JPH0378508B2 - - Google Patents

Description

[Detailed description of the invention]

[Industrial Application Field] The present invention relates to a shift control device for a V-belt continuously variable transmission for a vehicle to prevent a shock when shifting from N range to D range or R range. [Prior Art] A V-belt continuously variable transmission uses an endless metal V-belt invented by Van Doorne wrapped around two pulleys. The band is characterized in that a large number of V-shaped blocks are arranged without interruption from one another, and power is transmitted between the side surfaces of the pulley and the V-shaped blocks and by the pressing force between the V-shaped blocks. Conventionally, various proposals have been made for using this endless metal V-belt as a continuously variable transmission. The challenges for a continuously variable transmission are a means to accurately change the torque ratio according to the driving condition, a hydraulic control device to control power transmission according to the torque ratio, and a way to change the range from N to D.
There is a forward/reverse switching mechanism that prevents a shock when shifting to range or R range, but no effective proposal has been made yet. However, as a prior invention, British Patent No. 2058250 (Borghu Warner Limited, September 7, 1978)
(Application filed on April 8, 1982, Publication issued on April 8, 1982) is known. This consists of a torque converter connected to the output shaft of the engine, a continuously variable transmission mechanism with a fixed pulley and a movable pulley on the input and output shafts, and a V-belt wrapped between the input and output shafts, and a planetary gear. It has a forward/reverse switching mechanism using a unit and a reduction gear mechanism connected to a differential gear device, and a hydraulic control device moves the movable pulley to change the torque ratio, and also changes the torque ratio of the planetary gear unit. advance by engaging or disengaging a predetermined rotating element;
It controls reversing. [Problem to be Solved by the Invention] By the way, in the torque ratio change and forward/reverse switching control described above, this type of continuously variable transmission differs from a general automatic transmission in that the continuously variable transmission This is done with low frictional resistance between the metal pulley and the metal V-shaped block, which requires high differential hydraulic pressure.
Further, even when the engine is wide ring (throttle opening degree is 0), relatively high oil pressure is required to prevent the V-belt from slipping during engine braking. On the other hand, in the forward/reverse switching mechanism, a clutch or brake consisting of a large number of friction plates engages a predetermined rotating element of the planetary gear unit, as in a general automatic transmission. Therefore, in the prior patent, the high hydraulic pressure for changing the torque ratio is used as it is for the latch of the forward/reverse switching mechanism and the differential hydraulic pressure of the brake, so the range changes from the neutral range (N range) to the travel range ( When shifting to D range or R range),
The problem is that the clutch or brake engages too quickly, creating a shift shock. The present invention solves the above problems, and includes:
In addition to ensuring the working oil pressure of the hydraulic servo device to vary the effective diameters of the input and output pulleys of the V-belt type continuously variable transmission for vehicles, the friction engagement device of the forward/reverse switching mechanism has a level control system. It is an object of the present invention to provide a speed change control device that can gradually supply hydraulic pressure while making adjustments and can prevent a shift shock that occurs when shifting from a neutral range to a driving range. [Means for Solving the Problem] In order to achieve the above object, the present invention provides an input pulley, an output pulley, a V-belt stretched between these two pulleys, and an input pulley and an output pulley. a V-belt type continuously variable transmission having a hydraulic servo device for varying the effective diameter of the V-belt type continuously variable transmission; a planetary gear unit; and a forward/reverse switching mechanism having a friction engagement device for engaging or disengaging a predetermined rotating element thereof. and a hydraulic control device that supplies the hydraulic pressure supplied from a hydraulic source to the hydraulic servo device as hydraulic pressure adjusted to a required predetermined pressure. The hydraulic control device includes a manual valve that supplies operating pressure to a predetermined frictional engagement device in response to range switching, and a manual valve that adjusts the hydraulic pressure adjusted to a predetermined pressure by the hydraulic pressure adjustment device to a lower level. a shift control valve that supplies operating pressure to the friction engagement device; and a solenoid that controls hydraulic pressure that is applied as back pressure to the shift control valve to change and control the level of the low operating pressure regulated by the shift control valve. It is characterized by comprising a valve. [Operation and Effects of the Invention] In the V-belt continuously variable transmission for vehicles, when the engine is idling (for example, throttle opening is 0)
In order to prevent the V-belt from slipping, it is necessary to supply high working pressure to the hydraulic servo device that makes the effective diameter of the input and output pulleys variable. A friction engagement that engages or disengages a predetermined rotating element of the planetary gear unit of the forward/reverse switching mechanism via a manual valve when the hydraulic pressure for hydraulic servo operation adjusted to a high pressure is shifted from the neutral range to the travel range. If the operating pressure is supplied to the device, a switching shock will occur during shifting, but in the present invention, the friction engagement device of the forward/reverse switching mechanism has pressure regulated to a lower operating pressure by a shift control valve under the control of a solenoid valve. Since the hydraulic pressure is gradually supplied, it is possible to prevent switching shock from occurring due to the sudden engagement of the frictional engagement device that occurs when shifting from the neutral range to the driving range, thereby improving shift feeling. Not only this, but also the durability of the frictional engagement device can be improved. [Examples] Examples 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 the engine, 102 is the carburetor, 20
is a transmission device provided between the engine 100 and the drive side axle, which includes a fluid coupling 21 connected to the output side 101 of the engine, and a V-belt type continuously variable transmission connected to the fluid coupling 21. 30, a continuously variable transmission comprising a forward/reverse switching planetary gear transmission 40 connected to the output shaft 26 of the continuously variable transmission 30, and a reduction gear mechanism 23 connected to the output shaft 47 of the planetary gear transmission 40. It is made up of. The fluid coupling 21 is well known and consists of a pump impeller 211 connected to the output shaft 101 of the engine and a turbine runner 212 connected to the fluid coupling output shaft 214. Note that other fluid torque converters or mechanical clutches may be used in place of the fluid coupling. The V-belt continuously variable transmission 30 includes an input pulley 31 connected to an output shaft 214 of a 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 includes a fixed flange 311 connected to the output shaft 214, and a fixed flange 311 connected to the output shaft 214.
