JPH05256358A - Speed change control device of continuously variable transmission - Google Patents

Abstract

PURPOSE:Not to impose an excess burden to a belt by controlling a speed change ratio control valve so as to make a speed change ratio to a fixed value when the number of revolution of a detected wheel is below a prescribed value and the speed change ratio is below a prescribed value. CONSTITUTION:When such situation is detected as that the number of revolution Ns of a secondary pulley is below a prescribed value Nso and a speed change ratio R is below a prescribed value Ro at a step S400, a flag EMRBRF by which it is shown at a step S402 that an emergency brake is stepped on is set. The numbers of revolution Np, Ns are held as NPH, NSH at a step S404 in order to memorize this speed change ratio. A solenoid 55 is turned off in order to fix the speed change ratio at a step S406. Thereby, even if a speed change control system changes the speed change ratio, it is disappeared that a belt is moved along the surface of a pulley. Therefore, an excess burden is not imposed onto a belt and the prevention of a durability drop and the generation of a noise can be restrained.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control device for a continuously variable transmission using an endless belt, and more particularly to an improvement of the control device for improving the durability of the belt.

[0002]

2. Description of the Related Art Conventionally, as a control device for a continuously variable transmission configured to transmit an engine output obtained through a free coupling to wheels at a predetermined gear ratio,
For example, a "control device for a continuously variable transmission" disclosed in Japanese Patent Laid-Open No. 63-53356 is known. That is, by appropriately controlling the hydraulic pressure to be applied to the input side pulley (primary pulley) and the hydraulic pressure to be applied to the output side pulley (secondary pulley), the effective diameters of both pulleys are continuously changed and continuously changed. The gear ratio to be obtained is obtained.

In such a V-belt type continuously variable transmission, the gear ratio can be changed linearly and continuously, but on the contrary, it is difficult to fix the gear ratio to a constant value. In this regard, the above-mentioned JP-A-63-533
In No. 56, a gear ratio control valve for variably controlling the gear ratio, a gear ratio fixing valve for sealing the pressure oil in the pulley hydraulic chamber to fix the gear ratio, and switching between these valves for control. And a control solenoid.
Further, in this continuously variable transmission, when the gear ratio fixing valve is switched to the communication cutoff position, the hydraulic pressure in the hydraulic chamber of the pulley is confined, thereby keeping the gear ratio constant and improving the gear feel. It is something to try.

[0004]

On the other hand, in such a continuously variable transmission, the diameter of the pulley around which the belt is wound must be changed in order to perform the gear shift. In FIG. 1, when 600 is the fixed pulley plate and 700 is the movable pulley plate, when the pulley plate 700 moves left and right in FIG. 1, the belt 800 moves up and down while sliding on the wall surfaces of the pulley plates 600 and 700. As a result, the pulley diameter changes.

However, the rotation of the wheels stops when the brakes are suddenly applied. This means that when the vehicle restarts, it is necessary to set the gear ratio suitable for zero vehicle speed to the gear ratio of the input pulley and the output pulley. However, in the case of sudden braking, the gear ratio will not return to zero and the vehicle will stop. Therefore, when restarting, the gear ratio must be increased, that is, the so-called low speed gear must be changed. However, in this case, the vehicle is stopped and the pulley is not rotating,
For example, moving the movable pulley 700 to the left side in FIG. 1 by using hydraulic pressure forces the belt 800 to slide along the wall surfaces of the pulleys 600 and 700, and at this time, it is against the static friction force. Since the belt 800 slides, the load on the belt increases, the durability of the belt decreases, and abnormal noise may occur.

As a conventional technique proposed to solve such a problem, there is, for example, JP-A-61-52457. According to Japanese Patent Laid-Open No. 61-52457, when the foot brake is stepped on and the vehicle is about to stop suddenly, the line speed of the continuously variable transmission is increased to approach the movable pulley around which the endless belt is wound. By increasing the speed, the shift speed of the CVT is increased, and the speed is rapidly shifted to the maximum speed ratio (so-called low speed gear).
However, increasing the shift speed does not change the fact that the belt is overloaded, and therefore, this conventional technique is not suitable for the purpose of protecting the belt and preventing abnormal noise.

The present invention has been made in view of the above-mentioned problems, and provides a control device for a continuously variable transmission that does not apply an excessive load to the belt of the continuously variable transmission.

[0008]

[Means for Solving the Problems] and

The present invention for achieving the above object is a continuously variable transmission in which a V-belt is stretched between a pair of pulleys whose effective radius changes when the supplied hydraulic pressure is made variable. In the control device, the gear ratio control valve for controlling the hydraulic pressure supplied to the pulley according to the gear ratio, the detection means for detecting the rotational speed of the wheel, and the means for detecting the gear ratio are detected. Control means for controlling the gear ratio control valve so that the gear ratio becomes a fixed value when the wheel rotation speed is less than or equal to a first predetermined value and the detected gear ratio is less than or equal to a predetermined value. Is characterized by.

Since the gear ratio is fixed, the diameter of the pulley is not changed even if the metaphorical driving condition requires the change of the pulley (change of the gear ratio), and the belt is forced to move. It will no longer be moved along the wall.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT An embodiment of a control device for a continuously variable transmission according to the present invention will be described below in detail with reference to the accompanying drawings.
In particular, the continuously variable transmission of this embodiment is provided with a gear ratio fixing valve for fixing the gear ratio. Then, sudden braking is applied and the wheel rotation speed (secondary pulley rotation speed N _S ) is equal to or lower than a predetermined value (N _S0 ), and the gear ratio (ratio of N _S to primary pulley rotation speed N _P = N _P / N). _When the state in which _S ) is less than or equal to the predetermined value (R ₀ ) is detected, the gear ratio fixed by the gear ratio fixing valve is set to a fixed value. By analogy, even if the rotation of the secondary pulley stops in the worst case, the gear ratio is fixed,
The belt is not moved along the surface of the pulley, the belt is not overloaded, durability can be prevented from lowering, and noise can be suppressed. The gear ratio is fixed by the gear ratio fixing valve of the chainer until the rotation speed of the secondary pulley rises to a predetermined value (N _S1 ). That is, when the driver tries to start the vehicle again after the vehicle is suddenly braked and stopped, it is not preferable to start with the gear ratio fixed by the driver. However, it is not preferable to release the gear ratio from the fixed state immediately after starting, because it also unnecessarily burdens the belt, so the gear ratio is fixed by increasing the rotation speed of the secondary pulley to a predetermined value (N _S1 ). It continues until it comes. After that, after releasing the operation of the gear ratio fixed valve to make the gear ratio controllable, the gear ratio control is slowly returned from the previously fixed state to the normal control. If the number of revolutions of the secondary pulley has risen to a predetermined value (N _S1 ), the gear ratio should be returned to the gear ratio at that time and the vehicle should start, but by slowly returning it, the burden on the belt is lightened. ing.

According to this embodiment, the location of the gear ratio control valve and the fixed gear ratio valve will be apparent with reference to the description of FIG. Also, control of the fixed transmission ratio valve will be performed in connection with FIGS. 2, 24 and 25. <Structure of Example Device> First, FIG. 3 is a skeleton diagram showing the entire structure of the continuously variable transmission Z, and FIG.
The hydraulic circuit Q of the continuously variable transmission Z shown in FIG. Here, for convenience of description, the overall configuration of the continuously variable transmission Z will be briefly described with reference to FIG.
The configuration of the hydraulic circuit Q including the line pressure control device, which is the subject matter of the present invention, will be described with reference to. Overall Structure of Continuously Variable Transmission Z This continuously variable transmission Z is a continuously variable transmission for a front-wheel drive vehicle, and is a torque converter connected to an output shaft 1 of an engine A.
Basically, there are a motor B, a forward / reverse switching mechanism C, a belt transmission mechanism D, a speed reduction mechanism E and a differential mechanism F. Torque converter B The torque converter B is fixed to one side of a pump cover 7 connected to the engine output shaft 1 and is a pump impeller 3 that rotates integrally with the engine output shaft 1 and the pump impeller 3. So that it faces the pump cover 7
And a turbine runner 4 rotatably provided in a converter front chamber 7a formed inside the rotor, and a stator 5 interposed between the pump impeller 3 and the turbine runner 4 to perform a torque increasing action. ing. Further, the turbine runner 4 is a carrier 15 which is an input member of a forward-reverse switching mechanism C, which will be described later, via the turbine shaft 2, and the stator 5 is connected to a transmission case 19 via a one-way clutch 8 and a stator shaft 9, respectively. It is connected.

