
906 Speed detection processing circuit that converts the output of
907 is a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage; 908 to 911 are human interfaces for each sensor; 912 is a central processing unit (CPIJ), 91.3 is solenoid valve 74.84
．． Read-on memory (ROM), which stores programs to control the 85 and data necessary for control, 91
4 is a random access memory (RAM) for temporarily storing input data and parameters necessary for control, 9
15 is a clock, 916 is an output interface, 91
7 is a solenoid output driver, and output interface J
, Solenoid valve 85 for upshifting the output of chair 91G
, to the operating output of the downshift solenoid valve 84 and the shift control solenoid valve 74. Input interfaces 908 to 911, CPU 912, ROM
913, RAM 914, and output interface 916 are connected through a data bus 918 and an address bus 919.

Next, the operation of the hydraulic pressure adjusting device 60, which is a feature of the present invention and is comprised of a torque ratio valve 66, a detent valve 64, a throttle valve 65, a manual valve 62, and a regulator valve 61, will be explained.

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. The hydraulic pressure is defined by the hydraulic pressure at which the hydraulic servo can transmit torque without causing the V-belt 33 to slip.

When the engine is operated in a state that provides the best fuel efficiency, the minimum necessary line pressure for changes in the torque ratio T between the input and output shafts is shown by the solid line in FIG. 7 using the throttle opening θ as a parameter. When starting the vehicle, it is impossible to operate the engine at the best fuel efficiency within the range of torque ratio that can be achieved by both pulleys, so as shown by the dotted line, it is approximately 20% lower than the best fuel efficiency characteristic curve shown by the solid line above. It is desirable to set the line pressure to the line pressure shown by the large broken line, and even when the throttle opening is θ=0 during engine braking, there is one point! 1
It is desirable to have a higher line pressure characteristic expressed in wA.

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 position ff (L, D, N, R.

P) is adjusted as follows by changing the throttle opening θ and the torque ratio between the input and output shafts.

<Position D) As shown in Table 1 above, 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 depressurized. At this time, in the shift control TI mechanism 70, if the shift control solenoid valve 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 Oil road 1
3 is connected, and the line pressure supplied to oil path 3 is connected to oil path 1.
Hydraulic servo 49 of multi-disc clutch 45 for forward movement through 3
The vehicle is now able to move forward.

■ When the torque ratio T is t, ≦T≦t2.

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

As a result, regardless of the throttle opening θ, the oil passage 7
No detent pressure occurs. Further, the throttle valve 65 is
Since the boat 664 of the torque ratio valve 66 connected to the oil passage 9 is closed and the spool 651 receives feedback pressure from the land 657 in addition to the land 656, the difference in the throttle opening θ shown in FIG. A throttle pressure having the characteristics shown in is outputted to the regulating valve plunger 612 of the regulating valve 61 via the oil passage 8. As a result, the line pressure output from the regulator 61 is reduced to the area shown in FIG.
The result is as shown in Figure O (E).

■ Torque ratio T is t! When <1'≦t.

As shown in FIG. 5(B), the torque ratio valve 66 is connected to the boat 6.
63 is closed, allowing the oil passage 9 and the drain boat 666 to communicate with each other. The oil passage 6 is also depressurized through the boat 665. Therefore, no detent pressure is generated, and the throttle pressure increases by the amount that the pressure in the oil passage 9 is exhausted and the feedback pressure is no longer applied to the land 656 of the spool 651, and is expressed by the characteristic curve shown in FIG. 8(d). Ru. The line pressure at this time has the characteristics shown in area (R) in FIG. 9 and (G) in FIG.

■ When the torque ratio T is t, <'l'≦[,.

As shown in FIG. 5(C), the oil passage 9 is connected to a drain boat 666.
Therefore, the throttle pressure is the same as in ① above.
This is shown in figure (d). However, since the bow) 663 opens and the oil passage 1 and oil passage 6 communicate with each other, the throttle opening θ becomes 0.
While ≦θ≦θ1% and the spool 641 of the detent valve 64 is on the left side of the figure as shown in FIG. 4(A), the spool 641 closes the oil passage 6 and closes the oil passage 7. is discharged from the manual valve 62 via the oil path 5, but when the throttle opening θ is between 06% and 05100%, the spool 641 moves to the right in the figure as shown in FIG. 4(B). 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 is reduced to (
As shown in area (w) and (su) in Figure 1, θ=θ6%
The characteristic changes stepwise.

