JPH10103494A - Controller for automatic transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To complete speed changing at a shift point at all times, and solve engine rotation over failure and drive feeling failure by changing speed on the basis of a shift point corrected according to at least acceleration of a vehicle, car speed, and a throttle opening. SOLUTION: It is predicted whether acceleration of a vehicle is a positive value or not, and speed changing which is subsequently generated on the basis of a present speed number. Line pressure P0 at the time of speed changing which is predicted is predicted from a throttle opening in Step 35, and a pre- charge base flow rate Q0, a pre-charge flow rate Q1 in which a temperature is corrected are found out from the P0 in Steps 36, 37. A stroke time T0 of a frictional element is found out from the Q1 and the like in Step 38. The T0 and a vehicle acceleration are multiplied together, and it is set to a car speed correcting rate VSP' in Step 39. The VSP' is subtracted from a shift judging speed, a shift judging speed FVSP' which is corrected is found out in Step 40, and the number of speed is judged. A shift command is outputted early in the case of every speed changing operations, speed changing is completed on a shift line, and car speed over and the like are suppressed.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control device for an automatic transmission mounted on an automobile.

[0002]

2. Description of the Related Art Generally, an automatic transmission mounted on an automobile combines a torque converter and a transmission gear mechanism,
The power transmission path of the transmission gear mechanism is switched by a selective operation of a plurality of friction elements such as a clutch and a brake to automatically shift to a predetermined gear.

[0003] In this case, the optimum gear position to be shifted is determined by, for example, using a shift map in which the throttle opening and the vehicle speed are set as parameters, as disclosed in Japanese Patent Application Laid-Open No. H4-254063. It is known to apply the measured values of

According to this, for example, when the vehicle speed is increasing as shown by an arrow a in FIG. 24, the upshift from the second speed to the third speed is performed at a point a where the vehicle speed crosses the 2-3 shift line. If the determination is made and the vehicle speed is decreasing as indicated by the arrow b, the downshift from the third speed to the second speed is determined at the point a where the vehicle speed crosses the 3-2 shift line.

[0005]

Incidentally, this type of automatic transmission is provided with a hydraulic control circuit for supplying and discharging hydraulic oil to and from a hydraulic chamber of each friction element such as the clutch and brake. Once the gear is determined,
In order to realize the gear, the on / off state or duty ratio of a hydraulic control valve such as a plurality of solenoid valves and a duty solenoid valve also arranged in a hydraulic control circuit is controlled to engage or release a predetermined friction element. It has become.

[0006] Here, the following problem occurs. That is, as described above, since the engagement or release of the friction element is performed by the supply and discharge of the hydraulic oil, a time lag occurs between the output of the shift command and the end of the actual shift. In this case, since the fluidity of the hydraulic oil is affected by the temperature, the time lag fluctuates. For example, at the shift point a in FIG.
Even at a speed of km / h, at normal temperature when the fluidity of the hydraulic oil is high, the speed change ends at 51 to 52 km / h or the like, while on the other hand, at a low temperature where the fluidity of the hydraulic oil decreases, 54 to 55 k is higher.
The shift ends at m / h or the like, so that the vehicle speed at the end of the shift is not constant, which not only gives the driver a sense of discomfort, but also
If the shift end is delayed at a low temperature, the engine speed becomes excessively high and the engine load becomes excessive.

Similarly, also in the case of the downshift shown by the arrow b in FIG. 24, at a low temperature, a longer time after the 3-2 shift command is output at the shift point a than at a normal temperature, 2 The shift to the high speed is completed. In this case, there is no problem of over rotation of the engine, but the problem of poor drive feeling remains.

[0008] Such a problem does not occur only due to the properties of the hydraulic oil. For example, even if the time lag from the output of the shift command to the end of the shift is the same, the vehicle may not operate properly. When traveling downhill or the like, a positive acceleration is likely to be applied compared to when traveling uphill or the like, so that the vehicle speed at the end of shifting is increased, and the problem such as excessive engine rotation also occurs.

SUMMARY OF THE INVENTION The present invention addresses the above-described situation in an automatic transmission, and aims to solve the problems of over-rotation of the engine and poor drive feeling by always terminating the shift at a predetermined shift point. As an issue.

[0010]

In order to solve the above problems, the present invention uses the following means.