Further, the turbine runner 4 and the pump cover 7
A lock-up piston 6 is arranged between and. The lockup piston 6 is slidably attached to the turbine shaft 2 and comes into contact with the pump cover 7 by introducing or discharging hydraulic pressure into the converter / front chamber 7a and the converter / rear chamber 10. A lock-up state integrated with this and a converter state separated from the pump cover 7 are selectively realized. In the lock-up state, the engine output shaft 1 and the turbine shaft 2 are directly connected to each other without fluid, and in the converter state, the engine torque changes from the engine output shaft 1 to the turbine shaft via fluid. It is transmitted to the 2 side. Forward-reverse switching mechanism C The forward-reverse switching mechanism C has a forward drive state in which the rotation of the turbine shaft 2 of the torque converter B is directly transmitted to a belt transmission mechanism D side, which will be described later, and a reverse drive state in which the rotation is transmitted to the belt transmission mechanism D in a reverse rotation state. And are selectively set, and in this embodiment, the forward / reverse switching mechanism C is configured by a double pinion type planetary gear unit. That is, the carrier 15 splined to the turbine shaft 2
Includes a first pinion gear 13 that meshes with the sun gear 12.
And a second pinion gear 14 that meshes with the ring gear 11. The sun gear 12 is splined to a primary shaft 22 of a belt transmission mechanism D, which will be described later.

Further, an FWD / clutch 16 for connecting and disconnecting the ring gear 11 and the carrier 15 is provided between the ring gear 11 and the carrier 15.
Further, between the ring gear 11 and the mesh case 19, there are provided REV / clutch 17 for selectively fixing the ring gear 11 to the mesh case 19. Therefore, in the state where the FWD / clutch 16 is fastened and the REV / clutch 17 is opened, the ring gear 11 and the carrier 15 are integrated and
Since the ring gear 11 can be rotated relative to the mesh case 19, the rotation of the turbine shaft 2 is output as it is in the same direction from the sun gear 12 to the primary shaft 22 side (forward state).

On the other hand, when the FWD / clutch 16 is opened and the REV / clutch 17 is fastened, the ring gear 11 is fixed to the side of the case 19 and the ring gear 11 and the carrier 15 rotate relative to each other. Therefore, the rotation of the turbine shaft 2 is output from the sun gear 12 to the primary shaft 22 side (reverse drive state) in a state where the rotation of the turbine shaft 2 is reversed via the first pinion gear 13 and the second pinion gear 14.

That is, in the forward / reverse switching mechanism C,
The forward / reverse switching is executed by the selective operation of the FWD / clutch 16 and the REV / clutch 17. Belt transmission mechanism D The belt transmission mechanism D is a primary pulley 21 which will be described later and is arranged coaxially on the rear side of the forward-reverse switching mechanism C described above.
The belt 20 is stretched between a secondary pulley 31 and a secondary pulley 31 which will be described later and are arranged in parallel with and separated from the primary pulley 21.

The primary pulley 21 is arranged coaxially with the turbine shaft 2 described above, and one shaft end of the primary pulley 21 is spline-coupled to the sun gear 12 of the forward / reverse switching mechanism C. A fixed conical plate 23 having a predetermined diameter is provided integrally with the primary shaft 22, and a movable conical plate 24 is provided so as to be movable in the axial direction with respect to the primary shaft 22. The conical friction surface of the fixed conical plate 23 and the conical friction surface of the movable conical plate 24 form a belt receiving groove 21a having a substantially V-shaped cross section.

A cylindrical cylinder 25 is fixed to the outer surface 24a of the movable conical plate 24. Further, a piston 26 fixed to the primary shaft 22 side is oil-tightly fitted and inserted into the inner peripheral surface side of the cylinder 25, and the piston 26, the cylinder 25 described above, and the movable conical plate 2 are inserted.
The primary chamber 27 is configured by the three parties of No. 4 and No. 4.
A line pressure is introduced into the primary chamber 27 from a hydraulic circuit Q described later.

The primary pulley 21 is moved relative to the belt 20 by moving the movable conical plate 24 in the axial direction by the hydraulic pressure introduced into the primary chamber 27 to increase or decrease the distance between the fixed conical plate 23 and the fixed conical plate 23. The effective diameter is such that it is adjusted. The secondary pulley 31 is
Basically, it has the same configuration as the above-mentioned primary pulley 21, and the fixed conical plate 33 and the secondary shaft 32 are provided on the secondary shaft 32 that is arranged in parallel with the above-mentioned primary shaft 22 at a distance. The movable conical plates 34 are provided integrally and movably on the secondary shaft 32, respectively. The conical friction surface of the fixed conical plate 33 and the conical friction surface 44a of the movable conical plate 34 facing each other have a belt receiving groove 31a having a substantially V-shaped cross section.
Is configured.

Further, on the outer surface 34b side of the movable conical plate 34, a substantially cut cylindrical cylinder 35 is coaxially fixed. Further, on the inner peripheral surface side of the cylinder 35, a piston 36 fixed to the secondary shaft 32 is oil-tightly fitted in a portion near the axis thereof. The piston 36, the cylinder 35, and the movable conical plate 34 form a secondary chamber 37. Line pressure is introduced into the secondary chamber 37 from the hydraulic circuit Q, as in the primary pulley 21 side.

Like the primary pulley 21, the secondary pulley 31 also has its movable conical plate 34 brought into contact with and separated from the fixed conical plate 33 so that the effective diameter for the belt 20 is adjusted. At this time, the pressure receiving area of the movable conical plate 34 is set to be smaller than that of the movable conical plate 24 of the primary pulley 21.

The deceleration mechanism E and the differential mechanism F have conventionally known structures, and therefore the description of their structures will be omitted. Operation of continuously variable transmission The operation of the continuously variable transmission Z will be briefly described below. The torque transmitted from the engine A via the torque converter B is transmitted to the belt transmission mechanism D in the forward / reverse switching mechanism C with its rotation direction set to the forward direction or the reverse direction.

In the belt transmission mechanism D, when the effective diameter of the primary pulley 21 is adjusted by introducing or discharging the hydraulic oil into the primary chamber 27, the primary pulley 21 is interlocked with the primary pulley 21 via the belt 20. The effective diameter of the secondary pulley 31 thus adjusted is adjusted while following it. The ratio between the effective diameter of the primary pulley 21 and the effective diameter of the secondary pulley 31 determines the gear ratio between the primary shaft 22 and the secondary shaft 32.

The rotation of the secondary shaft 32 is further reduced by the reduction mechanism E and then transmitted to the differential mechanism F.
It is transmitted from the differential mechanism F to the front axle (not shown). Hydraulic Circuit Q The hydraulic circuit Q shown in FIG. 2 includes a converter front chamber 7a for connecting the lock-up piston 6 of the torque converter B in the continuously variable transmission Z and a converter front chamber 7a.
Rear chamber 10, FWD / clutch chamber 16 and REV / clutch chamber 17 of forward / reverse switching mechanism C, and belt transmission mechanism D
Primary chamber 37 for the primary pulley 21 of
It is for supplying a controlled hydraulic pressure to the secondary chamber 27 for the secondary pulley 31.

As a source of the source pressure of the entire hydraulic circuit Q,
An oil pump 40 driven by the engine A is provided. 2 and FIG. 4 which is a block diagram of the control system of the present embodiment, reference is made to each control valve in the hydraulic circuit Q and a duty solenoid or an on / off type solenoid for controlling them. When,
The correspondence with the chamber to which the hydraulic pressure regulated by these control valves is supplied becomes clear.