(L position) At the manual valve 62, the oil passage 5 communicates with the oil passage l.

Oil passage 3 and oil passage 4 are equivalent to the D position.

■ When the torque ratio T is tl≦T≦11.

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. Throttle opening θ is 01%
5100%, the oil passage 7 is the oil passage 6 and Fig. 4 (B)
As shown in the figure, the pressure is exhausted through the drain boat 665 of the torque ratio valve, so no detent pressure is generated, and the throttle pressure is the same as in the case of the D position. Therefore, the line pressure is the first
The characteristics are shown in (R) in Figure 1.

■ When the torque ratio T is t, <T≦t.

The difference from the above (■) is that in the torque ratio valve 66, the oil passage 9
communicates with the drain port 1666 and is exhausted, and the throttle pressure output from the throttle valve 65 to the regulating valve 61 via the oil passage 8 increases, and as a result, the line pressure reaches (h) in Fig. 11. It is expressed by the characteristic curve as shown.

■ When the torque ratio T is 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 exhausted from the drain port 666. Since line pressure is supplied to both the oil passage 6 and the oil passage 5, the detent valve 64 outputs the detent pressure regardless of the throttle opening, and the detent valve 64 outputs the detent pressure and the regulator valve that manually generates the same throttle pressure as in (2) above. 61 outputs the line pressure shown in FIG. 11 (N).

(R position) In the manual valve 62, the oil passage 4 and the oil passage 5 are oil passage l.
The oil passage 3 is depressurized. At this time, in the shift control mechanism 70, the shift control solenoid 74 is turned off.
When line pressure is being supplied to the oil chamber 713 in the F state, the spool 712 is positioned to the right, so that the oil passage 4 and the oil passage 14 are communicated with each other, and the line pressure supplied to the oil passage 4 is is supplied to the hydraulic servo 43 of the reverse multi-disc brake 42 through the oil path 14, and the vehicle enters the reverse state.

(P position and N position) Since both oil passages 3.4 and 5 are under pressure in the manual valve 62, 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 of P, the line pressure when the torque ratio T is in the range of t<Tst4 is set low at the throttle opening θ1% or less as shown in the characteristic curve (wa) in Figure 1O. The reason for this is to increase the line pressure in operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low. This is to prevent this from happening, since it becomes difficult to maintain the pressure, and furthermore, the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which tends to cause trouble.

In addition, when the manual valve 62 is shifted to the R and R shift positions, the characteristic curves (H) and (L) in FIG.
As shown in , when the torque ratio T is 1. The line pressure was set high under operating conditions in the range of ≦T≦t2 and the throttle opening θ is 0.1% or less because relatively high oil pressure is required even at low throttle opening during engine braking. by.

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 pump 524
Since the power loss caused by this can be reduced, fuel efficiency can be improved.

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 flowchart shown in FIG.

After the throttle opening θ is read by the throttle sensor 904 (step 921), the shift lever position is determined by the shift lever switch 901 (step 922). As a result of the determination, if the shift lever is in the P position or N position, the solenoid valves 84 and 85 are activated by the P position or N position processing subroutine shown in FIG.
0FFL (step 931), P or N state is 1? Store it in AM914 (step 932)
.

As a result, the human cub-931 is brought into a nietral state. When the shift lever changes from the P position or N position to the R position, and from the N position to the D position, a shift shock is applied to alleviate the shift shock associated with the N-R shift and N-D shift, respectively. Control processing is performed (steps 940 and 950).

This shift track control process will be described in detail below.

First, the shift control mechanism 70 includes the electric control circuit 90 described above.
By the action of the solenoid valve 74 controlled by the output of It has the function of keeping the upper limit of the hydraulic pressure supplied to the hydraulic servos 48 and 49 below a set value, and limits the engagement pressure of the clutch and brake.

In this embodiment, as shown in FIG. 14, the pressure receiving area of the land provided on the spool 712 of the shift control valve 71 is
S, , S, , S, , Sz, spring 7 in order from the left side of the figure
11 is F and the oil pressure of the oil chamber 713 is P, the hydraulic servo 4 of the multi-disc clutch 45 that is engaged when moving forward
The hydraulic pressure PC supplied to 9 and the hydraulic pressure P supplied to the hydraulic servo 48 of the multi-disc brake 42 that is engaged during reversing are the 0th equation and the 0th equation, respectively, which are hydraulic balance equations of the shift control valve 71.
From the equation, it is given as follows. .