and a movable flange 312 provided to face the movable flange 312 to form a V-shaped space, the movable flange 3
12 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 30, and a movable flange 322 provided opposite to the fixed flange 321 to form a V-shaped space. , the movable flange 322 is provided so as to be movable in the axial direction by a hydraulic servo 323. The above input pulley 31 and output pulley 32
The amount of variation L of the movable flanges 312 and 322 is 0 to l ₂ to _{l 3} to _{l 4} (0<l ₂ <l ₃ <l ₄ ), which causes torque to be generated between the input shaft 214 and the output shaft 26. ratio T
Continuously variable speed is performed in which the speed changes within the range of _t1 to _t2 to _{t3 to t4} ( _t1 < _t2 < _t3 < _t4 ). 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 forward/reverse switching planetary gear transmission 40 includes a sun gear 41, a ring gear 43, these sun gears 41,
It is composed of a double planetary gear 44 that meshes with a 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, and the carrier 46 is used for forward and reverse switching. planetary gear transmission 40
is connected to an output shaft 47 of. Sun gear 41 and carrier 46 are removably connected by a multi-plate clutch 45, and ring gear 43 is removably connected to a transmission case 400 by a multi-plate 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 is directly changed to the planetary gear transmission for forward/reverse switching. The signal is transmitted to the output shaft 47 of the machine 40 to enable forward running. 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 a V-belt continuously variable transmission 3
This is to compensate for the fact that the shift range obtained with 0 is lower than the shift range achieved by a normal vehicle transmission, and for example, the torque is increased by reducing the speed by a reduction ratio of 1.45. 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 V-belt continuously variable transmission shown in FIG. The hydraulic control circuit includes a hydraulic power source 50 and a hydraulic adjustment device 6.
A shift control mechanism 70 that alleviates shock during 0, N-D, and N-R shifts, and a torque ratio control device 80
Consisting of The oil pressure source 50 is connected to an oil strainer 51 from an oil sump.
Hydraulic oil sucked up by a pump 52 driven by the engine is supplied to a regulator valve 61 through an oil passage 11 to which a relief valve 58 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. A torque ratio valve 66 works in conjunction with the movable flange 322 of the side pulley 32 to supply line pressure to the detent valve 64 according to the amount of displacement thereof, and discharges pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65; It is comprised of a regulator valve 61 that regulates the hydraulic pressure supplied from the source 50 and supplies it to the hydraulic servo 323 as line pressure. As shown in FIG. 3, the manual valve 62 has a spool 621 that is set to R, R, N, and R, corresponding to the shift positions P, R, N, D, and L of the shift lever provided at the driver's seat.
It is set at each position D and L, and connects the oil passage 1 to which line pressure is supplied and the output oil passages 3 to 5 as shown in the table.

[Table] In the table, ○ indicates the connection status with oil channel 1,
× indicates that the 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. The line pressure is output from the output port 616 to the oil passage 1. Oil discharged from port 614 is supplied via oil line 12 to the fluid coupling, oil cooler, and lubrication points. The detent valve 64 is linked to and interlocks with the throttle opening θ of the rear valve of the carburetor 102.