The main components of the hydraulic circuit are a line pressure adjusting valve 41, a pressure reducing valve 42, a gear ratio control valve 43,
It is composed of a fixed gear ratio valve 44, an oil pressure correction valve 45, a clutch valve 46, a manual valve 47, a converter relief valve 48, a lockup control valve 49 and the like. Further, as is clear from FIGS. 2 and 4, the gear ratio control valve 43 includes the primary duty solenoid 52.
Directly controlled by. The fixed gear ratio valve 44 is directly controlled by an on / off type solenoid 55.
The clutch valve 46 is directly controlled by the clutch duty solenoid 53. The lockup control valve 49 is directly controlled by the on / off type solenoid 54.

The hydraulic oil discharged from the oil pump 40 is first adjusted to a predetermined line pressure by the line pressure adjusting valve 41 and then supplied to the secondary chamber 37 of the secondary pulley 31 via the line 101. The other output of the line pressure adjusting valve 41 is sent to the clutch valve 46 via the line 102. The clutch valve 46 is
The hydraulic pressure in this line 102 is transferred to the duty solenoid 53.
The pressure is adjusted to a predetermined pressure by means of the line 103 and the manual valve 47 and the lockup control valve 4
Send to 9.

The pressure reducing valve 42 reduces the line pressure supplied to the secondary chamber 37 to generate pilot pressures for the pressure correction valve 45, the gear ratio control valve 43, the gear ratio fixed valve 44, and the clutch valve 46. The pilot pressure for controlling the line pressure (secondary pressure) is adjusted by electrically controlling the duty ratio of the duty solenoid 51. That is, the pressure controlled by the solenoid 51 is introduced into the pilot chamber of the correction valve 45, and the correction valve opens and closes according to the pressure. The pressure on the line 104 controlled according to the open / closed state is introduced into the pilot chamber of the line pressure adjusting valve 41 to obtain a desired line pressure. The line pressure adjusting valve 41 may be directly pressure-controlled by a duty solenoid or the like, but by providing the pressure correction valve 45, an appropriate pressure that compensates for oil leak in the hydraulic circuit can be obtained. become.

The gear ratio control valve 43 is controlled by the primary duty solenoid 52. The pressure on the line 106 generated by the gear ratio control valve 43 is sent to the primary chamber 27 via the gear ratio fixed valve 27. The fixed gear ratio valve 27 is controlled by an on / off type solenoid 55. When the solenoid 55 is in the on state, the line 107 going to the primary chamber 27 is connected to the line 106, and in the off state, the line 107 is closed. In other words, by turning off the solenoid 55, the pressure applied to the primary chamber 27 is fixed to the value at the time when the fixed valve 27 is closed regardless of the operation of the gear ratio control valve 43, that is, The gear ratio can be fixed.

The gear ratio control valve 43 is controlled by the primary duty solenoid 52. When the solenoid 52 is on, the hydraulic pressure in the primary chamber 27 is equal to that of the lines 107, 106, 108 and the relief ball 10.
It is drained through 5. That is, no pressure is generated in the primary chamber. On the contrary, when the solenoid is off, the drain passage 108 is closed and, conversely, the line pressure is introduced into the primary chamber 27 through the orifice 109 and the line 106. The two spools contained in the control valve 43 are tapered. Therefore, while the solenoid 52 is not always on or always off, that is, when the solenoid 52 is in the duty state, the valve 43 opens at an opening ratio according to the duty ratio of the solenoid 52. However, since the pressure passes through the orifice 109, the pressure increase in the primary chamber does not become abrupt.

The duty solenoid 53 controls the clutch valve 46. The line pressure controlled by the solenoid 53 is sent to the manual valve 47 and the lockup control valve 49 via the line 103. In the forward state, the line pressure is applied to the FWD clutch chamber 16 through the line 103, the valve 47, and the line 110. On the other hand, the pressure in the REV clutch chamber 17 is released via the line 112.

On the contrary, in reverse, the line pressure is not only sent to the FWD clutch chamber 16 via line 110. That is, as long as the lockup control valve 49 is in the specific lockup state, the line pressure is equal to the line 10
It is also sent to the REV clutch chamber 17 via 3 → line 113 → line 112. Thus, the reverse drive state is obtained.

The lockup control valve 49 is turned on / off.
It is controlled by the solenoid 54. Valve 49 is locked
If it is in operation, contact the converter rear chamber 10.
The following line 116 is relieved via the relief line 115.
It communicates with the leaf valve 48. Lock up
The state is obtained. The above is the hydraulic control in the hydraulic circuit Q.
Is a schematic description of. <Details of line pressure control>Control system connection FIG. 4 shows a control unit 78 for controlling the line pressure.
The input signal and the output signal are shown. Each source
As shown in FIG. 4, the renoid valves 51, 52, 53, etc.
A control unit 78 is connected to each solenoid
51, 52, 53, etc. are driven by this control unit 78
It is controlled. This control unit is shown in FIG.
The shift position (D, 1, 1,
2, R, N, P) shift position from sensor 82 for detecting
Position signal RANGE and rotation speed N of the primary shaft 22_PTo
The rotation speed sensor 83 (not shown in FIG. 3) for detecting
Primary pulley rotation speed signal N_P Of the secondary shaft 32
Number of rotations N_S Rotation speed sensor 8 for detecting (or vehicle speed)
4 (not shown in FIG. 3) secondary pulley rotation speed signal
Issue N_S And the throttle opening TVO of the engine A is detected.
From the opening sensor 85 (not shown in FIG. 3)
Opening signal TVO and engine engine speed N _E Detect
Rotation speed from the rotation speed sensor 86 (not shown in FIG. 3)
Signal N_E And the turbine shaft 2 in the torque converter B
Number of revolutions N_T Turbine speed sensor 87 (Fig.
Turbine speed signal N from (not shown in FIG. 3)_TAnd hydraulic
Oil temperature from sensor 88 that detects the oil temperature of circuit Q
THO and sensor 8 for measuring the secondary chamber pressure
The hydraulic signal P from 9 is input.

The torque transmitted from the engine output to the forward / reverse switching mechanism C via the torque converter B is the converter transmission torque transmitted via the turbine runner 4 and the lockup torque transmitted via the lockup piston 6. It means that there are two clutch transmission torques. Therefore, the turbine torque transmitted to the primary shaft 22 must be determined in consideration of these two torques. This is line 1 by clutch valve 46.
This is the reason why the line pressure on 02 (hereinafter abbreviated as clutch pressure) must be precisely controlled. However, as can be seen from FIG. 2, the line pressure on the line 101 by the line pressure adjusting valve 41 and the clutch pressure by the clutch valve 46 are intimately entwined. Therefore, the entire line pressure and clutch pressure control according to this embodiment will be described below.

5 and 6 are overall block diagrams of the line pressure control according to this embodiment. As shown in these figures, this control outputs the following three duty signals. These signals are the signal TLUP to the duty solenoid 53 for controlling the hydraulic pressure acting on the piston 6 for locking up, that is, the pilot pressure of the control valve 41, and the hydraulic pressure acting on the secondary chamber 37, That is, the signal TSEC to the duty solenoid 51 for controlling the basic line pressure and the duty solenoid 52 for controlling the hydraulic pressure acting on the primary chamber 27, that is, for controlling the pilot pressure of the control valve 44. Signal T
It is PRM.

The gear ratio is determined by the signal TSEC applied to the solenoid 51 and the signal TPRM applied to the solenoid 52.
Determined by In the present embodiment, as will be described in detail later with reference to FIG. 6, a target primary revolution speed PREVT that matches various driving conditions at that time is determined, and feedback control is performed toward the target value. Signal TPRM. In addition, the signal TSEC takes into consideration the effect of the lockup clutch,
Is generated according to the gear ratio based on. Lockup clutch control schematic According to FIG. 5, the generation control of the signal TLUP is as follows. That is, when it is determined by 200 that the current operating state is within the lock range, at 201, the differential pressure signal DP
The LUP is calculated. Note that FIG. 7 shows map characteristics for determining the lock range. That is, if it is determined by the logic shown in FIG. 7 that it is within the lockup range, 201 calculates the lockup clutch differential pressure DPLUP from the lockup transmission torque initial value TQINT calculated from the engine output torque TQPUMP (described later). .. This DPLUP
Is the target value of the hydraulic pressure difference before and after the piston 6.

