JPH02118260A - Device for controlling speed change of continuously variable transmission - Google Patents

Abstract

PURPOSE:To improve accelerating performance by restraining the controlling action of increase in gear ratio by a gear-ratio restraining means when an input torque indicates a rate of change in a negative direction at the time of shift down accompanying an accelerating operation. CONSTITUTION:An input torque rate-of-change operating portion 71 obtains the rate of change of input torque from a present input torque value by a primary pulley input torque operating portion 69 and a previous input torque value stored in a memory means 70. At the time of sudden acceleration, in the case of the increasing zone of pulley ratio accompanying the shift down of a transmission with a negative rate of change of input torque, a target engine speed determining portion 66 correct an original target engine speed which is obtained from a map to be smaller by a defined value and, after that, set same to a final target engine speed, and a duty ratio operating portion 73 calculates a duty ratio for controlling gear ratio corresponding thereto. Thereby, the deficiency of torque at the time of accelerating can be prevented.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to a speed change control device for a continuously variable transmission capable of continuously variable control of a speed change ratio.

(Prior art) For example, as shown in Japanese Patent Application Laid-Open No. 60-222647, (1) a drive pulley (primary pulley) and a driven pulley (secondary pulley) are connected to each other via a belt. Each effective diameter can be changed arbitrarily by a hydraulic control means (a combination of a hydraulic cylinder and an electromagnetic control valve), and the hydraulic control means is electronically controlled by, for example, a transmission control unit constituted by a microcomputer. Continuously variable transmissions, which allow the gear ratio of the transmission to be continuously varied within the entire shift range from the highest gear ratio to the lowest gear ratio, have recently been widely used in automobiles. It's starting to look like this.

However, in the case of such a variable pulley type continuously variable transmission,
Normally, the line pressure of the hydraulic system line corresponding to the oil discharged from the oil pump is first regulated to a predetermined pressure with a pressure regulating valve, and then the line pressure is applied to the hydraulic cylinder for variable effective diameter of the driven pulley, and the line pressure is adjusted to the flow rate. The gear ratio is continuously changed by changing the effective diameter of each pulley by supplying the oil to the variable effective diameter hydraulic cylinders of the driving pulleys through control valves. Therefore, even if the driver rapidly depresses the accelerator during sudden vehicle acceleration, for example, as long as the opening degree of the flow control valve remains constant, the gear ratio will not drop rapidly and a certain delay time will inevitably occur.

Therefore, there is a problem that the response during sudden acceleration is poor.

Therefore, as a countermeasure to this problem, for example, in the configuration of the above-mentioned Japanese Patent Application Laid-Open No. 60-222647, which was previously described as a conventional example, when the driver suddenly depresses the accelerator as described above, the state is Detected by a sensor,
Within a predetermined period of time from the time of detection, the opening degree of the flow rate control valve is controlled to be larger than the normal opening degree by a predetermined opening degree to rapidly lower the gear ratio and quickly increase the torque.

(Problems to be Solved by the Invention) According to the configuration of the prior art described above, it is thought that it is possible to reduce the gear ratio quickly during sudden acceleration.

However, if we examine the relationship between the reduction in the gear ratio and the engine speed in detail, we will find the following.

That is, during sudden acceleration, the throttle opening degree increases rapidly and the engine speed increases significantly. As a result, in the sudden acceleration state, the original engine torque is absorbed by the inertia of the engine rotation, and even though the engine speed increases, the output torque from the engine does not increase and the drive side pulley Such a reduction may occur that the engine torque input to the engine is substantially reduced. Therefore, even if the gear ratio is simply lowered in response to sudden acceleration as in the prior art described above, it is not possible to achieve sufficient torque increase and sufficient acceleration performance.

(Means for Solving the Problems) The present invention has been made for the purpose of solving the above problems, and includes a driving side pulley and a driven side pulley, each of which is configured to have an effective diameter that can be changed arbitrarily. At the same time, a belt for transmitting driving force is stretched between the driving pulley and the driven pulley, and the effective diameters of the driving pulley and the driven pulley are variably controlled by the gear ratio control means, thereby continuously transmitting the driving force. In a continuously variable transmission configured to change a gear ratio, input torque change rate detection means detects a rate of change in the amount of input torque input from the engine to the driving pulley, and the gear ratio is controlled by the gear ratio control means. and a gear ratio suppressing means for suppressing the increase control, the change rate of the input torque to the driving pulley detected by the input torque change rate detecting means at the time of downshifting of the transmission accompanying acceleration operation of the vehicle is provided. The present invention is characterized in that when a rate of change in the negative direction is indicated, the speed ratio suppressing means suppresses the speed ratio increasing control operation by the speed ratio controlling means.