It is equipped with a spool 641 that moves as shown in the figure, and when the throttle opening degree is 0≦θ≦ _θ1 , the detent pressure is connected to the oil passage 5 and the input port 616' provided in the regulator valve 61 as shown in FIG. 4A. When θ ₁ <θ≦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. 4B. The throttle valve 65 includes a spool 651 that is arranged in series with the spool 641 of the detent valve via a spring 645 and has the other spring 652 placed behind it, and the throttle opening is transmitted via the spool 641 and the spring 645. The opening area of the port 653 communicating with the oil passage 1 is adjusted by the action of the spool 651 that moves according to the fluctuation of the degree θ, and the throttle pressure output oil passage 8 communicating with the input port 618 provided in the regulator valve 61 is adjusted. Outputs throttle pressure. Spool 6
51 branches from the oil passage 8, and supplies the output oil pressure to a land 656 and a land 657 having a larger pressure receiving area than the land 656 via output oil pressure feedback oil passages 9 and 10 provided with orifices 654 and 655, respectively. We are receiving feedback from The torque ratio valve 66 includes a spool 662 linked to the movable flange 322 of the output pulley 32 via a connecting rod.
The amount of movement L is l ₃ ≦L≦l ₄ (torque ratio T is t ₂ ≧T≧t ₁ )
In this case, the spool 622 is located on the left side of the figure as shown in FIG. The output oil passage 6 is communicated with the drain port 665 to discharge pressure. The moving amount L of the movable flange 322 is smaller than the first set value _l3 , and _l2 ≦L< _l3 (t3 _≧
When T>t ₂ ), the spool 662 is located at the intermediate portion as shown in FIG. 5B, the port 664 connected to the oil passage 9 and the drain port 666 communicate with each other, and the oil passage 9 is depressurized. Movement amount L of movable flange 322
is smaller than the second set value l ₂ and 0≦L≦l ₂ (t ₄ ≧T
>t ₃ ), the spool 6 is closed as shown in Fig. 5C.
A port 663 connected to the oil passage 1 communicates with the oil passage 6, and line pressure is supplied to the oil passage 6. The shift control mechanism 70 has a spring 71 on one side.
A shift control valve 71 equipped with a spool 712 that receives line pressure from an oil chamber 713 provided at the other end, an orifice 72 provided in the oil passage 1 that supplies line pressure to the oil chamber 713, and the orifice 7
2 and an 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. The solenoid valve 74 is turned on and the drain port 7
41 is opened to discharge pressure from the oil chamber 713, the spool 712 of the shift control valve 71 is connected to the spring 7.
The oil passage 13 is moved to the left in the figure by the action of 11, and is connected to the hydraulic servo 49 that operates the multi-disc clutch 45 of the planetary gear transmission 40. The oil passage 14 is connected to the hydraulic servo 48 that operates the multi-disc brake 42. are connected to drain ports 714 and 715, respectively, to discharge pressure and release the multi-disc clutch 45 or the multi-disc brake 42. When the solenoid valve 74 is off, the drain port 741 is closed, and the spool 712 is located on the right side in the figure due to 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 to engage the multi-disc brake 42 or the multi-disc clutch 45. In this embodiment, the shift control valve 71 includes an oil passage 13 and an oil passage 1.
An oil chamber 717 and an oil chamber 716 are provided to feed back the output hydraulic pressure of the multi-disc clutch 45 and the multi-disc brake 42, thereby alleviating the rise in the output hydraulic pressure and preventing shock when the multi-disc clutch 45 and the multi-disc brake 42 are engaged. The torque ratio control device 80 includes a torque ratio control valve 81, orifices 82 and 83, a downshift solenoid valve 84, and an upshift solenoid valve 85. The torque ratio control valve 81 has a spring 8 on one side.
a spool 812 with a spool 11 installed behind it, 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;
An input port 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 the V-belt continuously variable transmission 3
An oil chamber 819 is provided with an output port 818 that communicates with the hydraulic servo 313 of the input pulley 31 of No. A drain port 813 is provided to evacuate the pressure in the oil chamber 815 according to the movement of the oil chamber 815. The downshift solenoid valve 84 and the upshift solenoid valve 85 are the torque ratio control valve 81, respectively.
is attached to an oil chamber 815 and an oil chamber 816, both of which 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 downshift solenoid valve 8 of the torque ratio control device 80 in the hydraulic control circuit shown in FIG.