The DPLUP is subjected to the limit correction at 202 based on the characteristic shown in FIG. This is because excessive hydraulic pressure is not applied. And, further, the clipped DPLUP is corrected at 203, taking into account the line pressure (which is defined by the signal PSEC defining the hydraulic pressure applied to the secondary chamber 37). That is, according to FIG. 8C, the line pressure PS
The maximum value of the clutch pressure is calculated by EC, and the calculated amount is set as the maximum clutch pressure PLUOFF _MAX . Then, as shown in FIG. 8B, the DPLUP is set to the PLUO
While controlling the limit with FF _MAX as the maximum value, DPL
Based on the characteristics of Fig. 8 (b), the clutch pressure PLUO
Convert to FF. The clutch pressure PLUOFF is converted into a duty ratio at 204, and power source voltage correction or the like is performed at 205. Further, at 206, the corrected duty is converted into a cycle and output to the solenoid 53. The details of this duty conversion will be described later. Outline of Primary Hydraulic Control The outline of the hydraulic control of the primary chamber will be described with reference to FIG.

First, at 220, the shift signal RAN
GE, throttle opening TVO, operation mode MODE,
Based on the current secondary pulley rotation speed N _S, etc.,
The target value PREVT of the primary rotation speed is determined. Then, the difference DNP between the target rotation speed PREVT and the current rotation speed N _P of the primary pulley is calculated by 221 and the solenoid 52 is subjected to the duty control while performing the feed back correction and the non-feed back control correction on this DNP. .. Here, the execution condition of the feed back control is determined by 224, and the selector 225 selects the feed back corrected if the execution condition is satisfied, and selects the feed back corrected if not satisfied, and the duty conversion stage 226. Output to. The conversion to the duty ratio and the like are the same as in the case of the lockup. The determination of PREVT in 220 is made based on the characteristics shown in FIG. Abnormal Gear Shift Detection FIG. 6 also illustrates how the solenoid 55 is controlled when a gear shift abnormality is detected. In the present embodiment, the abnormal gear shift means the rotation speed N _{P of the} primary pulley.
Alternatively, it is detected as an abnormality in the target primary rotation speed PREVT. That is, as will be apparent from the following description based on the flow chart of FIG. 23, in the present embodiment, three operating states are set for abnormal gear shifting. One is when the rotation speed NP itself is going to exceed a certain threshold value N _TH , that is, when N _P ≧ N _TH , and the second is the absolute deviation of the primary rotation speed NP from the target primary rotation speed PREVT. If the absolute value of ΔN _P = PREVT-N _P is about to exceed a predetermined threshold Δ _TH (> 0), that is, | ΔN _P | ≧ Δ _TH , the absolute value of the change rate of the deviation, _{i.e., δN P = △ n P (} n) - △ when the absolute value of n _P (n-1) is about to exceed the predetermined threshold value δ _TH (> 0), That is, | δN _P | ≧ δ _TH . This is because the primary revolution speed N _P , the target primary revolution speed PREVT, the deviation ΔN _P , the deviation change speed δN _P, etc. do not change significantly. Here, n represents the amount measured in the current cycle, and n-1 represents the amount measured in the previous cycle.

There are various possible causes of these abnormal gear shifts. For example, when the gear ratio control valve 43 becomes sticky or when the solenoid 52 for controlling the gear ratio control valve 43 is broken, the solenoid 52 is used. The driver circuit (not shown) may be short-circuited, the primary pulley rotation speed sensor may fail (sensor fail signal in FIG. 6), or the like. The hydraulic circuit is designed so that a fail-safe function automatically works against the disconnection of the solenoids 52 and 55. That is, when the solenoid 52 is disconnected, the gear ratio control valve 43 is always in the off position, so that hydraulic pressure is applied to the primary chamber and the secondary
The rotation speed of the pulley is controlled downward. When the solenoid 55 is disconnected, the gear ratio fixing valve 44 is always in the off position, so the pressure in the primary chamber 27 is fixed. That is, the rotation speed of the secondary does not go to the high rotation side. Further, the failure detection itself of the pulley rotation speed sensor can be easily performed by monitoring the magnitude of the value of the output signal and indirectly measuring the impedance of the sensor circuit. When such a state is detected, the sensor fail signal becomes 1. Be done. The target rotational speed PREVT of the returned primary pulley after a sudden stop is determined by two routes as shown in FIG. One is a shift signal RANGE, a throttle opening TVO, and an operation mode M.
PREVT is determined based on ODE, the current rotational speed N _S of the secondary pulley, and the like. For convenience, the determination of PREVT by this route is called “normal control”.

The second route is to determine PREVT based on "return control" 300. In this return control, when the vehicle is suddenly stopped, the gear ratio is fixed according to the control of this embodiment. Therefore, it is an excessive burden on the belt to rapidly return the gear ratio from the fixed state to the "normal control". Because. PRE determined by "normal control"
The selector 234 determines which of the VT and the PREVT determined by the “return control” is selected. The actual operation of the selector is realized by the control procedure of FIG. Schematic of secondary hydraulic control Return to the description of FIGS. 5 and 6.

The hydraulic pressure applied to the secondary chamber 37 (corresponding to the line pressure in this embodiment) is schematically shown in FIG. 5B, and is controlled as follows. First,
At 210, the engine output torque TQPUMP is calculated.
As shown in FIG. 10A, the calculation of TQPUMP is performed with respect to the engine output torque TQENG calculated based on the throttle opening TVO and the engine speed N _E.
Load compensation TQLD and compensation amount T due to oil pump loss
It is a reduction of QOPMP. As the engine load, there is an air conditioner, a power steering, etc., as shown in FIG. 10 (b), and in the present embodiment, correction by the engine water temperature THW is also taken into consideration. Then, the correction of the pump loss is calculated based on the line pressure PSEC to the secondary pulley and the engine speed N _E as shown in (c) of FIG. 10.

Engine torque T taking these corrections into consideration
At 211, the torque TQI input from the forward / reverse switching mechanism C to the primary pulley based on QPUMP.
N is calculated. The calculation of this TQIN will be described in detail later, but the engine output torque TQPUMP is
Torque TQL transmitted through the lockup clutch
Torque TQCVD transmitted through UP and converter
Is accurately separated, and the combined torque transmitted to the primary pulley is accurately calculated. 21
In 2, the force F necessary to press the pulley is input from the input gear ratio RATIO by inputting this TQIN.
Calculate SEC. The calculation of FSEC will be described in detail later with reference to FIG. This pressing force FSEC is
After being corrected for centrifugal force at 213, it is converted to pressure PSECO at 214. Therefore, this PSECO
The output torque output via the torque converter and the lockup mechanism is changed to the primary torque according to the gear ratio RATIO.
For the force FSEC required to properly transmit to the pulley,
It has been corrected by centrifugal force.

On the other hand, at 215, the minimum pressure PSMIN required to operate the gear shifting operation is calculated based on the gear ratio RATIO and the primary pulley rotation speed N _P.
Then, the selector 232 selects the larger value of PSECO and PSMIN. This is a signal for driving the duty solenoid 51 that controls the line pressure. How the centrifugal pressure is generated and the selector 232
Then, the reason why the larger one of PSECO and PSMIN is selected will be described with reference to FIGS. 12 and 13 to 16.

Now, the pulley input torque calculation performed at 211 will be described in detail with reference to FIG. As described above, when the engine output is transmitted to the primary pulley of the belt transmission mechanism D through the torque converter B and the lockup clutch, the torque transmitted through the turbine runner 4 and the lockup torque are transmitted. This is done so that two of the torques transmitted through the piston 6 are taken into consideration. That is, the engine output torque TQENG is divided into two directions, that is, the torque converter and the lockup mechanism. However, these two divided torques have different torque transmission ratios to the primary shaft 22 because their transmission paths are different. Therefore, in order to accurately grasp the torque transmitted to the shaft 22, it is necessary to accurately grasp how the engine output torque TQENG is divided into two.