First, in the invention described in claim 1 of the present application (hereinafter referred to as "first invention"), a vehicle speed detecting means for detecting a value relating to a vehicle speed, and a value relating to a throttle opening are detected. A throttle opening detecting means for determining a gear position based on a shift point predetermined by the throttle opening and the vehicle speed and a detection result of the two detecting means, and executing a shift to the gear position A control device for an automatic transmission, comprising: correction means for correcting a shift point according to at least an acceleration of a vehicle so that the shift ends at the shift point; and detecting the corrected shift point and both of the detection means. And a shift executing means for executing a shift based on the result.

According to a second aspect of the present invention, in the first aspect, an acceleration detecting means for detecting a value relating to the acceleration of the vehicle and a value relating to a time required for shifting are provided. A shift time calculating means for determining the shift point is provided, and the correcting means corrects the shift point according to the detection result of the acceleration detecting means and the calculation result of the shift time calculating means.

According to a third aspect of the present invention (hereinafter referred to as a "third invention"), the correction means determines whether the next shift is an upshift or a downshift based on vehicle acceleration. The first detecting unit detects the shift line pressure predicted from the throttle opening, the capacity of hydraulic oil to be supplied or discharged to or from a friction element to be engaged or released in the next shift, and the oil temperature to determine the friction. A second detection unit for detecting a time required for fastening or releasing the element, a third detection unit for detecting a vehicle speed correction amount at a shift point from the time required for the fastening or release, and the vehicle acceleration; and It is characterized by having a correction unit for correcting the shift point by using the correction unit.

On the other hand, according to the invention described in claim 4 (hereinafter referred to as "fourth invention"), in the third invention, the correction means is arranged such that when the absolute value of the vehicle acceleration is large, it is smaller than when the absolute value is small. The invention described in claim 5 (hereinafter referred to as a "fifth invention") is characterized in that the correction amount of the shift point is increased, and in the third invention, the correction means is required for fastening or releasing the friction element. When the time is long, the correction amount of the shift point is made larger than when the time is short.

According to a sixth aspect of the invention (hereinafter referred to as a "sixth aspect"), in the third aspect, the correction means suppresses detection of a shift point correction amount during a shift operation. It is characterized by.

According to a seventh aspect of the invention (hereinafter referred to as a "seventh invention"), there is provided an engine speed detecting means for detecting a value relating to the engine speed, and a shift predetermined by the engine speed is provided. A control device for an automatic transmission that determines a shift speed based on a point and a detection result of the detection means and executes a shift to the shift speed, wherein the shift position is determined so that the shift ends at the shift point. Is provided in accordance with at least the acceleration of the vehicle, and shift executing means for executing a shift based on the corrected shift point and the detection result of the detecting means is provided.

By using the above means, each invention of the present application operates as follows.

According to the first aspect of the present invention, when the shift speed is determined based on a shift point predetermined based on the throttle opening and the vehicle speed, at least the vehicle acceleration is adjusted so that the shift ends at the shift point. , The shift point is corrected in accordance with, so that the shift can be ended at a fixed shift point, and the problems of overspeed of the engine and poor shift feeling can be solved.

In this case, according to the second aspect, the shift point is corrected in accordance with the vehicle acceleration and the time required for the shift, so that the amount of change in the vehicle speed until the end of the shift is detected,
As a result, the corrected shift point can be determined by back calculation from the target vehicle speed at the end of the shift.

According to the third aspect of the invention, the second aspect of the invention is further embodied, and the type of the shift that occurs next is detected by the first detection unit, and the time required for the detected shift is detected by the second detection unit. Is detected by the third detection unit, the amount of change in the vehicle speed occurring during the shift is detected as a vehicle speed correction amount, and the correction unit corrects the shift point using the vehicle speed correction amount, and executes these processes. After that, the shift ends at a certain shift point.

In this case, according to the fourth and fifth aspects, the correction amount of the shift point is increased as the absolute value of the vehicle acceleration is increased or as the time required for the shift is increased. In any case, when the amount of change in vehicle speed that occurs during shifting is large,
The shift command will be output earlier,
The problem that the vehicle speed or the engine speed at the end of the shift is excessively increased and the sense of incongruity due to a shift in shift feeling are suppressed.

On the other hand, according to the sixth aspect, the so-called hunting of the shift point correction control is prevented, and the smoothness of the control is ensured. That is, as an example, consider a case in which the driver applies a large brake, and then immediately stops the brake operation. In this case, the vehicle acceleration fluctuates greatly within a short period of time, and specifically increases once a large negative acceleration is detected. By the way, according to the control, the shift point is corrected according to the vehicle acceleration. In this case, at the time when a large negative acceleration is first detected, the shift command is output earlier so that the shift command is output earlier. Is changed to a higher vehicle speed side, and as a result, a downshift gearshift command is output, and the gearshift operation starts.