4 and an electric control circuit 90 that controls the upshift solenoid valve 85. 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; 905 is a speed detection processing circuit that converts the output of the rotational speed sensor 902 into voltage; and 906 is the converter that converts the output of the vehicle speed sensor 903 into voltage. a vehicle speed detection circuit; 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 for each sensor, 912 is a central processing unit (CPU), and 913 is a read-on memory (ROM) that stores programs for controlling the solenoid valves 74, 84, and 85 and data necessary for control. , 914 is a random access memory (RAM) for temporarily storing input data and parameters necessary for control, 915 is a clock, 916 is an output interface, 917
is a solenoid output driver, which converts the output of the output interface 916 into operating outputs of the upshift solenoid valve 85, the downshift solenoid valve 84, and the shift control solenoid valve 74. Input interface 908-9
11 and CPU912, ROM913, RAM914,
Communication with the output interface 916 is via a data bus 918 and an address bus 919. Next, torque ratio valve 66, detent valve 6
4. The operation of the hydraulic pressure adjusting device 60 of this embodiment, which is composed of the throttle valve 65, the manual valve 62, and the regulator valve 61, will be explained. The hydraulic oil supplied to the hydraulic control path 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 for the vehicle to run with low fuel consumption, it is necessary to bring the line pressure supplied to the hydraulic control circuit close to the necessary minimum.
And each hydraulic servo of the output pulley 32 is defined by a hydraulic pressure that allows torque to be transmitted 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 the vehicle is started, 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, 20 It is desirable to set the line pressure to the line pressure as shown by the broken line which is about % larger, and it is desirable to set the line pressure to a higher line pressure characteristic as shown by the dashed line even when the throttle opening θ=0 during engine braking. In this embodiment, the line pressure, which is the output of the regulator valve 61, is adjusted to the shift position of the manual valve 62 (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 the table 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 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 the oil The line pressure supplied to the oil passage 3 is connected to the hydraulic servo 4 of the multi-plate clutch 45 for forward movement through the oil passage 13.
9, and the vehicle becomes ready to move forward. When the torque ratio T is t1 _≦ T≦ _t2 . As shown in FIG. 5A, the torque ratio valve 66 is
Close port 663 connected to oil passage 1, and close oil passage 6.
is communicated with the drain port 665 to exhaust pressure. As a result, no detent pressure is generated in the oil passage 7 regardless of the throttle opening degree θ. Further, in the throttle valve 65, the port 664 of the torque ratio valve 66 that communicates with the oil passage 9 is closed, and the spool 651 receives feedback pressure from the land 656 and the land 657.
Throttle pressure having the characteristic shown in FIG. As a result, the output of the regulator 61 and the line pressure become as shown in area F of FIG. 9 and area E of FIG. 10. When the torque ratio T is _t2 ≦T≦ _t3 . As shown in FIG. 5B, the torque ratio valve 66 closes the port 663 and allows the oil passage 9 to communicate with the drain port 666. 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 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. 8D. Ru. The line pressure at this time has the characteristics shown in area R in FIG. 9 and area G in FIG. When the torque ratio T is _t3 ≦T≦ _t4 . As shown in FIG. 5C, the oil passage 9 is evacuated from the drain port 666, so that the throttle pressure is expressed in FIG. 8D as above. However, since the port 663 opens and the oil passages 1 and 6 communicate with each other, the throttle opening θ is within the range of 0≦θ≦θ ₁ %, and the spool 641 of the detent valve 64
4A, the oil passage 6 is closed by the spool 641, and the oil passage 7 is connected to the manual valve 62 via the oil passage 5.
However, the throttle opening θ is θ ₁
When %<θ≦100%, the spool 641 moves to the right in the figure as shown in FIG. 4B, and the oil passage 6 and the oil passage 7 communicate with each other, so that detent pressure is generated in the oil passage 7. As a result, the line pressure has a characteristic that changes stepwise at θ=θ ₁ %, as shown in area 2 of FIG. 9 and area 1 of FIG. 10. (L position) The oil passage 5 communicates with the oil passage 1 at the manual valve 62. Oil passage 3 and oil passage 4 are equivalent to the D position. When the torque ratio T is t1 _≦ T≦ _t2 . When the throttle opening degree θ is 0≦θ≦θ ₁ %, 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. . When the throttle opening θ is θ ₁ %<θ≦100%, 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. 4B, and no detent pressure is generated. First, the throttle pressure is the same as in the D position. Therefore, the line pressure has the characteristics shown in FIG. When the torque ratio T is _t2 ≦T≦ _t3 . The difference from the above is that in the torque ratio valve 66, the oil passage 9 communicates with the drain port 666 and the pressure is discharged, and the throttle pressure that the throttle valve 65 outputs to the adjustment valve 61 via the oil passage 8 increases. As a result, the line pressure is expressed by a characteristic curve as shown in FIG. When the torque ratio T is _t3 ≦T≦ _t4 . Oil passage 6 and oil passage 1 are connected by the torque ratio valve 66.
and the oil passage 9 is connected to the drain port 666.