In the present embodiment, as shown in FIG. 11, as a method of calculating the above-mentioned bisected torque, first, the torque TQLUP transmitted by the lockup is calculated, and this torque TQLUP is subtracted from the engine output torque TQPUMP, Torque transmitted to the torque converter TQCV
I asked for D. The calculation method may be the reverse of the above, that is, the torque transmitted only to the torque converter may be obtained, and then the torque transmitted to the lockup may be obtained.

First, based on the engine output TQPUMP, the initial value DPINT of the differential pressure before and after the lockup piston is calculated according to the characteristic shown in FIG.
The characteristic (a) is uniquely determined according to the lockup mechanism. As described above, DPINT takes into consideration the determination condition of the lockup range (FIG. 7),
It is converted into the above-mentioned DPLUP by 201 of FIG. This differential pressure DPLUP indicates the pressure transmitted to the lockup piston 6. By the way, since the torque actually transmitted to the piston 6 depends on the viscosity of the oil in the converter front chamber 7a (that is, the temperature of the oil), it follows the characteristic shown in FIG. Depending on the oil temperature THO, the lockup transmission ratio H _LU is set so that a lower torque is transmitted to the piston 6 as THO increases.
This transmission ratio H _LU is an amount of 0 to 1 as shown in FIG. Then, from the transmission ratio HLU and the engine torque TQPUMP, as shown in (c), the torque TQLUP actually transmitted to the lockup piston 6 is obtained from TQLUP = TQPUMP × H _LU . Therefore, of the engine output, the torque TQCVD input to the torque converter alone is TQCVD = TQPUMP-TQLUP. Therefore, the torque TQCNVT output from the converter alone is calculated in consideration of the torque ratio TR of (e), and TQCNVT = TQCVD × TR = (TQPUMP-T
QLUP) × TR. Therefore, the combined torque TQTRBN from the lockup and the converter is TQTRBN = TQLUP + TQCNVT. This torque is input to the switching mechanism C and transmitted to the primary pulley. This switching mechanism C
Since the planetary mechanism is mounted on the engine, the final torque value TQIN is calculated from TQTRBN in consideration of the speed reduction ratio KSRDCT by this mechanism.

TQIN = TQTRBN × Reduction Ratio The value of this reduction ratio KSRDCT is shown in (f) of FIG. Now, here, the calculation of the torque ratio TR will be described with reference to (d) and (f) of FIG. 11. The torque ratio is the ratio of the torque transmitted by the converter to the input torque, and is calculated from the speed ratio E. The speed ratio E is the ratio (N _P / N _E ) of the rotational speed N _P of the primary pulley to the engine rotational speed N _E and the speed change range (RANG).
E) and is determined as shown in FIG. 11 (e). In addition, when reversing (R), KLR
If RT is a deceleration constant during reverse movement of the forward / reverse switching mechanism, the speed ratio ER is N _P / N _E / KLRRT. RAN
When GE is in P and N, naturally E is "0".
The torque ratio is determined according to the characteristic such as (d). This characteristic is set to the maximum value "2" when the speed ratio E is zero, that is, when the vehicle is stopped, and gradually decreases from "2" as the speed ratio E gradually increases from zero.
When the speed ratio E reaches about 0.8, the torque ratio TR is set to "1", and thereafter, the torque ratio TR is set to be maintained at "1".

As described above, the pulley input torque calculation performed in 211, that is, the torque TQIN transmitted to the primary pulley shaft 22 was accurately obtained. The details of the secondary pressure PSEC calculation will be described by describing in detail the control performed in steps 212 to 216. First, the calculation of the pressing force performed at 212 in FIG. 5 will be described with reference to FIG. This pressing force is a force that should be applied to the piston 36 so as to keep the effective diameters of the primary pulley and the secondary pulley at required amounts against the tension of the belt. Since the effective diameter changes according to the gear ratio, the pressing force is the gear ratio RAT.
Must change according to IO.

The gear ratio RATIO is calculated as follows according to FIG. First, as shown in FIG. 12B, the current primary pulley rotation speed N _P and the target pulley rotation speed PREVT calculated at 220 are compared in magnitude. This is because in order to obtain the required pressing force, it is necessary to generate the larger pressing force of the pressing force required to maintain the current gear ratio and the pressing force required after the gear shift. If the larger value selected by the selector 231 out of N _P and PREVT is NP, RATIO becomes RATIO = N _P / N _S by the multiplier 230. Then, the gear ratio RATIO and the input torque T
The pressing force FSEC is calculated from QIN as shown in FIG.
Calculate according to the characteristics of. In this case, the larger the gear ratio, that is, the overdrive, the larger the pressing force is required.

The details of the centrifugal hydraulic pressure correction will be described. Details of this correction control are shown in parts (c) to (g) of FIG. 12. First, the reason why centrifugal oil pressure correction is necessary and the problems associated with this correction are illustrated. 13 to 16 will be described. As shown in FIG.
When the primary pulley having an effective area A _P and the secondary pulley having A _S are rotated by the belt, the line pressure P _S commonly applied to these pulleys (the line pressure is defined by the solenoid 51 on the secondary side). , Etc., the pressing forces F _P and F _S against each pulley are: F _P = A _P × P _S + K _P × N _P ² (1) F _S = A _S × P _S + K _S × N _S ² + F _SP (2) In each equation, each K is a predetermined constant,
The second term is the force due to the centrifugal force, and the third term in the second equation is the force due to the spring (38 in FIG. 3). Then, maintaining the ratio H _F of the pressing forces between the pulleys (shown by the following equation) at a predetermined value is a premise for performing an appropriate gear shift. If this value is not set properly, for example,
This is because there is a possibility that a situation may occur in which the effective diameter of the secondary pulley does not change even though the effective diameter of the primary pulley becomes smaller.

H _F = F _P / F _S (3) FIG. 14 is a graph showing the change characteristics of F _P with respect to the primary pulley rotation speed N _P. Similarly, FIG.
Shows that of F _S. As shown in these figures, the pressing force due to the centrifugal force increases as the number of revolutions N _P (N _S ) increases as the square. Since this centrifugal force works in the direction of pressing the pulley, if this amount is subtracted from the line pressure, the load on the hydraulic pump is reduced, which contributes to fuel consumption reduction. Figure 1
What is shown by the broken line 5 is the pressure A _S × P when the line pressure P _S is reduced by the contribution of the pressure due to this centrifugal force.
_This shows the change in _S. By the way, since the area ratio between the pulleys is A _P : A _S ≈2: 1 as described above, when the line pressure P _S is reduced by the centrifugal correction, F _S becomes substantially constant. Even if it can be maintained, the decrease amount of A _P × P _S becomes large and the pressing force ratio H _F
Will not be maintained at an appropriate value. This is because the contribution of the centrifugal force is effective in the square of the rotation speed, but the absolute value itself is small. Therefore, when the centrifugal correction becomes large, the line pressure is significantly reduced, and in such a case, the line pressure is reduced to less than the required minimum amount.・ Because the rotation speed N _{S of the} pulley rises, it becomes large).
That is, if the line pressure P _S is reduced by ΔP due to the centrifugal force correction, the change in the pressing force due to this is A _P × Δ
P, and as a result, F _P is considerably reduced. Therefore, in this embodiment, the decrease in P _S is monitored, and when the pressure falls below the minimum line pressure, the limit correction is performed. Lower limit value PSMIN at this time
Is defined by the following equation obtained by solving the equation (3) for P _{S. PSMIN = {(K S · N} S 2 + F SP) × H F -K P · N P 2} / {A P -A S · H F} ... (4) continued return connexion described in Figure 12. (C) and (d) of FIG.
Thus, the centrifugal force and the spring force are corrected with respect to the required pressing force FSEC. Then, the pressure PSECO applied to the piston 36 of the secondary pulley is calculated from the piston area A _S. That is, PSECO = {FSEC- (K _S N _S ² + F _SP )} / A _S. Then, the selector 232 selects the smaller one of PSECO and PSMIN. And (h) of FIG.
In the above, PSEC is obtained by multiplying the selected one by a safety factor K _SF for ensuring the certainty of the operation.