When moving forward pg・s+ = pc, st+ptt ”・■P
c= (Sl /5t)Ps (Fs+/St)P when moving backward, ・Sl =Pb (Sl St)+Fi+
■P, -(Sl/(Sl-sz)) ps(F s+/
(S + Sz)) Also, the pressure receiving area of the valve body 731 inserted in the pressure limiting valve 73 is S
l. If the elastic force of the spring 732 provided in front of the valve body 731 is F,2, then the pressure limiting valve 73
is the highest pressure pj of Ps according to the hydraulic equilibrium equation #2! 1m1t
It operates with.

P Ii + mit XS 3 = P sz
■P 11m1t ”P sz/S
3 At this time, PC and P are the highest pressures Pc 1iaIit , Pb li+mit according to formulas II and 0.
is limited.

When moving forward P c 11m1t = (S + / S t) P
li+wit-F sr/S z ■P when moving backward, 11m1t = (S +/(S I-S z
) ) P 11m1t-(ps+/(S I-5
t)) ■The solenoid valve 74 generates a solenoid pressure P in the oil chamber 713 according to the duty (%) given by the following equation.

Duty = (Solenoid ON time in one cycle/Solenoid operating period) x 100 (%) This duty control is calculated such that during the pulse at IJillJIK'' shown in Fig. 15, L''-nM'' (n = 1, 2.3, ・・・・
), and is achieved by applying gradually decreasing pulses to the shift control solenoid valve 74 shown in FIG. By controlling the shift control solenoid valve 74 in this way,
A hydraulic pressure P adjusted according to the duty is generated in the oil chamber 713 of the shift control valve 71.

The solenoid pressure P shown in FIG. 17 is the shift control valve 71.
The hydraulic pressure Pc or P supplied to the hydraulic servo 48 or 49 shown in FIG.

is obtained.

When mitigating locking shock during N-D shift and N-R shift, hydraulic servo 48 or 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.
The engagement of the multi-disc clutch 45 or the multi-disc brake 42 between C and C is completed. In this way, hydraulic servo 48 or 4
Shift shock control processing 940 that controls the solenoid valve 74 for controlling the oil pressure supplied to
．． A program flowchart of 950 is shown in FIG.

Figure 19 shows each parameter K of the waveform diagram shown in Figure 15.
, L'', and M'' are shown. In step 941, it is determined whether or not the PLUG during shock control processing is on, and the
If LUG is on, shift shock control processing is in progress and the process advances to step 946; if PLUG is not on, the shift lever position stored in RAM 914 and the current shift lever position are used to start shift shock control processing. By comparing the values, it is possible to determine whether the shift lever has changed from the P or N position to the R position.
- is determined (step 942) and whether there is a change from the N position to the D position is determined (step 943). If any changes have occurred, step 944.94
5, the corresponding parameters K", L"
, M" and set the parameter K to 0,
P indicates that shock control processing is in progress.
Turn on the LUG (step 955). If no change has occurred, the process returns and no shock control processing is performed.

In step 946, it is determined whether the parameter K that determines the end of the period K'' is greater than 0, and if K is not thicker than O, then K is set to K''=1, and then set to L.
In step 947), it is determined whether L is less than or equal to 0, and if ■ is less than or equal to O, the process proceeds to step 951, and if L is less than or equal to 0, the process proceeds to step 951. , assumes that all shock control processing has been completed, and turns off the PLUG.In step 946, the PLUG is
When the parameter K, which determines the end of the ON time, is greater than 0, K-1 is set to K (step 950), and then it is determined whether the parameter L, which determines the end of the on time in one period K, is 0 (step 950). 951). When L is O, a command to turn off the solenoid valve 74 is issued and the process proceeds to step 95.
2) If L is not 0, issue an on command and step 9
53) After that, set L-1 to L and return.

Similar shift shock control processing can also be performed using the programmable timer 920 shown in FIG.

Next, referring back to FIG. 12, shift control, which is a feature of the present invention, will be explained.