It is excluded from. 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 regulator valve 61 inputs the detent pressure and the same throttle pressure as above. outputs the line pressure shown in FIG. (R position) Oil passage 4 and oil passage 5 at manual valve 62
is in communication with the oil passage 1, and the oil passage 3 is depressurized. In the shift control mechanism 70 at this time, if the shift control solenoid 74 is in the OFF state and line pressure is supplied to the oil chamber 713, the spool 71
2 is located on the right side, oil passage 4 and oil passage 1
The line pressure supplied to the oil passage 4 is supplied to the hydraulic servo 48 of the multi-disc brake 42 for reversing through the oil passage 14, and the vehicle enters the reversing state. (P position and N position) Oil passages 3, 4 and 5 in manual valve 62
Since both are exhausted, the regulator valve 61
The line pressure that is the output of is the same as at the D position. In this line pressure adjustment, the manual valve 62
is shifted to the D, N, and P shift positions, the line pressure when the torque ratio T is in the range of t ₃ < T ≤ t ₄ is determined by the throttle opening as shown in the characteristic curve R in Figure 10. The reason for setting θ so low at ₁ % or less is that if the line pressure is set high in operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low, oil will leak from various parts of the hydraulic circuit due to high oil temperature. When there is a large leak, it becomes difficult to maintain line pressure, and the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which can easily cause trouble, so this is designed to prevent this. be. Further, when the manual valve 62 is shifted to the L and R shift positions, the torque ratio T is in the range of t ₁ ≦T≦t ₂ and the throttle opening is The reason why the line pressure is set high under the operating condition where θ is θ ₁ % or less is because relatively high oil pressure is required during engine braking even when the throttle opening is low. In this way, as shown in Fig. 9, by bringing the line pressure close to the minimum necessary oil pressure shown in Fig. 7,
Since the power loss caused by the pump 52 can be reduced, fuel efficiency can be improved. Next, the operations of the shift control mechanism 70 and the torque ratio control device 80 controlled by the electric control circuit 90 explained in FIG. 6 will be explained with reference to the program flowchart in FIG. 12. 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 first
Both solenoid valves 84 and 85 are controlled by the P position or N position processing subroutine shown in Figure 3.
OFF (step 931) and save the P or N state to RAM.
914 (step 932). As a result, a neutral state of the input pulley 31 is obtained. When the shift lever changes from P position or N position to R position, and from N position to D position, N-R shift and N
- Shift shock control processing is performed to alleviate the shift shock associated with the D shift (steps 940 and 950). This shift shock control processing is a feature of the present invention and will be described in detail below. First, the shift control mechanism 70 includes a solenoid valve 7 controlled by the output of the electric control circuit 90 described above.
The action of 4 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49 of the planetary gear transmission 40 to prevent shock during shifting, and the action of the pressure limiting valve 73 adjusts the timing of supplying and discharging hydraulic pressure to the hydraulic servos 48 and 49.
It has the function of keeping the upper limit of the oil pressure supplied to the clutches 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 ₁ , S _{1 , S 1 , S 2 and the elastic force of the spring 711 is defined as S 1 , S 1} , S ₁ , S ₂ in order from the left in the figure. F _S1 , and assuming that the oil pressure in the oil chamber 713 is P _s , the oil pressure P _c is supplied to the hydraulic servo 49 of the multi-disc clutch 45 that is engaged when moving forward, and the hydraulic servo 48 of the multi-disc brake 42 is engaged when moving backward. The supplied hydraulic pressure P _b is given as follows from the equations 1 and 2, which are oil pressure balance equations for the shift control valve 71, respectively. When moving forward P _s × S ₁ = P _c × S ₂ + F _s1 P _c = S ₁ /S ₂ P _s −F _s1 /S ₂ When moving backward P _s × S ₁ = P _b × (S ₁ − S ₂ ) + F _s1 P _b = S ₁ /S ₁ -S ₂ P _s -F _s1 /S ₁ -S ₂ In addition, the pressure receiving area of the valve body 731 inserted in the pressure limiting valve 73 is S ₃ , and the medium 73
The elastic force of the spring 732 installed behind 1 is F _s2
Then, the pressure limiting valve 73 operates at the maximum pressure Plimit of _Ps according to the hydraulically balanced equation. Plimit+S ₃ =F _s2 Plimit+S ₃ =F _s2 /S ₃ At this time, the maximum pressures P _c limit and P _b limit of P _c and P _b are limited according to the formulas 1 and 2. When moving forward, P _c limit = S ₁ / S ₂ Plimit - F _s1 / S ₂ When moving backward, P _b limit = S ₁ / S ₁ - S ₂ P limit - F _s1 / S ₁ - _{S 2} The solenoid valve 74 is given by the following formula. A solenoid pressure _Ps is generated in the oil chamber 713 according to the duty (%). Duty (%) = Solenoid ON time in one cycle / Solenoid operating time x 100 This duty control has ^a pulse width of L ^* -nM ^* (n=
1, 2, 3, . . . ), and the pulse width gradually decreases, is applied to the shift control solenoid valve 74 shown in FIG. By controlling the duty of the shift control solenoid valve 74 in this manner, a hydraulic pressure P _s adjusted according to the duty is generated in the oil chamber 713 of the shift control valve 71. The solenoid pressure P _s shown in FIG. 17 is amplified by the shift control valve 71 to obtain the hydraulic pressure P _c or P _b shown in FIG. 18 to be supplied to the hydraulic servo 48 or 49. When relaxing the engagement shock during N-D shift and N-R shift, the rise of the hydraulic pressure P _b or P _c supplied to the hydraulic servo 48 or hydraulic servo 49 is controlled as shown in the hydraulic characteristic curve shown in Fig. 16. , In the figure, multi-plate clutch 4 between AC
5 or complete the engagement of the multi-disc brake 42. A program flow chart of shift shock control processing 940, 950 for controlling the solenoid valve 74 for controlling the hydraulic pressure supplied to the hydraulic servo 48 or 49 is shown in FIG. FIG. 19 shows a program flowchart when control is performed using the parameters K ^* , L ^* , and M ^* in the waveform diagram shown in FIG. 15. In step 941, it is determined whether or not FLUG is on during shock control processing. If FLUG is on, it means that shift shock control processing is in progress and the process proceeds to step 946.