(Function) In the configuration of the speed change control device for the continuously variable transmission of the present invention, the input torque change rate detection means detects the change rate of the input torque input from the output shaft of the engine to the drive pulley, and the acceleration operation is performed. When it is detected that the input torque has changed in a negative direction (decreasing direction) during a transmission downshift corresponding to do. Therefore, even when there is a possibility that the engine's output torque will substantially decrease due to the increase in inertia caused by the increase in the engine speed, such as during sudden acceleration, resulting in a torque shortage, the engine torque will be reduced as much as possible. This makes it possible to prevent a decline in

(Effects of the Invention) Therefore, according to the shift control device for a continuously variable transmission of the present invention,
The lack of torque during acceleration is resolved, and it becomes possible to provide a continuously variable transmission-equipped vehicle with high acceleration responsiveness.

(Embodiments) Hereinafter, embodiments of the present invention will be described in detail with reference to FIGS. 1 to 10 of the drawings.

First, FIG. 1 is a skeleton diagram mainly showing the overall configuration of the transmission section of a speed change control device for a continuously variable transmission according to an embodiment of the present invention, and FIG. The heavy pressure circuit diagram of the step-change transmission section and the circuit configuration diagram of the control section thereof are shown in detail in FIG. 3, respectively. Therefore, first, the overall configuration of the continuously variable transmission will be explained with reference to FIG. 1, and then the configuration of the hydraulic circuit section will be explained with reference to FIG. The configuration and operation of the control circuit shown in FIG. 3 will be explained with reference to the flowcharts and graphs shown in FIG. 4 and subsequent figures.

First, in FIG. 1, the continuously variable transmission is configured as a continuously variable transmission for a front wheel drive vehicle, for example, and the continuously variable transmission is configured as a continuously variable transmission for a front wheel drive vehicle.
A torque converter B, a forward/reverse switching mechanism C, a belt transmission mechanism, a deceleration mechanism E, and a differential mechanism F are connected to the output shaft L of the
It is equipped with

The torque converter B includes a pump impeller 3 that is fixed to one side of a pump cover 7 connected to an output shaft l of the engine and rotates integrally with the engine output shaft l, and a pump impeller 3 that faces the pump impeller 3. A turbine runner 4 is rotatably provided in a converter chamber 7a formed inside the pump cover 7, and a turbine runner 4 is interposed between the pump impeller 3 and the turbine runner 4 to perform torque multiplication. The stator 5 has a stator 5 for carrying out the operation. Further, the turbine runner 4 is connected to 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 the minyon case 19 via a one-way clutch 8 and a stator shaft 9. Each is connected.

Furthermore, a lock-up piston 6 is arranged between the turbine runner 4 and the pump cover 7. This lock-up piston 6 is slidably attached to the turbine shaft 2, and when hydraulic pressure is introduced into or discharged from the lock-up chamber 10, the pump cover 7
A lock-up state in which the converter is in contact with and integrated with the pump cover 7 and a converter state in which the converter is separated from the pump cover 7 can be selectively realized. In the lock-up state, the engine output shaft 1 and the turbine shaft 2 are directly connected without fluid, and in the converter state, the engine torque is
and are transmitted to the turbine shaft 2 side via fluid.

Next, the forward/reverse switching mechanism C is connected to the torque converter B.
This mechanism is for selectively setting a forward state in which the rotation of the turbine shaft 2 is directly transmitted to the belt transmission mechanism, which will be described later, and a reverse state, in which the rotation is transmitted in reverse to the belt transmission mechanism, which will be described later. In the embodiment, this forward/reverse switching mechanism C is comprised of a double pinion type planetary gear unit. That is, 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 are attached to the carrier +5 spline-coupled to the turbine shaft 2. Incidentally, the sun gear 12 is spline-coupled to a primary shaft 22 of a belt transmission +MD, which will be described later.