Next to the N-D shift shock control processing 950, the input pulley rotation speed sensor 902 reads the actual input pulley rotation speed N (step 923), and then the throttle opening θ read in step 921 is set to O.
It is determined whether or not (step 924), and when θ≠O, a subroutine 960 is executed to set the input pulley target rotation speed N" to the input pulley rotation speed for the best fuel efficiency, and
= 0 and the throttle is fully closed, in order to determine the necessity of engine braking, it is determined whether the shift lever is set to the D position or the L position (
Step 926) When the shift lever is set to the D position, a D position engine brake processing subroutine 970 is executed, and when the shift lever is set to the L position, an L position engine brake processing subroutine 980 is executed. Then, set the input end pulley target rotation speed N° to a value suitable for each.

A subroutine 960 for setting the input pulley target rotation speed N0 described above to the input pulley rotation speed for the best fuel efficiency will be explained.

Generally, in order to operate the engine at the best fuel efficiency,
It is preferable to operate according to the best fuel efficiency power line shown by the broken line in FIG. In this figure 20, the horizontal axis is the engine speed (
rpm), and the vertical axis is the engine output shaft torque (kg-rn
), and the best fuel efficiency power line can be obtained as follows. That is, from the constant fuel consumption rate curve (in g/ps-h) of the engine shown by the solid line in Fig. 20 and the constant force curve (in ps) shown by the two-dot chain line, point A in the figure can be found. The fuel consumption rate Q (g/p s - h) and the horsepower P (ps>
Then, at point A, 5=QXP (g/h) of fuel will be consumed per hour. By determining the fuel consumption of 51 s per hour at all points on each equal horsepower curve, the point where S is the minimum on each equal horsepower curve can be determined,
By connecting these points, a best fuel efficiency power line indicating the engine operating condition that provides the best fuel efficiency for each horsepower 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 throttle opening θ shown in FIG.
From the engine output performance curve shown in FIG. 22, the fluid coupling performance curve shown in FIG. The output line can be found. FIG. 25 shows the best fuel efficiency fluid coupling output line shown in FIG. 24 replaced by the relationship between throttle opening and fluid coupling output rotation speed. In the continuously variable transmission device of this embodiment, this fluid coupling output rotation speed directly becomes the input pulley rotation speed Nm.

For this purpose, the input pulley target rotation speed N0 shown in FIG.
In the subroutine 960 for setting the best fuel efficiency input pulley rotation speed N, the best fuel efficiency input pulley rotation speed N1 corresponding to the throttle opening degree θ shown in FIG. Step 961) Set the data address, read the best fuel efficiency input pulley rotation speed N from the set address Step 962), 3! The data of the best fuel consumption input pulley rotation speed Nl corresponding to the calculated throttle opening degree θ is set as the input pulley target rotation speed N1 (step 963).

Next, the engine brake processing subroutine 9 in FIG.
70.980 will be explained.

The engine brake processing subroutine 970 at position D is as follows:
As shown in FIG. 27, the vehicle speed V is detected by the vehicle speed sensor 903.
(971), calculate the acceleration α at that point (972), and then calculate the acceleration α, which is an appropriate acceleration A for the vehicle speed.
It is determined whether or not it is (973). When acceleration α is greater than acceleration A, input pulley target rotation speed N1 is changed to current input pulley rotation speed N1 in order to downshift.
Please set it to a larger value (974), so that the acceleration α is the acceleration A.
If it is not greater than the input pulley target rotation speed N°
is set to the best fuel consumption input end pulley rotation speed N corresponding to the throttle opening degree θ.975) Return. The relationship between vehicle speed and appropriate acceleration variable is determined in advance through experiments or calculations for each vehicle, as shown in FIG.

The engine brake processing subroutine 970 for the L position is as follows:
As shown in FIG. 29, the heavy speed V is detected by the vehicle speed sensor 903.
is read (981), and then the current torque ratio T is calculated from the vehicle speed V and the input pulley rotation speed N using the following formula (9
82).

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, it is determined whether the current torque ratio T is smaller than the torque ratio T1 that provides safe and appropriate engine braking for the vehicle speed.
When is smaller than the torque ratio T4, the input end pulley target rotation speed N1 is set to a value larger than the current input end pulley rotation speed N in order to downshift (984), and the process returns. If the torque ratio T is not smaller than the torque ratio T6, the input pulley target rotation speed N1 is set to the current input pulley rotation speed N (985), and the process returns. The torque ratio T1 that provides safe and appropriate engine braking for the vehicle speed is determined in advance through experiments or calculations for each vehicle, as shown in FIG.