However, when it is detected that the acceleration has increased again, the shift line is returned to the low vehicle speed side. In this case, as shown by the reference numeral c in FIG. 24, the driving state of the vehicle is changed to a downshift shift line (2-1 shift line shown by a chain line) provided to prevent hunting of the shift determination itself. Line) and an upshift shift line (1-2 shift line shown by a solid line), and the 2-1 downshift shift line is corrected to the position indicated by the symbol c due to rapid deceleration. When the 2-1 shift is started and the 2-1 shift line is restored, there is no problem because the first speed is still determined. However, as shown by reference numeral d, the driving state of the vehicle is 1-2.
In the case where the vehicle is on the higher vehicle speed side than the upshift line, and the degree of deceleration is still large, the 2-1 downshift line is corrected to the high vehicle side to the position indicated by the symbol d. When the 2-1 shift is started, if the 2-1 shift line is returned again, the second speed is determined. As a result, immediately after the downshift command is output, the upshift command is executed again. May be output.

The sixth aspect of the present invention addresses this problem, and aims to solve the hunting problem of the shift point correction control by suppressing the detection of the shift point correction amount during the shift operation. is there. Therefore, in the above example, once the setting of the 2-1 downshift shift line is changed to the high vehicle speed side and the 2-1 downshift is being executed,
The return of the 2-1 downshift line to the low vehicle speed side is avoided, and the hunting of the control is eliminated.

According to the seventh aspect, the shift determination is not made based on the shift point predetermined by the throttle opening and the vehicle speed, but based on the shift point predetermined by the engine speed. When performed, the same effects as those of the above-described inventions are obtained, and in particular, the problem of excessive engine speed is directly solved.

[0026]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The embodiments of the present invention will be described below in terms of a mechanical configuration, a hydraulic control circuit, and control operations such as gear shifting.

Mechanical Configuration First, the overall mechanical schematic configuration of the automatic transmission 10 according to the present embodiment will be described with reference to the skeleton diagram of FIG.

The automatic transmission 10 includes, as main components, a torque converter 20 and a transmission gear mechanism driven by an output of the converter 20 (hereinafter, the engine side is referred to as front and the anti-engine side as rear). And a plurality of friction elements 51 to 40 such as clutches and brakes for switching a power transmission path formed by the planetary gear mechanisms 30 and 40.
55 and a one-way clutch 56, whereby the first to fourth speeds in the D range and the first to third speeds in the S range are provided.
The first and second speeds in the speed range and the L range and the reverse speed in the R range are obtained.

The torque converter 20 is a pump 2 fixed in a case 21 connected to the engine output shaft 1.
2 and the pump 22
And a stator 25 interposed between the pump 22 and the turbine 23 and supported by the transmission case 11 via a one-way clutch 24 to increase the torque. And a lock-up clutch 26 provided between the case 21 and the turbine 23 and directly connecting the engine output shaft 1 and the turbine 23 via the case 21. And
The rotation of the turbine 23 is output to the planetary gear mechanisms 30 and 40 via a turbine shaft 27.

Here, behind the torque converter 20, an oil pump 12 driven by the engine output shaft 1 via a case 21 of the torque converter 20 is arranged.

On the other hand, the first and second planetary gear mechanisms 30,
Reference numeral 40 denotes sun gears 31 and 41, and a plurality of pinions 32 ... 32 meshed with the sun gears 31 and 41.
42 ... 42 and these pinions 32 ... 32, 42 ... 4
2 and pinion carriers 33, 43, and ring gears 34 meshed with the pinions 32,.
44.

The turbine shaft 27 and the first
A forward clutch 51 is provided between the planetary gear mechanism 30 and the sun gear 31, a reverse clutch 52 is provided between the turbine shaft 27 and the sun gear 41 of the second planetary gear mechanism 40, and a turbine shaft 27 is provided with the second planetary gear mechanism. A 3-4 clutch 53 is interposed between the pinion carrier 43 and the pinion carrier 40, and a 2-4 brake 54 for fixing the sun gear 41 of the second planetary gear mechanism 40 is provided.

Further, the ring gear 34 of the first planetary gear mechanism 30 and the pinion carrier 43 of the second planetary gear mechanism 40
The low reverse brake 55 and the one-way clutch 56 are arranged in parallel between the transmission case 11 and these components, and the pinion carrier 33 of the first planetary gear mechanism 30 and the second planetary gear mechanism Forty ring gears 44 are connected, and the output gear 13 is connected to them.