Further, between the ring gear 11 and the carrier 15, there is a clutch 16 for connecting and disconnecting the two, and between the ring gear 11 and the transmission case 19, there is a clutch 16 for selectively connecting the ring gear 11 to the transmission case I9. A brake I7 for fixing the vehicle is provided respectively.

Therefore, when the clutch 16 is engaged and the brake 17 is released, the ring gear II and carrier 15
Since the ring gear II is made to be able to rotate relative to the transmission case 19, the rotation of the turbine shaft 2 continues in the same direction, and is connected from the sun gear 12 to the primary shaft (driving side primary pulley). 2
output to the 22 side (which is the axis of I) (forward state)
.

On the other hand, when the clutch 16 is released and the brake 17 is engaged, the ring gear 11 is fixed to the transmission case 19 side and the ring gear 1X and the carrier 15 can rotate relative to each other, so that the turbine shaft 2 The rotation of the first pinion gear 13 and the second pinion gear 1
4 and is output to sun gear I2 in an inverted state (reverse state). That is, in this forward/reverse switching mechanism C, switching between forward and backward travel is performed by selectively operating the clutch 16 and brake 17.

Furthermore, the belt transmission mechanism includes a primary pulley 21, which will be described later, which is coaxially arranged on the rear side of the forward/reverse switching mechanism C, and a primary pulley 21, which will be described later, which is spaced apart in a direction parallel to the primary pulley 21. A V-belt 20 is stretched between it and a secondary pulley 31.

The primary pulley 21 has a predetermined diameter on a primary shaft 22 which is disposed coaxially with the turbine shaft 2 and whose one shaft end is spline-coupled to the sun gear 12 of the forward/reverse switching mechanism C. A fixed conical plate 23 is integrally mounted on the primary sleeve 22, and a movable conical plate 2 is integrally mounted on the primary sleeve 22.
4 are provided on the primary shaft 22 so as to be movable (slidable) in the axial direction thereof. The conical friction surface of the fixed conical plate 23 and the movable conical plate 24
The conical friction surface forms a belt groove 21a having a substantially V-shaped cross section.

Further, a cylindrical hydraulic cylinder 25 is fixed to the outer surface 24a side of the movable conical plate 24. Furthermore, a piston 26 fixed to the primary shaft 22 side is oil-tightly fitted into the inner circumferential surface of the cylinder 25, and the piston 26, the cylinder 25, and the movable conical plate 24 are the main components. A primary chamber 27 is configured. Note that a predetermined line pressure is introduced into this primary chamber 27 from a hydraulic circuit Q, which will be described later.

The primary pulley 21 is moved by the movable conical plate 2 by the hydraulic pressure introduced into the primary chamber 27.
The effective diameter of the belt 20 can be adjusted by moving the belt 4 in the axial direction to increase or decrease the distance between the belt 4 and the fixed conical plate 23.

The secondary pulley 31 basically has the same structure as the primary pulley 21, and is arranged at a distance of F square from the primary shaft 22 in a straight line. On the secondary shaft 32, the fixed conical plate 33 is integrated with the secondary sleeve 32,
Further, a movable conical plate 34 is provided so as to be movable on the secondary shaft 32. The conical friction surfaces of the fixed conical plate 33 and the movable conical plate 3 face each other.
4 and a belt groove 3 having a substantially V-shaped cross section as in the case of the primary pulley 21 side described above.
1a is formed.

Further, a substantially stepped cylinder 35 is coaxially fixed to the outer surface 34b of the movable conical plate 34. Also,
A piston 36 whose axial portion is fixed to the secondary sleeve 32 is fitted into the inner peripheral surface of the cylinder 35 in an oil-tight manner.

This piston 36, the above cylinder 35, and the movable cone @34
The secondary chamber 37 is configured with the champion of the above-mentioned primary pulley 2.
Line pressure is introduced from the hydraulic circuit Q similarly to the first side. Similar to the primary pulley 21, the secondary pulley 31 and the movable conical plate 34 are connected to the fixed conical plate 33.
The effective diameter of the belt 20 is adjusted by moving it toward and away from the belt 20.

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.

Further, the speed reduction mechanism E and the differential mechanism F are constructed from conventionally known mechanisms.