When the input pulley target rotation speed N0 is set as shown in L, the actual input pulley rotation speed N and the best fuel efficiency input pulley rotation speed N4 are compared in FIG. 12 (step 927), when N<Nd-, an operation command for the downshift solenoid valve 84 is issued and the process proceeds to step 9.
28), NUN", an operation command is issued to the upshift solenoid valve 85, and step 929), N=N"
In this case, an OFF command is issued for both solenoid valves 84 and 85 (step 920).

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 to optimize the actual input pulley rotation speed.

It is made to match the fuel consumption input side pulley rotation.

(During constant torque ratio running) As shown in FIG. 31(A), solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned off. 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. The spool 812 is moved to the left in the figure because of the pressing force P exerted by the spring 811. The spool 812 is moved to the left and the oil chamber 81
When 5 and the drain 1813 communicate with each other, the pressure P2 is exhausted, so the spool 812 is moved to the right in the figure by the oil pressure P in the oil chamber 816. When the spool 812 is moved to the right, the drain boat 813 is closed. In this case, by providing a flat notch 812b on the land edge of the drain boat 813 and the spool 812, the spool 812 can be held in a more stable state as shown in FIG. 31(A).
In this state, the oil passage 2 is closed, and the oil pressure of the hydraulic servo 313 of the input pulley 31 is transferred to the output side boolean IJ32.
Due to the line pressure applied to the hydraulic servo 323, it is compressed via the belt 33, and as a result, the oil pressure of the hydraulic servo 323 is balanced. Actually, since there is oil leakage in the oil passage 2 as well, the input end pulley 31 is gradually expanded and the torque ratio 'r changes in the direction of increasing. Therefore, as shown in FIG. 31(A), the spool 812
When the drain boat 814 is in equilibrium, the spool 812 is closed so that the oil passage 1 is slightly open.
A flat notch 812a is provided at the land edge of the oil passage 2 to compensate for oil leakage in the oil passage 2.

(Upshift! - time) As shown in FIG. 31(B), the solenoid valve 85 is turned on by the output of the electric control circuit. As a result, the pressure in the oil chamber 816 is evacuated, so the spool 812 is moved to the left in the drawing, and as the spool 8!2 moves, the pressure in the oil chamber 815 is also evacuated from the drain boat 8!3, but the action of the spring 8110 The spool 812 is set at the left end in the figure.

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

(During downshift) As shown in FIG. 31(C), the solenoid valve 84 is turned on by the output of the electric control circuit, and 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 end 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 output hydraulic pressure of the torque ratio control valve 81 is supplied to the hydraulic servo 3 (3) of the input end (drive side) pulley 31, and the hydraulic servo 323 of the output side (driven side) pulley 32 is supplied with a line. The hydraulic pressure of the input side hydraulic servo 313 is P8, and the hydraulic pressure of the output side hydraulic servo 323 is P. Po/P has a characteristic as shown in the graph of FIG. 32 with respect to the torque ratio T. For example, throttle opening θ...50%, torque ratio T=1.5 (a in the figure)
While driving at point), release the accelerator and set θ-30.
%, when Po/Pi is maintained as it is, it shifts to the point in the figure where the torque ratio T = 0.87, and conversely, when the torque ratio T = 1.5 is maintained, the input terminal The values of P and /P are increased by the output of the torque ratio control mechanism 80 that controls the pulley and changed to the value of point C in the figure. In this way, P,
By controlling the value of /P as necessary, it is possible to set an arbitrary torque ratio corresponding to all load conditions.

[Brief explanation of the drawing]