Thus, the gear ratio, that is, the pulley rotation speed N
Regardless of S, the line pressure PSEC for ensuring the shifting operation could be secured. Next, the control performed on the solenoids 51, 52, 53 will be described. In this solenoid control, in order to reduce the pulsation of hydraulic pressure and ensure the durability of the solenoid, the drive frequency of the duty solenoid is based on the discharge amount of the hydraulic pump (that is, the engine speed N _E ) and the duty ratio. Control it.

FIG. 17 shows the whole of this control. Since the controls for the three solenoids are the same, the duty conversion for the solenoid 52 will be described as a representative of the solenoids. First, in FIG. 17A, the signal CSTRK is converted into a duty. At this time, when the oil temperature THO is high, the output DU
TY is large. In (b), the battery power supply voltage is corrected for this duty. And
Upper and lower limit clip processing is performed in FIG. 17C, and failsafe processing is further performed in FIG. 17D. Where XSF
T1F is a flag for storing the occurrence of a fail state. This fail safe is a safe control for the fail including the above-described abnormal gear shift.

On the other hand, in (e), the mode of the driving cycle of the solenoid is set based on the engine speed N _E and the DUTY obtained in (a). In this embodiment, this periodic mode is
3 of LOW, MIDDLE, and HIGH depending on zone judgment
I made it to the street. Then, in (f), the drive cycle is set to 10.5 ms (LOW), 21.0 based on the zone determination.
ms (MIDDLE) or 31.5 ms (HIGH). As described above, the duty control of the present embodiment controls the drive frequency based on the engine speed N _E and the DUTY value. The frequency control is performed because the pulsation of the pump output easily occurs when the duty ratio is variable while the frequency is constant. The drive frequency is higher as the engine speed N _E is higher, in other words,
The larger the discharge amount from the hydraulic pump 40, the higher the discharge amount. This is because the amplitude of the hydraulic pressure pulsation tends to increase as the discharge amount increases, so that the drive cycle is shortened and the ripple is reduced. <Control Procedure> The above is the line pressure control in the hydraulic control device of the continuously variable transmission including the torque converter having the lockup, and the duty control of the duty solenoid used in the hydraulic circuit. The control described above can be implemented by the control unit 78 whether it is a digital computer or an analog computer. Therefore, next, a control procedure when such control is realized by a digital computer will be described with reference to FIGS. 18 to 25.

18 and 25 show the control procedure of the entire control.
18 is a flow chart, and in particular FIG. 18 shows a secondary pool.
Fig. 25 shows the hydraulic control of the primary pulley.
Will be explained. The control procedures in FIGS. 25 and 18 are independent of each other.
Will be done. But the primary hydraulic control and the second
Dali hydraulic control is the secondary pressure control in Figs.
The pressing force FSEC of the pulley is the primary speed PRE
VT (or N_P) Is calculated through RATIO
As they are, they are related to each other. Secondary control First, referring to FIG. 18, the hydraulic control of the secondary pulley is performed.
I will explain to you. In step S100, the secondary chamber 3
Target value of line pressure in line 101 supplied to 7
Calculate PSEC. This detail is shown in FIG.
It In step S200, the forward / reverse clutch chamber
The target value P of the clutch pressure that the pressure supplied to
_CLTIs calculated. Here, the target value P of the clutch pressure_CLTIs
Set according to the torque input via the Luk converter
Of the line pressure control procedure (FIG. 20) described later.
The value corresponding to TQTRBN (step S122) is set.
To be done. In step S202, PSEC, P_CLTTo
For example, consider the oil temperature THO according to the characteristics shown in FIG.
Considering each, the target duty ratio D_L, D_CConvert to. Su
In step S204, feedback control is performed.
Hydraulic pressure P required for_{OI L}Is read from the sensor 89. That
Then, in step S206, the target values PSEC and P_OILWhen
From the difference of the duty value D_LFeedback system against
Control amount for control △_LIs calculated.

Δ _L = k (PSEC-P _OIL ) Then, in step S208, D _L is corrected according to D _L = D _L −Δ _L. In step S210, the map shown in FIG. 22, fed back control amount for the line pressure △ based on the L and the target clutch pressure P _CLT, calculates the correction amount △ _C for Deyutei D _C of clutch pressure,
In step S212, D _C = D _C - according △ _C, calculates a Deyutei value corresponding to the final clutch pressure. In step S214, these duty values D _L and D _C are output to the solenoids 51 and 53, respectively. The significance of the control in the PSEC calculation steps S204 to S212 will be described later, and the details of the calculation of the target line pressure PSEC performed in step S100 will be described with reference to FIG.
Follow the instructions below.

In FIG. 20, first, step S102.
Then, the shift position signal RANGE is read from the sensor 82. When the shift position is in P or N, the transmission does not operate, so in step S23 the turbine torque TQTR is set.
BN is set to "0", and the process proceeds to step S124. When the shift position is D, 1, 2, or R, step S106
The engine speed N _E and the throttle opening TVO are read from the sensors 86 and 85, respectively. Then, in step S107, the engine torque TQENG is calculated based on N _E, TVO, etc., in accordance with (a) of FIG. In step S108, the load is corrected in accordance with (b) in the figure, and in step S109 the torque cross is corrected by the oil pump to obtain the engine torque output TQPUMP.

Next, at step S110, the oil temperature THO is read from the sensor 88, and the initial value DPINT of the differential pressure of the lockup clutch is calculated based on TQPUMP according to (a) of FIG. This DPINT
As described above, the differential pressure DPLUP is calculated by the control for calculating the duty on the primary side based on the above. Therefore, b look up class Tutsi transmission ratio H _LU from the DPLUP is accordance connexion calculated in (b) of FIG. 11, in step S112, based on these DPLUP and H _LU, Russia look up class Tutsi transmission torque TQLUP is calculated. That is, TQLUP = DPLUP × H _LU . In step S114, the turbine speed NT of the torque converter is read from the sensor 87. Then, in step S116, the speed ratio E is obtained from E = N _T / N _E. Here, N _T is equivalent to N _P (or N _P × KLRRT) in (e) of FIG. Then, in step S118, the torque ratio TR is calculated. They are,
This has been described with reference to (d) and (e) of FIG. Next, in step S120, the converter transmission torque TQCNVT is transmitted.
Is calculated. That is, TQCNVT = (TGPUM-PTQLU) * TR. Next, in step S122, a combined torque TQTRBN of the transmission torque TQLUP via the lockup clutch and the transmission torque TQCNVT via the converter.
Is calculated. That is, TQTRBN = TQLUP + TQCNVT. By correcting the planetary reduction ratio,
TQIN.

In step S124, the sensors 84 and 83 are
To the secondary pulley speed N _SAnd the primary
-Revolution N_PAnd read, and abnormal change occurs at step 125.
Performs speed detection control. Details of the control of this step 125
The details are shown in the flow chart of FIG. 23 and will be described later.
In step S126, the gear ratio RATIO (= N_P/ N_S)
calculate. Here, the selector 230 is the primary pool.
Actual rotation speed N_PAnd target speed PREVT
The larger one is N_PSo the primary pulley
The number of rotations of is reflected in the number of rotations of the secondary pulley.

Then, in step S127, step S
"Sudden braking detection" control is performed based on N _S obtained at 124. This sudden braking detection control will be described later in detail. Then, in step S128, TQIN and RATI
From O and the force FSEC required to press the pulley is calculated. This detailed description has been described with reference to FIG. Then, in steps S130 and S32, as shown in FIG.
According to 2 (c) and (d), centrifugal force correction and spring correction are performed to obtain the line pressure PSECO. This PSECO becomes the piston pressure on the secondary pulley side.