Now, to briefly explain the operation of this continuously variable transmission, firstly, the driving torque transmitted from the engine 8 through the torque converter B is transferred to the forward/reverse switching mechanism C so that the rotation direction is either the forward direction or the reverse direction. It is transmitted to the belt transmission mechanism in a state set to either of the following.

In belt transmission mechanism, primary pulley 2
The effective diameter of the hydraulic fluid is adjusted by introducing or discharging the hydraulic oil into the primary chamber 27 of the primary chamber 27, and the secondary pulley 31, which is interlocked and connected to the primary pulley 21 via the belt 20, follows suit. Its effective diameter is adjusted with . The ratio of the effective diameter of the primary pulley 21 to the effective diameter of the secondary pulley 31 (pulley ratio) determines the gear ratio between the primary shaft 22 and the secondary shaft 32.

The rotation of the secondary shaft 32 is further decelerated by the deceleration mechanism E, and then transmitted to the differential mechanism F, and from the differential mechanism F to the front axle (not shown).

Next, the hydraulic circuit shown in FIG. 2 will be explained.

The hydraulic circuit shown in FIG. 2 includes a lock-up piston 6 of the torque converter B in the continuously variable transmission, a clutch 16 and a brake 17 of the forward/reverse switching mechanism C, and a primary pulley 21 of the belt transmission mechanism. The oil pump 40, which is driven by the engine A, is provided to control the operations of the pulley 31 and the secondary pulley 31, respectively.

The hydraulic oil discharged from the oil pump 40 is first adjusted to a predetermined line pressure by the line pressure regulating valve 41, and then is sent to the secondary chamber 37 of the secondary pulley 31 mentioned above via the main line 101. The primary chambers 27 of the primary pulleys 21 are supplied through branch lines 102 branched from the main line IOI.

Line pressure control in the line pressure regulating valve 41 is performed by controlling the pilot pressure introduced into the pilot chamber 41a. That is, the line pressure regulator 41 includes a spool 41b and a spring 41c that biases the spool 41b, and a pressure regulating boat 41d to which the oil discharged from the oil pump 40 is guided, and the oil pump 40. A drain boat 41e communicating with the suction side is provided. Further, hydraulic oil, which is branched from the branch line 102 and reduced to a predetermined pressure by the reducing valve 42, is introduced into the pilot chamber 41a as pilot pressure via the branch line 103. Therefore, in this line pressure regulating valve 41, the spool 41b introduces the hydraulic pressure in the main line 101 applied to one end thereof, the spring force of the spring 41c applied to the other end, and the pilot chamber 41a. The drain boat 41e is moved to the pressure regulating boat 41 by sliding in accordance with the balance between the pilot pressure and the resultant force.
By communicating or cutting off communication with d, a line pressure corresponding to the pilot pressure is generated. Further, the pilot pressure for controlling this line pressure is specifically adjusted by controlling the duty ratio of the first electromagnetic solenoid valve 51 provided in the branch line 103.

Further, the branch line 102 is provided with a gear ratio control valve 43 that operates in response to pilot pressure, and the gear ratio of the above-mentioned continuously variable transmission is controlled by this gear ratio control valve 43. 21 primary room 2
This is done by controlling the supply and discharge of hydraulic oil to and from 7. That is, the gear ratio control valve 43 includes a spool 43a that is always pressed toward the negative side by a spring 43b, and a line pressure boat 4 that communicates with the line 102.
3c, the drain boat 43d, and the above spring 43b.
When a shift valve 45 (described later) is set to the reverse position (R), line pressure is supplied to the pilot boat 43e which opens to the pilot chamber 43f formed on the opposite end surface side, and which opens to the spring 43b side and which will be described later. It has a reverse boat 43g which is introduced. When moving forward (when the shift valve 45 is in the soft position of either shift range D or 2.1), the reverse port 43
From where g is drained via the shift valve 45,
The spool 43a becomes slidable in the axial direction in response to pilot pressure introduced into the pilot chamber 43f. Then, the line pressure boat 43 is operated by the spool 43a.
c and the drain boat 43d are selectively connected to the primary chamber 2.
7, the primary chamber 27
Control of the supply and discharge of hydraulic oil to and from the engine, ie, gear ratio control, is carried out. On the other hand, when traveling in reverse, line pressure is introduced from the reverse port 43g, and the spool 43a receives this line pressure and is fixed in a state where it is fully pressed to the right in the figure.