If FLUG is not on, the shift lever is moved from the P position or It is determined whether there is a change to the position (step 942) and whether there is a change from the N position to the D position (step 943). If any changes occur, step
Each parameter corresponding to 944 and 945
Set K ^* , L ^* , and M ^* , set parameter K to 0, and turn on FLUG, which indicates that shock control processing is performed (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 one period K ^* is greater than 0. If K is not greater than 0, K is set to K ^* -
1. Set L as L ^* and L ^* as L ^* − M ^* (step
947), in step 948 it is determined whether L is less than or equal to 0, and if L is less than or equal to 0, the process proceeds to step 951;
If L is less than 0, it is assumed that all shock control processing has been completed and FLUG is turned off.
If the parameter K, which determines the end of one cycle K ^* , is greater than 0 in step 946, K-1 is set to K (step 950), and then the parameter L, which determines the end of the on time in one cycle K, is 0 or not. A determination is made as to whether or not this is the case (step 951). When L is 0, a command to turn off the solenoid valve 74 is issued (step
952), if L is not 0, an ON command is issued (step 953), then L-1 is set to L and the process returns. Similar shift shock control processing can also be performed using programmable timer 920 shown in FIG. The explanation will be given by returning to FIG. 12. N-D shift shock control processing 950
Next, the input side pulley rotation speed sensor 902
The actual input side pulley rotation speed N is read (step 923), and then it is determined whether the throttle opening degree θ read in step 921 is 0 or not (step 924). If θ≠0, the input side Subroutine 960 is executed to set the target pulley rotation speed N ^* to the pulley rotation speed on the input side for best fuel efficiency, and when θ=0 and the throttle is fully closed, the shift lever is set to the D position in order to determine the necessity of engine braking. If the shift lever is set to the D position, the engine brake processing subroutine 970 for the D position is executed, and the shift lever is set to the L position. If so, the engine brake processing subroutine 980 for the L position is executed, and the input pulley target rotation speed N ^* is set to an appropriate value. A subroutine 960 for setting the input pulley target rotation speed N ^* 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 the engine along the best fuel efficiency power line shown by the broken line in FIG. In FIG. 20, the horizontal axis shows the engine rotation speed (rpm) and the vertical axis shows the torque (Kg·m) of the engine output shaft, and the best fuel efficiency power line is obtained as follows. In other words, from the engine's equal fuel consumption rate curve (in g/ps/h) shown by the solid line in Figure 20 and the equal horsepower curve (in ps) shown by the two-dot chain line, the fuel at point A in the figure can be calculated. Assuming that the consumption rate is Q (g/ps·h) and the horsepower is P (ps), at point A, fuel will be consumed per hour by S=Q×P (g/h). 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. A best power consumption line indicating the engine operating state resulting in consumption is 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 the throttle opening θ shown in FIG. The best fuel efficiency fluid coupling output line can be found on the fluid coupling output performance curve shown in FIG. 24 from the fluid coupling performance curve shown in FIG. 22 and the engine equal fuel consumption rate curve shown in FIG. 23. FIG. 25 shows the best fuel efficiency fluid coupling output line shown in FIG. 24 replaced by the throttle opening and fluid coupling output rotational speed. This fluid coupling output rotation speed directly becomes the input pulley rotation speed N _B in the continuously variable transmission of this embodiment. To this end, in a subroutine 960 shown in FIG. 26 that sets the input pulley target rotation speed N ^* to the input pulley rotation speed N _B for the best fuel economy, the throttle opening θ is preliminarily stored in the ROM 913 as data. Set an address for the best fuel economy input side pulley rotation speed N _B corresponding to the throttle opening θ (step 961), read out the best fuel economy input side pulley rotation speed N _B from the set address (step 962), and The data of the best fuel consumption input pulley rotation speed N _B corresponding to the throttle opening θ thus obtained is set as the input pulley target rotation speed N ^* (step 963). Next, the engine brake processing subroutines 970 and 980 shown in FIG. 12 will be explained. D position engine brake processing subroutine 9
70 is a vehicle speed sensor 90 as shown in FIG.