Since this PSECO may drop below the minimum line pressure due to centrifugal force correction, the minimum line pressure PSM according to the gear ratio and the like at that time may be reduced by the following procedure.
Calculate IN. That is, in step S136, the pulley pressing force ratio HF is calculated according to the above equation (3). Then, in step S138, according to the equation (4),
Calculate the minimum line pressure PSMIN required for shifting. Next, in step S140, the larger one of PSECO and PSMIN is selected. In step S142, FIG.
The correction by the safety factor KSF is performed in accordance with (h).

PSECO and P obtained as described above
The larger SMIN is the target line pressure PSEC in step S100. Then, this PSEC is converted to a duty ratio D _L in step S202,
Further, in steps S204 to S208, the feed back control is performed based on the actual line pressure P _OIL as described above.

FIG. 21 shows the target line pressure PSEC.
Duty ratio D_LHowever, due to feedback control, △_LIs
The figure shows how the correction is performed. By the way, this △
_LIs the length of the line in the hydraulic circuit Q and the line resistance.
And a solenoid 51 for controlling the line pressure adjusting valve 41
When the characteristics of are decided, they are decided by themselves. And Ku
The clutch pressure control in the latch valve 46 also controls the line pressure.
The operation is performed with the hydraulic pressure from the adjusting valve 41 down. Paraphrase
Line pressure PSEC and clutch pressure P_CLTClosely with
Are related to each other, and that relationship can be known in advance.
Wear. That is, the above △ _CIs closely related to ΔL above.

[0063] Figure 22, △ _C, and explaining the state obtained based on the fed back control amount △ _L and the target clutch pressure P _CLT and oil temperature THO. In FIG. 22, the broken line for obtaining Δ _C is the feedback control amount Δ _L and the target clutch pressure P _CLT.
And the oil temperature THO. Thus, the clutch pressure is precisely feedforward controlled with substantially the same precision as the feedback control based on the sensor 89 measuring the line pressure supplied to the secondary chamber. In other words, only one hydraulic sensor is needed. Safe Control for Abnormal Gear Shift The details of the "abnormal gear shift detection" in step S125 of FIG. 20 will be described with reference to FIG.

When the pulley rotation speeds NS and NP are obtained by the sensors in step S124 of FIG. 20, step S298 or step S300 of FIG. 23 is executed. Step S298 is for checking whether or not a defect has been detected in the sensor for detecting the number of rotations of the two pulleys. Step S300 is a flag XSFT1F for storing the fail state.
Check the set state of. If the sensor has no fail and no fail has occurred in the past, step S3
Go to 02. Steps S302 to S312 are procedures for detecting the above-described three abnormal gear shift occurrence states.

In step S302, the abnormal rotation of the rotation speed NP is detected. That is, if N _P ≧ N _TH , a failure occurs. In step S304,
Target primary speed PREV of primary speed N _P
The deviation from T, ΔN _P = PREVT−N _P, is calculated, and if the absolute value is determined to be | ΔN _P | ≧ Δ _TH in step S306, it is determined that an abnormal gear shift has occurred.
Further, the rate of change of the step S308 the deviation _{△ NP δN P = △ N P} (n) - △ N to calculate the _P (n-1). Step S310 prepares for the next detection,
The current ΔN _P (n) is saved as ΔN _P (n-1). Then, in step S312, | δN _P | ≧ δ _TH is determined, and if YES, it is determined that an abnormal gear shift has occurred. If all of the above three judgments are NO, it means that an abnormal gear shift has not occurred, and therefore the flow returns to the original step S126 of the flowchart shown in FIG.

If even one of the above three judgments is YES, or if the sensor has not detected a failure, that is, if YES is determined in step S298, the current pulley is determined in step S320. Number of rotations N
Retain _S and N _P as N _PH and N _SH . Then, in step S322, the fail flag XSTF1F is set. In step S324, the solenoid 55 is turned off. As explained in connection with FIG.
When the solenoid 55 is turned off, the primary chamber 27
Is closed, the current gear ratio is forcibly maintained. From the flow chart of FIG. 23 to step S1 of FIG.
Returning to step 26, the above control procedure is repeated. Here, for example, in step S126, the pulley rotation speeds NS and NP when the gear shift abnormality is detected are used in the calculation of RATIO, but since these are the rotation speeds when the abnormal gear shift is first detected, Since the values themselves do not indicate a runaway condition, there is no problem in using these rotational speeds for calculating RATIO.

Once the fail flag XSTF1F is set in step S322, the rotational speed NP at that time, NP,
NS is held as N _PH and N _{SH in} step S320. Then, the process proceeds from step S300 to step S330, and the N _PH and N _SH saved in step S320 are
It is used as the subsequent pulley rotation speeds N _P and N _S. Also,
The OFF state of the solenoid 55 is also maintained in step S332.

Thresholds N _TH , Δ _TH and δ to be compared in step S302, step S306 and step S312.
_The problem is how to set _TH etc. First,
These values must be such that fail occurrence can be detected. If a small value is set in order to increase the detection sensitivity, false detection will be caused. If a large value is set, the detection will be delayed and the values of N _P and N _S at that time cannot be specified in step S330. Therefore, these threshold values are determined in consideration of them. Control at the time of sudden braking Whether or not the sudden braking operation is performed by the driver is made at step S127 of FIG.
The details of the control are shown in FIG. Figure 2
In step S400 of No. 4, the rotational speed N _S of the secondary pulley is less than or equal to a predetermined value (N _S0 ) and the gear ratio RATI
When the state in which O (abbreviated as R) is equal to or less than the predetermined value (R ₀ ) is detected, the flag EMRBRF indicating that the sudden braking is applied is set in step S402. Then, in order to store the gear ratio at this time, step S4
At 04, the rotational speeds NP and NS are held as N _PH and N _SH .
Then, in order to fix the gear ratio in step S406,
The solenoid 55 is turned off. The fixed gear ratio prevents the belt from moving along the surface of the pulley no matter how the gear ratio is changed by the gear change control system, so that the belt is not overloaded and is durable. It is possible to prevent deterioration of the property and suppress the generation of noise.

The fixing of the gear ratio is continued until the rotation speed of the secondary pulley rises to a predetermined value (N _S1 ).
The control for this continuation will be described with reference to the primary hydraulic control of FIG. Control of primary hydraulic pressure In step S500 of FIG. 25, the above-mentioned flag EM is used.
The set status of the RBRF is checked. First, in the case of normal driving without sudden stop (EMRBRF = 0)
Will be described.

In step S520, it is determined whether the conditions for performing feedback control of the primary hydraulic pressure are met. If this condition is not met, the open control is performed in step S550. This open control is as described with reference to 223 in FIG. If it is determined in step S520 that the feedback control execution conditions are met, the process proceeds to step S522 to determine whether the timer TMR has timed out. This timer is a timer that defines the execution period of the "return control" and is set in step S504. Therefore, at the time of normal running, since TMR = 0, step S
Proceeding to 540, the pulley rotation speeds N _P and N _S and traveling conditions, that is, the throttle opening TVO, the traveling range, and the like are read from the sensor. In step S542, the target rotation speed PREVT is calculated based on these values.

Next, the case where the sudden braking is applied will be described. In this case, since EMRBRF is set in step S402 (FIG. 24), step S50
From 0, the process proceeds to step S502. In step S502, the pulley rotation speeds N _P and N _S are read. In step S503, it is checked whether the secondary rotation speed N _S has reached the predetermined value N _S1 . This is to detect that the vehicle has started again after a sudden stop and reached a predetermined vehicle speed. The above-mentioned N _S0 is a threshold value for detecting a sudden braking state, and this N _S1 is a rotational speed at which it is considered that the belt will not be burdened even if the gear ratio is returned from the fixed state to the normal state. Therefore, N _S0
The relationship between N _S1 and N _S1 is N _S0 <N _S1 .

After the engine is started, the pulley rotation speed increases, and if N _S > N _S1 , step S50
In step 4, the timer TMR is set to a predetermined time. Since it is not necessary to set the gear ratio to a fixed value any more, EMRBRF is reset in step S506, and the solenoid 55 is turned on in step S508. Once the timer TMR is set, YES is determined in step S522 until the timer times out, so the process proceeds to step S524. Step S522 to Step S53
The control of No. 2 is the above-mentioned “return control” (300 in FIG. 9). This "return control" is to calculate how much the current gear ratio R is different from the normal gear ratio R ', and perform feedback control so that this difference is reduced according to this difference.