Therefore, regardless of the pilot pressure, the line pressure boat 43c and the drain boat 43d are always in communication, and the gear ratio is kept fixed at the maximum gear ratio.

Incidentally, in this embodiment, two pilot pressure supply systems to the speed ratio control valve 43 are provided, and these are selectively used by a switching valve 44, which will be described later. That is, the switching valve 44 includes a spool 44a and a spring 44b that biases the spool 44a toward one side. The switching valve 44 also connects the pilot boat 44c, which has an opening at the end opposite to the spring 44b, to the line 1.
Connect to line 105 branched from 03 and connect to spool 44
Pilot pressure reduced by the reducing valve 42 is applied to one end of a.

Furthermore, the line 10 is located in the middle part of this switching valve 44.
5, a second pilot pressure introduction boat 44e that communicates with the pitot pressure generating means, and a pilot boat 4 of the gear ratio control valve 43.
A pilot pressure supply boat 44f is adjacent to the pilot pressure supply boat 44f, which communicates with the pilot pressure supply boat 3e via a line IO4. A second electromagnetic solenoid valve 52 is provided in the line 105 communicating with the first pilot pressure introduction boat 44d, and the pressure is regulated by the second electromagnetic solenoid valve 52 depending on the operating state of the second electromagnetic solenoid valve 52. The generated oil pressure and pitot pressure generated in accordance with the rotational speed of the engine are selectively supplied as pilot pressure to the pilot chamber 43 of the gear ratio control valve 43 to perform predetermined gear ratio control. .

On the other hand, the hydraulic oil whose pressure has been regulated by the line pressure regulating valve 41 is introduced into the boat A of the switching valve 44 via the line 106. The hydraulic oil supplied to boat A is
When the reverse gear is set, the brake chamber 62 of the brake 17 is connected to the brake chamber 62 of the brake 17 through the line 107, and when the forward gear is set, the clutch chamber 6 of the clutch 16 is connected through the line 10B.
1, respectively, to put the forward/reverse switching mechanism C into the reverse or forward operating state. In this embodiment, the accumulator 1 is connected between the line 107 and the line 108.
8 is provided, and this one accumulator 18 effectively relieves the engagement shock of both the clutch 16 and the brake I7.

In addition, the hydraulic oil whose pressure is regulated by the line pressure regulating valve 41 is
After being regulated to a predetermined clutch pressure by the clutch pressure regulating valve 46, the clutch pressure is introduced into the lock-up control valve 47 via the line 109. By controlling the pilot pressure of the lock-up control valve 47 with the third electromagnetic solenoid valve 53, the hydraulic oil introduced into the lock-up control valve 47 is applied to the lock-up engaging side (LOCK) or the lock-up releasing side (LINL).
OCK).

Further, in FIG. 2, reference numeral 48 is a relief valve.

Next, a method of controlling the gear ratio control valve 43 will be described in detail with reference to FIGS. 2 to 7.

First, the control range of the second electromagnetic solenoid valve 52 and the switching valve 4
4, the second electromagnetic solenoid valve 52 adjusts the hydraulic pressure (pilot pressure) in the line 105 in response to the duty ratio changing from 0 to 100% as shown in FIG. can be changed in the range from 0 to P.

On the other hand, in the switching valve 44, the spool 44a moves in the axial direction according to the oil pressure applied to the pilot boat 44c, and selectively changes the pilot pressure to the first pilot pressure introduction boat 44d and the second pilot pressure introduction boat 44e. In this embodiment, the eleventh second pilot pressure introduction boat 44d is configured to communicate with the supply boat 44r.