Fig. 1 is a schematic diagram of a V-belt type continuously variable transmission for vehicles, Fig. 2 is a diagram showing one embodiment of a hydraulic control circuit in the continuously variable transmission of the present invention, and Fig. 3 explains the operation of a manual valve. 4 is a diagram for explaining the operation of the detent valve and the throttle valve. FIG. 5 is a diagram for explaining the operation of the torque ratio valve. FIG. 6 is a diagram for explaining one embodiment of the present invention. 7 is a diagram showing the required line pressure characteristics of the hydraulic control circuit, FIG. 8 is a diagram showing the throttle pressure characteristics, and FIGS. 9, 10, and 11 are diagrams showing the characteristics of the present invention. FIG. 12 and FIG. 13 are diagrams for explaining the flow of processing in the electric control circuit, and FIG. 14 is a diagram for explaining the operation of the shift control mechanism. , FIG. 15 is a waveform diagram of the control pulse, and FIG. 16 is a waveform diagram of the control pulse.
Figure 17 shows the characteristics of the hydraulic pressure supplied to the input and output side hydraulic servos, Figure 17 shows the characteristics of the solenoid pressure, Figure 18 shows the characteristics of the output oil pressure of the shift control valve,
Figure 9 is a diagram for explaining shift thin control processing, Figure 20 is a diagram showing the best fuel efficiency power line of the engine,
Figure 21 is a diagram showing the characteristics of engine output performance, Figure 22
Figure 23 shows the performance curve of the fluid transmission mechanism, Figure 23 shows the equal fuel consumption rate curve of the engine, Figure 24 shows the best fuel efficiency fluid coupling output curve, and Figure 25 shows the best fuel efficiency fluid coupling. Figures 26, 27, and 29 are diagrams showing the characteristics of the output rotation speed, and Figures 28 and 30 are diagrams for explaining the flow of processing in the electric control circuit.
FIG. 31 is a diagram for explaining the operation of the torque ratio control device; FIG. 32 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo; FIG. 33 is a schematic diagram of a conventional continuously variable transmission. 30...V belt type continuously variable transmission, 31...Input side pulley, 32...Output side pulley, 33...V belt,
60... Hydraulic adjustment device, 80... Torque ratio control device, 311.321... Fixed flange, 312.322
...Movable flange, 313.323...Hydraulic servo. Applicant Aisin AK Daplu Co., Ltd. Figure 3 Figure 1 Figure 4 Figure (A) (Snake) Figure 5 (Snake) Figure 6 Figure Throttle, Iθ Figure (Ni9/Cm2) υ ja (maJ) (ry+yn) Input side 7 - tL amount and Fig. 9 Fig. 12 Throttle wholesaler O Straw 13 Fig. 15 Fig. 16 Fig. 4 Yasu 1 Darkness Pd or Pb (kg/cm") Figure 19 s Season-F-%8 Figure 30 Torque ratio, T

Claims (4)

[Claims] (1) Input and output pulleys with variable effective diameters each having a fixed flange, a movable flange, and a hydraulic servo installed on the movable flanges, each attached to the input shaft and output shaft, and a pulley stretched between these pulleys. The clamping force of the drive belt is generated by supplying hydraulic pressure to the hydraulic servo on either one of the input side or output side pulleys, and the clamping force is generated on the hydraulic servo on the other of the input side or output side pulleys. A hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission that controls the torque ratio between input and output shafts by hydraulically adjusting the effective diameter of both pulleys by selectively supplying hydraulic pressure, comprising: a hydraulic pressure source; A regulator valve that adjusts the pressure of hydraulic fluid from the hydraulic source to a line pressure that is supplied to the one hydraulic servo, and a control means that controls the line pressure to increase in accordance with an increase in the torque ratio between the input and output shafts. A hydraulic V-belt type continuously variable transmission for a vehicle, characterized in that the control means generates hydraulic pressure according to a torque ratio and includes means for securing the minimum necessary belt clamping force at a predetermined torque ratio. Adjustment device. (2) The control means is a torque ratio valve that generates oil pressure according to the torque ratio between the input and output shafts, and the regulator valve inputs the output oil pressure of the torque ratio valve to generate a torque between the input and output shafts. The hydraulic pressure adjustment device for a V-belt type continuously variable transmission for a vehicle according to claim 1, characterized in that the line pressure is output in accordance with the ratio. (3) The torque ratio valve has a plurality of oil supply passages to which oil pressure is supplied according to the torque ratio between the input and output shafts, and the oil supply passages are connected to the oil chamber on the pressure increasing side of the regulator valve, 3. The hydraulic pressure adjustment device for a V-belt type continuously variable transmission for a vehicle according to claim 2, wherein the line pressure is controlled according to the torque ratio between the input and output shafts. (4) The torque ratio valve operates to detect the amount of displacement of the movable pulley of the pulley and supply line pressure to the supply oil passage in accordance with the amount of displacement. The hydraulic pressure adjustment device for a vehicle V-belt type continuously variable transmission according to item 2.