That is, in step S524, the pulley rotation speeds N _P and N _S and traveling conditions, that is, the throttle opening TVO, the traveling range, and the like are read from the sensor. Then, in step S526, the current actual gear ratio R is calculated from the pulley rotation speeds N _P and N _S. In step S528, the target gear ratio R'is temporarily calculated based on the TVO, the driving range, and the like. In step S530, the target gear ratio R _T is calculated as follows.

If R> R ′, R _T = R + ΔR / ΔT If R <R ′, R _T = R−ΔR / ΔT Here, ΔT is a calculation cycle. This _RT is the target value PRE
VT. Thus, the target gear ratio PREVT is gradually feedback-controlled from the gear ratio when the sudden braking is applied to the target gear ratio R _T that is optimum for the operating condition at that time, so the belt gradually moves to the pulley. Since it slides in the radial direction, there is no strain on the belt. <Effects of the Embodiment> As described above, according to the hydraulic control device of the present embodiment, when the braking is suddenly applied, the gear ratio is fixed so that the belt is placed on the surface of the pulley. The belt will not be moved along with it, the belt will not be overloaded, and durability can be prevented from lowering and noise can be suppressed. H: Since the gear ratio is fixed until the number of rotations of the secondary pulley rises to a predetermined value (N _S1 ), the gear ratio is not released immediately after starting the vehicle. There is no undue burden. T: After the secondary pulley rotation speed has risen to the predetermined value (N _S1 ), the gear ratio control is slowly returned from the previously fixed state to the normal control to reduce the burden on the belt. is doing. <Modifications> The present invention can be variously modified without departing from the scope thereof.

For example, N _{S in} step S503 of FIG.
Instead of> N _S1 , R ′ <(N _PH / N _SH ).

[0076]

As described above, according to the present invention,
Under certain conditions where the belt will be burdened, the gear ratio is fixed, so even if the metaphorical driving condition requires a change of the pulley (change of the gear ratio),
The diameter of the pulley is not changed, and the belt is not forced to move along the wall surface of the pulley. For this reason, the durability of the belt is not reduced and no abnormal sound is generated.

[Brief description of drawings]

FIG. 1 is a diagram illustrating an inconvenience that occurs when the present invention is not applied.

FIG. 2 is a circuit diagram showing a configuration of a hydraulic circuit connected to the continuously variable transmission when the hydraulic control device for a hydraulic differential transmission according to the present invention is applied to the continuously variable transmission;

FIG. 3 is a skeleton diagram schematically showing the configuration of a continuously variable transmission to which an embodiment of a hydraulic control device for a hydraulic differential transmission according to the present invention is applied;

FIG. 4 is a block diagram showing a connection configuration of a control device 78,

FIG. 5 is a diagram showing the overall control of one embodiment according to the present invention,

FIG. 6 is a diagram showing the overall control of one embodiment according to the present invention,

FIG. 7 is a diagram showing a characteristic that determines a control range of a lockup operation;

FIG. 8 is a diagram illustrating control for calculating the duty of a solenoid for a lockup clutch,

FIG. 9 is a diagram illustrating control when a target rotation speed of a primary pulley is calculated,

FIG. 10 is a diagram illustrating control when calculating an engine torque TQPUMP,

FIG. 11 is a diagram illustrating control when calculating turbine torque TQTRBN;

FIG. 12 is a diagram illustrating control when performing centrifugal force correction,

[Fig. 13]

FIG. 14

FIG. 15

FIG. 16 is a diagram illustrating a situation in which inconvenience occurs when centrifugal force correction is performed,

FIG. 17 is a diagram for explaining a method for controlling the drive cycle of the duty solenoid,

FIG. 18 is a flowchart showing the overall control procedure according to the embodiment.

FIG. 19 is a diagram for explaining an operation of converting a target pressure into a duty ratio,

FIG. 20 is a flowchart showing details of line pressure control,

FIG. 21

FIG. 22 is a diagram for explaining the correspondence between feedback control and feedforward control;

FIG. 23 is a flow chart showing a procedure of detection control and prevention control of abnormal gear shift,

FIG. 24 is a flowchart showing a control procedure for detecting the occurrence of a sudden braking state,

FIG. 25 is a flowchart showing a control procedure for determining a primary hydraulic pressure.

[Explanation of symbols]

A ... Engine, B ... Torque converter, C ... Forward / reverse switching mechanism, D ... Belt transmission mechanism, E ... Reduction mechanism, F ... Differential mechanism, RATIO ... Gear ratio, E ... Speed ratio, TR ... Torque ratio, NE ... engine speed, NP ... primary shaft speed, NS ... secondary shaft speed, NT ... turbine speed, P ... line pressure, Q ... hydraulic circuit, R ... electric control circuit, Z ... continuously variable transmission, 1 ... Output shaft, 2 ... Turbine shaft, 3
... Pump impeller, 4 ... Turbine runner, 5 ... Stator, 6 ... Lockup piston, 7 ... Pump cover, 7
a: converter front room, 8: one-way clutch,
9 ... Stator shaft, 10 ... Converter rear chamber, 11 ... Ring gear, 12 ... Sun gear, 13 ... First pinion gear, 1
4 ... 2nd pinion gear, 15 ... Carrier, 16 ... Clutch, 17 ... Brake, 18 ... Accumulator, 19 ... Mission case, 20 ... Belt, 21 ... Primary pulley, 21a ... Belt receiving groove, 22 ... Primary shaft, 23
... Fixed conical plate, 24 ... Movable conical plate, 24a ... Outer surface, 2
5 ... Cylinder, 26 ... Piston, 27 ... Primary chamber,
31 ... Secondary pulley, 31a ... Belt receiving groove, 32
... secondary shaft, 33 ... fixed conical plate, 34 ... movable conical plate, 34a ... conical friction surface, 34b ... outer surface, 35 ... cylinder, 36 ... piston, 37 ... secondary chamber, 38 ...
Pressing spring, 40 ... Oil pump, 41 ... Line pressure adjusting valve, 42 ... Pressure reducing valve, 43 ... Gear ratio control valve, 44 ... Gear ratio fixed valve, 47 ... Mamual shift valve, 46 ... Clutch pressure adjusting valve, 49 … Lockup control valve, 48… Relief valve,
51, 52, 53, 54, 55 ... Solenoid, 78 ... Control unit, 82 ... Shift position sensor, 83 ... Primary speed sensor, 84 ... Secondary speed sensor, 85
... Slot throttle sensor, 86 ... Rotation speed sensor, 87 ...
Turbine speed sensor, 89 ... Oil pressure sensor

Claims (4)

[Claims] 1. A shift control device for a continuously variable transmission, wherein a V-belt is stretched between a pair of pulleys whose effective radius is changed by changing the supplied hydraulic pressure. A gear ratio control valve for controlling the supplied hydraulic pressure according to the gear ratio, a detection unit for detecting the number of rotations of the wheel, a unit for detecting the gear ratio, and the detected number of wheel rotations is the first. A shift control device for a continuously variable transmission, comprising: control means for controlling the gear ratio control valve so that the gear ratio becomes a fixed value when the detected gear ratio is less than or equal to a predetermined value .. 2. The shift control device for a continuously variable transmission according to claim 1, wherein the control means is a valve operatively located upstream of the gear ratio control valve and is supplied to the pair of pulleys. It is a fixed gear ratio valve that closes the hydraulic pressure supply path. 3. The shift control device for a continuously variable transmission according to claim 1, wherein the control means fixes the gear ratio and then the gear ratio detected by the detecting means has a preset gear ratio characteristic. When the gear ratio approaches the basic gear ratio, the gear ratio fixation is released. 4. The shift control device for a continuously variable transmission according to claim 3, wherein the control means gradually feedback controls the gear ratio toward the target gear ratio after the fixation of the gear ratio is released. To do.