3, the vehicle speed V is read (971), the acceleration α is calculated at that point (972), and it is then determined whether the acceleration α is an appropriate acceleration A for the vehicle speed (973). When acceleration α is greater than acceleration A,
In order to downshift, input pulley target rotation speed N ^* is set to a value larger than the current input pulley rotation speed N (974), and when acceleration α is not larger than acceleration A, input pulley target rotation speed N ^* is set. Set to the best fuel economy input side pulley rotation speed N _B corresponding to the throttle opening θ (975) and return. The relationship between vehicle speed and appropriate acceleration A is determined in advance by experiment or calculation for each vehicle, as shown in FIG. Engine brake processing subroutine 9 for L position
70 is a vehicle speed sensor 90 as shown in FIG.
3, read the vehicle speed V (981), then read the vehicle speed V
The current torque ratio T is calculated from the input pulley rotation speed N using the following formula (982). T=(N/V)×k k is the reduction gear mechanism 2 inside the transmission
3, 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 T ^* that provides safe and appropriate engine braking for the vehicle speed (983), and if the torque ratio T is smaller than the torque ratio T ^* , a downshift is performed. In order to do so, the input pulley target rotation speed N ^* is set to a value larger than the current input pulley rotation speed N (984), and the process returns. If the torque ratio T is not smaller than the torque ratio T ^* , the input pulley target rotation speed N ^* is set to the current input pulley rotation speed N (985), and the process returns. The torque ratio T ^* 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. As above, input pulley target rotation speed N ^*
Once established, in FIG. 12, the actual input pulley rotation speed N and the best fuel efficiency input pulley rotation speed N ^* are compared (step 927), and N<N ^*.
When N=N*, a command to operate the downshift solenoid valve 84 is issued (step 928), when N>N ^* , a command to operate the upshift solenoid valve 85 is issued (step 929), and when N=N ^* , a command to operate the upshift solenoid valve 85 is issued (step 928). 8
Issue OFF commands of 4 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. Control device 80
This is done by actuating two solenoid valves 84 and 85 provided in the input side, so that the actual input pulley rotation speed matches the input pulley rotation speed for the best fuel efficiency. (During constant torque ratio running) As shown in FIG. 31A, the solenoid valves 84 and 85 controlled by the output of the electric control circuit are
It will be turned off. As a result, the oil pressure P ₁ in the oil chamber 816
is the line pressure, and the oil pressure _P2 in the oil chamber 815 is also 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 drawing because of the pressing force _P3 caused by the spring 811. The spool 812 is moved to the left and the oil chamber 815
When the drain port 813 and the drain port 813 communicate with each other, the pressure P ₂ is exhausted, so the spool 812 is moved to the right in the figure by the oil pressure P ₁ in the oil chamber 816. Spool 8
When 12 is moved to the right, the drain port 813
will be closed. In this case, drain port 813
By providing a flat notch 812b at the land edge of the spool 812 and the spool 812, it becomes possible to more stably hold the spool 812 at an intermediate equilibrium point as shown in FIG. 31A. In this state, the oil passage 2 is closed,
The hydraulic pressure of the hydraulic servo 313 of the input pulley 31 is
The line pressure applied to the hydraulic servo 323 of the output pulley 32 causes the V-belt 33 to meet and be compressed, resulting in equilibrium with the hydraulic pressure of the hydraulic servo 323. Since there is actually 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 FIG. 31A, a flat notch 812a is provided at the land edge with the spool 812 so that the drain port 814 is closed and the oil passage 1 is slightly open when the spool 812 is in equilibrium. , to compensate for oil leakage in the oil passage 2. (During upshift) As shown in FIG. 31B, 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 figure.As the spool 812 moves, the pressure in the oil chamber 815 is also evacuated from the drain port 813, but due to the action of the spring 811, the spool 812
12 is set at the left end in the figure. In this state, the line pressure of oil passage 1 is at port 818.