The relative positions of the pilot pressure supply boat 44e and the pilot pressure supply boat 44f are set as follows in accordance with the magnitude of the pilot pressure, that is, the duty ratio of the second electromagnetic solenoid valve 52. That is, the duty ratio of the second electromagnetic solenoid valve 52 is ~D
When it is within the range of 1%, that is, the pilot pressure is Po=P,
When the pressure is within the range, the spool 44a is located between the positions shown in the upper and lower rows of FIG. Introduction boat 44
e is in the closed state, and when the duty ratio of the second N magnetic solenoid valve 52 is within the range of O-D,%, that is, when the pilot pressure is within the range of P o-P +, as shown in the upper part of Fig. 2. , the spool 44a moves fully to the right and the first
The eleventh second pilot pressure introduction boats 44d and 44e and the pilot pressure supply boat are connected so that the second pilot pressure introduction boat 44e communicates with the pilot pressure supply boat 44f while the pilot pressure introduction boat 44d is closed. 44'' is set relative to the duty ratio of the second electromagnetic solenoid valve 52, that is, the oil pressure in the line 105.The speed change control by this second electromagnetic solenoid valve 52 is as shown in FIG. As shown in FIG. 2, the vehicle speed (i.e., the rotational speed of the secondary pulley 31) and the throttle opening are set as parameters for each shift position in advance to correspond to the current driving state determined from the map of the target primary pulley rotational speed. The duty ratio for operating the second electromagnetic solenoid valve 52 corresponding to the target gear ratio is calculated from the deviation between the target primary pulley rotation speed and the actual primary pulley rotation speed, and the duty ratio is activated by a signal of the duty ratio. The pilot pressure of the gear ratio control valve 43 is adjusted by the second electromagnetic solenoid valve 52, and the primary
By controlling the supply and discharge of hydraulic oil to the chamber 27, the gear ratio is controlled in accordance with the target value.

The duty ratio of the control signal, which is the basis for controlling the operation of the second electromagnetic solenoid valve 52, is determined, for example, according to the flowchart shown in FIG. That is, first, in steps S+ and S-, various states are initialized and signals from each sensor (throttle opening, engine speed, vehicle speed, etc.) are read, and then, in step S, clutch pressure control is performed. Further, in step S4, lock-up control is performed to engage and release the lock-up clutch depending on the operating state. Further, determine the duty ratios of the electromagnetic solenoid valves 5t to 53 using the gear ratio control (step S) and secondary pressure control (step S8), which will be described in detail later.

Then, in step S7, a duty signal is outputted to each of the electromagnetic solenoids 51 to 53 by -4° so as to achieve the set duty ratio.

Further, the gear ratio control in step S is specifically performed as follows according to the flowchart shown in FIG.

In this control, first, in step S, the soft position is
After reading the throttle opening of the engine and the rotation speed of the secondary pulley 21, in step S+t, the target primary at the soft position is determined based on the vehicle speed corresponding to the rotation speed of the secondary pulley 31 and the throttle opening. -Calculate the pulley rotation speed No. As mentioned above, as shown in Figure 6, this calculation is performed using a map of the target primary pulley rotation speed, which is preset for each shift position using vehicle speed and throttle opening as parameters, and the actual vehicle speed. This is done by comparing and referencing (interpolating) and the throttle opening.

Next step S13. In S L4, the actual rotation speed Np of the primary pulley 21 is read, and in the following step srs, this and the target primary pulley rotation speed N are combined.
The deviation ΔNl) from po is calculated. Further, in step S16, the duty ratio of the second electromagnetic solenoid valve corresponding to this deviation ΔN is determined according to the characteristics shown in FIG.

In this way, by changing the duty ratio of the second electromagnetic solenoid valve 52 according to the deviation ΔNp, hydraulic oil is supplied to the hydraulic cylinder 25 of the primary pulley 21 via the gear ratio control valve 43. Under the control, the effective pitch diameter of the primary pulley 21 is changed, and accordingly, the rotation speed of the primary pulley 21 changes in a direction that matches the target rotation speed Npo. Thereby, the rotational speed of the primary pulley 21, in other words, the gear ratio of the belt transmission mechanism, is optimally controlled in accordance with the vehicle speed and the throttle opening.

Next, the structure and operation of the essential parts of the embodiment of the present invention in the transmission control unit that controls the transmission as described above is particularly shown as a functional block diagram in FIG.

In FIG. 3, reference numeral 61 is an engine rotation speed detection means (in this embodiment, since a lock-up state is assumed, the engine rotation speed detection means is ultimately a rotation speed detection means of the primary pulley 21). 62 is a vehicle speed detection means, and 63 is a throttle opening detection means. Further, reference numeral 64 indicates the engine rotation speed detection means 6.
The actual engine speed detected by 1 (here, we assume a lock-up state as described above, so
The engine speed NE is naturally the primary pulley 21.
This is an engine rotation speed storage unit (RAM) that stores NF (Np) (equal to the rotation speed of the engine) within a predetermined period (at least one control cycle or more).

On the other hand, reference numeral 65 denotes a target rotation speed calculation unit that calculates the target rotation speed Npo of the above-mentioned primary pulley 21 for gear ratio control, and the target rotation speed calculation unit 65 is connected to the vehicle speed detection means 62 and the throttle opening. The detection outputs V and θTVQ of the detection means 63 are inputted, and the target rotational speed Npo of the primary pulley 21 is calculated (map interpolation) for controlling the gear ratio according to the operating state specified by the two inputs (as described above). (See Figure 6). The calculated value Npo is input to the final target rotational speed determination section 66 and the rotational speed change rate calculation section 67 of the next primary pulley 21, respectively.

Further, reference numeral 68 is an engine torque calculation section, and the engine torque calculation section 68 uses the output Nε (Np) of the engine rotation speed detection means 61 and the throttle opening detection means 6.
The input torque calculation unit 69 of the drive-side primary pulley 21 calculates the actual engine torque (engine output shaft torque) ETQ at the time based on the output θTVO of No. 3.
supply to.

The primary pulley man torque calculation section 69 calculates PTQ=ETQ-π/30 from the rate of change ΔNp=dNp/dt of the rotational speed N+ calculated by the rotational speed change rate calculation section 67 and the engine torque ETQ. - Calculate the actual input torque PTQ to the primary pulley 21 by calculating dN/di (in the above calculation formula,
π/30·Ie indicates engine inertia based on angular velocity fluctuations). Therefore, this calculated value P T Q is obtained after the inertia torque of the engine is subtracted. Then, this calculation (direction P'rQ) is stored in the primary pulley driver torque storage section 70 for -1 time.

Further, reference numeral 71 indicates the current input torque value P T calculated by the primary pulley torque calculation unit 69.
Qn and the previous input torque value PTQ(n-1) stored in the bipedal input torque storage means 70 are input, and the rate of change of the input torque -)TQ is calculated as ΔP'rQ=dPTQ/d.
This is an input torque change rate calculation unit that calculates t. Calculated value ΔP 1' Q calculated by the input torque change rate calculation section 71
is supplied to the target rotation speed determining section 66 in the same way as the calculated value Npo of the bipedal target rotation speed. Target rotation speed determination unit 66
corrects the target rotational speed Npo calculated by the target rotational speed calculation section 65 according to the input torque change rate ΔPTQ calculated by the input torque change rate calculation section 71, and calculates the final target rotational speed Npo. ′ is determined.

That is, as mentioned earlier, when the throttle opening rapidly increases and the engine speed NE rapidly increases, such as during sudden acceleration, the inertia torque (zr /
30 ・I e-dNE/dt), the original output torque of the ennon is absorbed, resulting in a torque shortage, that is, a phenomenon occurs in which the amount of change in the input torque to the primary pulley 21 changes in the negative direction. .

Therefore, the torque PTQ actually input to the primary pulley 21 due to this phenomenon is: PTQ=engine torque (ETQ) - engine inertia torque (π/30·1e
-dN/dt), which is reduced by the amount absorbed by the inertial torque.

Therefore, the target rotation speed determination unit 66 determines whether the rate of change ΔPTQ of the input torque PTQ is larger or smaller than O, and determines whether the rate of change ΔPTQ of the input torque PTQ is smaller than 0 (
In other words, in the region where the pulley ratio increases due to transmission downshift (see Figure 8), the original target rotation speed Npo determined from the map in Figure 6 above is corrected to a predetermined rotation speed smaller. Final target rotation speed Npo' (Np
o'< Npo), and the duty ratio calculation unit 73
Then, the corresponding duty ratio for gear ratio control is calculated. Then, the second electromagnetic solenoid valve 52 (gear ratio control valve 43) described above is driven. On the other hand, in the normal case where the rate of change ΔPTQ of the input torque is larger (increase) than O (near the overdriving state), the target rotation speed is controlled as usual based on the map value Npo shown in Fig. 6 above. conduct.