Hydraulic servo 31
3 increases, the input pulley 31 operates in the closing direction, and the torque ratio T decreases. Therefore, by controlling the ON time of the solenoid valve 85 as necessary, the upshift is performed by reducing the torque ratio by a desired amount. (During downshift) As shown in FIG. 31C, 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 port 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 downshift is performed. As mentioned above, input side (drive side) pulley 3
The output hydraulic pressure of the torque ratio control valve 81 is supplied to the hydraulic servo 313 of No. 1, and line pressure is led to the hydraulic servo 323 of the output side (driven side) pulley 32, which controls the hydraulic pressure of the input side hydraulic servo 313. When P _i is the hydraulic pressure of the output side hydraulic servo 323, P _p /P _i has a characteristic as shown in the graph _of FIG. 32 with respect to the torque ratio T, for example, the throttle opening θ = 50%. , if you loosen the accelerator while driving with torque ratio T = 1.5 (point a in the figure) and set θ = 30%, when P _p /P _i is maintained as it is, the torque ratio T = 0.87 in the figure. Shifts to point b, and conversely the torque ratio T
= 1.5, P _p /P _i is maintained by the output of the torque ratio control mechanism 80 that controls the input pulley.
The value of is increased and changed to the value of point c in the figure. By controlling the value of P _p /P _i as necessary in this manner, it is possible to set an arbitrary torque ratio corresponding to any load condition.

[Brief explanation of drawings]

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 the manual valve. FIG. 4 is a diagram for explaining the operation of the detent valve and 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. FIG. 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. 12 and 13 are diagrams for explaining the flow of processing in the electric control circuit.
The figure is a diagram for explaining the operation of the shift control mechanism.
Figure 15 is a waveform diagram of control pulses, Figure 16 is a diagram showing the characteristics of the oil pressure supplied to the input and output side hydraulic servos, Figure 17 is a diagram showing the characteristics of solenoid pressure, and Figure 18 is a diagram showing shift control. Figure 19 is a diagram showing the characteristics of the output oil pressure of the valve, Figure 19 is a diagram to explain the shift shock control process, Figure 20 is a diagram showing the engine's best fuel consumption force line, and Figure 21 is the diagram showing the characteristics of the engine's output performance. 22 is a diagram showing the performance curve of the fluid transmission mechanism, FIG. 23 is a diagram showing the equal fuel consumption rate curve of the engine, FIG. 24 is a diagram showing the best fuel consumption fluid coupling output curve, and FIG. 25 is a diagram showing the fluid coupling output curve. Figures 26, 27, and 29 are diagrams for explaining the flow of processing in the electric control circuit. Figures 28 and 30 are for control. FIG. 31 is a diagram for explaining the setting data, FIG. 31 is a diagram for explaining the operation of the torque ratio control device, and FIG. 32 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo. 21... Fluid transmission device, 30... Continuously variable transmission mechanism, 33... V-belt, 40... Forward/forward switching mechanism, 42, 45... Friction engagement device, 70... Shift control mechanism, 74... Solenoid valve.

Claims (1)

[Claims] 1. An input pulley, an output pulley, a V-belt stretched between these pulleys, and a hydraulic servo device for varying the effective diameters of the input pulley and output pulley. a V-belt continuously variable transmission having a planetary gear unit, a forward/reverse switching mechanism having a friction engagement device that engages or disengages a predetermined rotating element thereof; and a hydraulic control device that supplies hydraulic pressure adjusted to a predetermined pressure to the hydraulic servo device. a manual valve that supplies working pressure to a predetermined frictional engagement device; and a manual valve that adjusts the hydraulic pressure adjusted to a predetermined pressure by the hydraulic pressure adjustment device to a lower level and supplies it to the frictional engagement device as a working pressure. A vehicle V characterized by comprising: a shift control valve; and a solenoid valve that controls hydraulic pressure applied as back pressure to the shift control valve in order to change and control the level of the low operating pressure regulated by the shift control valve. Speed change control device for belt type continuously variable transmission. 2 The shift control valve includes a first oil chamber to which operating pressure is supplied to the frictional engagement device, a second oil chamber to which back pressure controlled by the solenoid valve is supplied, and the first oil chamber. and a second oil passage communicating with the friction engagement device and a first oil passage to which the oil pressure regulated by the oil pressure adjustment device is supplied by applying the oil pressure supplied to the oil pressure chamber and the second oil chamber facing each other. and a spool that controls communication with the oil passage, and the second oil chamber communicates with the first oil passage via an orifice, and a pressure valve that maintains the oil pressure of the second oil chamber below a predetermined value. The speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 1, characterized in that the device is equipped with a shifting valve. 3. The oil pressure adjustment device includes a regulator valve that increases the oil pressure supplied from the oil pressure source in accordance with an increase in throttle opening and an increase in the torque ratio of the input pulley and the output pulley. A speed change control device for a vehicle V-belt type continuously variable transmission according to claim 1. 4. The solenoid valve is controlled by an electric control circuit that changes the energization time within the unit time, and the electric control circuit gradually changes the energization time of the solenoid valve when the manual valve is switched to the travel range. 2. A speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 1, wherein