Therefore, in the case of the configuration of this embodiment, if the conventional configuration is used, for example, as shown in the graphs (a) to (C) of FIG. With the target primary pulley rotation speed Npo set high based on the map characteristics shown in FIG. 6, the gear ratio is adjusted to sufficiently increase the pulley ratio as shown in FIG. Control takes place. However, in this case, due to the rapid increase in engine speed, the engine torque is essentially absorbed by the inertial force of the increase in engine speed, so that the input to the primary pulley 2I increases accordingly. Torque becomes insufficient. As a result, as shown in the acceleration inertia characteristic in FIG. 11(b), there has been a problem in which the acceleration force temporarily downshoots during sudden acceleration, making the driver feel a torque shock.

However, in the case of this embodiment, as described above, in such a case, the target rotation speed Np of the primary pulley 21 itself
o is changed to a target rotational speed Npo' that is a predetermined rotational speed lower than the original value such that the input torque to the primary pulley 21 substantially increases. As shown in the characteristics, the pulley ratio of the continuously variable transmission itself that is controlled becomes lower than in the conventional case. As a result, even if the input torque to the primary pulley 21 is lower than the desired characteristic value, the transmission output torque required for acceleration can be obtained, as shown in FIG. 10(b). As a result, at least the down shoot of the acceleration force as in the conventional case does not occur (see Fig. 1O (a) to (c)).

[Brief explanation of the drawing]

FIG. 1 is a skeleton diagram showing the overall structure centering around the transmission section of a shift control device for a continuously variable transmission according to an embodiment of the present invention, and FIG. 2 is a hydraulic circuit diagram of the same embodiment. Figure 3 is a functional block diagram of the gear ratio control section of the same embodiment device; Figure 4;
FIG. 5 is a 3-flow chart showing the general duty ratio control operation of the device of the above embodiment, and FIG. 6 is a target rotation speed map of the primary pulley used in the gear ratio control of the device of the embodiment. FIG. 7 is a duty ratio-pilot pressure map, FIG. 8 is a graph showing the control region of the embodiment of the present invention, and FIG. 9 is a graph showing the relationship between rotational speed deviation and duty ratio of the embodiment. FIG. 1O is a graph showing the effects of the present invention in a time chart 1 to 1, and FIG. 1I is a graph similar to FIG. 10 showing the problems of the conventional example. 16...Clutch 17...Brake I8...Accumulator 20...Belt 21...Primary pulley 3 l ・ ・ ・ ・ 41 ・ ・ ・ ・ 42 ・・ ・ ・ 43 ・ ・ ・ ・ 44 ・ ・ ・ ・ 45 ・ ・ ・ 46 ・ ・ ・ 47 ・ ・ ・ 48 ・ ・ ・ ・ 51~53 ・ 61 ・ ・ ・ ・ 62 ・ ・ ・ ・ 63 ・ ・...・Switching valve・Knot valve・Clutch pressure adjustment valve・Lockup control valve・Relief valve・Electromagnetic solenoid valve・Engine rotation speed detection means・Vehicle speed detection means・Throttle opening detection means・Target rotation speed calculation unit・Target rotation speed determination・Rotational speed change calculation section ・Engine torque calculation section ・Primary pulley torque calculation section ・Input torque change rate calculation section A...Engine B...Torque converter C...・Forward/forward switching mechanism D...Belt transmission mechanism E...Deceleration mechanism F...Differential mechanism Time vehicle speed Fig. 6 Duty ratio Fig. 7 (deviation)

Claims (1)

[Claims] 1. A drive side pulley and a driven pulley each having an effective diameter that can be changed arbitrarily are provided, and a belt for transmitting driving force is stretched between the drive side pulley and the driven side pulley, and a gear ratio control means is provided. In a continuously variable transmission in which the gear ratio is continuously changed by variably controlling the mutual effective diameters of the driving pulley and the driven pulley, the amount of input torque input from the engine to the driving pulley is An input torque change rate detecting means for detecting the rate of change, and a speed ratio suppressing means for suppressing the increase control of the speed change ratio by the speed ratio control means, and the above-mentioned change rate is provided when the transmission is downshifted due to acceleration operation of the vehicle. When the rate of change of the input torque to the driving pulley detected by the input torque change rate detection means indicates a rate of change in the negative direction, the speed ratio suppressing means performs an operation to increase the speed ratio by the speed ratio control means. 1. A speed change control device for a continuously variable transmission, characterized in that it suppresses.