The output gear 13 is meshed with the first intermediate gear 62 on the idle shaft 61 constituting the intermediate transmission mechanism 60, and the second intermediate gear 63 on the idle shaft 61 and the differential gear. The input gear 71 of the differential gear 70 meshes with the rotation of the output gear 13 and is input to the differential case 72 of the differential 70.
The left and right axles 73, 74 are driven via the zero.

The relationship between the operating states of the friction elements 51 to 55 such as the clutches and brakes and the one-way clutch 56 and the gears is summarized in Table 1 below.

[0036]

[Table 1] Note that the transmission gear mechanism portion of the automatic transmission 10 shown in the above skeleton diagram is specifically configured as shown in FIG. 2, but as shown in FIG. A turbine rotation sensor 305 used in the control is performed.

This sensor 305 has a forward clutch 51 whose tip rotates integrally with the turbine shaft 27.
The rotation speed of the turbine shaft 27 is detected by detecting a periodic change in a magnetic field generated by a spline provided on the outer peripheral surface of the drum 51a. I have.

Further, the torque converter 20 shown in the skeleton diagram of FIG. 1 is specifically configured as shown in FIG. 3, and will be described in detail.
Numeral 0 denotes a pump 22 composed of a large number of blades integrally provided on the case 21 attached to the engine output shaft 1 on the opposite side to the engine in the case 21, and a half on the engine side in the case 21. A turbine 23 having a plurality of blades, which is rotatably mounted on the case 21 and opposed to the pump 22;
And a stator 25 composed of a large number of blades supported by the transmission case 11 via a one-way clutch 24 and rotatable only in a predetermined direction.
And The boss 23a of the turbine 23 is spline-coupled to the turbine shaft 27, and the rotation of the turbine 23 is taken out to the side opposite to the engine via the turbine shaft 27.

In the case 21, the turbine 2
3 so as to rotate integrally with the turbine 23 and to be slidable in the axial direction with respect to the turbine 23.
Is built-in. This lock-up clutch 26
The case 21 is disposed so as to face the engine-side flat portion 21a, and when the lock-up clutch 26 is fastened to the case flat portion 21a, the case 21
, The engine output shaft 1 and the turbine shaft 27 are connected to each other.

Then, a case 2 is provided by the engine output shaft 1.
When the pump 22 is driven via the
The hydraulic oil on the second side is pushed down to the outer peripheral side by centrifugal force, and as shown by the arrow A, the hydraulic oil flows from the outer peripheral side toward the inner peripheral side into the turbine 23, so that the driving force is applied to the turbine 23. Is given. In that case, the pump rotation speed of the turbine rotation speed Nt (that is, the engine rotation speed Ne)
When the speed ratio e (= Nt / Ne) is equal to or less than a predetermined value, the stator 25 is locked by the one-way clutch 24 to apply a reaction force to the flow of hydraulic oil, thereby increasing the torque on the pump 22 side. Then, the power is transmitted to the turbine 23.

The lock-up clutch 26 is
Due to the pressure of the hydraulic oil in the chamber behind the clutch 26 in the case 21, that is, the rear chamber 26a, the case flat portion 21
a, and is released by an operating pressure supplied to a chamber provided between the lock-up clutch 26 and the case flat portion 21a, that is, a front chamber 26b. The slip state is controlled by adjusting the operating pressure supplied to the front chamber 26b.

Hydraulic control circuit Next, a hydraulic control circuit for supplying and discharging working pressure to and from hydraulic chambers provided in the friction elements 51 to 55 shown in FIGS. 1 and 2 will be described.

Of the friction elements, the 2-4 brake 54, which is a band brake, has a fastening chamber 54a and a release chamber 54b as hydraulic chambers to which operating pressure is supplied, and operates only in the fastening chamber 54a. When pressure is supplied,
-4 The rake 54 is fastened and the operating pressure is supplied only to the release chamber 54b, the operating pressure is not supplied to both the chambers 54a and 54b, and both the chambers 54a and 54b
When the operating pressure is supplied, the 2-4 brake 5
4 is released.

The other friction elements 51 to 53, 55
Has a single hydraulic chamber, and the friction element is fastened when operating pressure is supplied to the hydraulic chamber.

Here, the specific structure of the hydraulic actuator of the 2-4 brake 54 will be described. As shown in FIG. 4, the hydraulic actuator comprises a transmission case 11 and a cover member 54c fixed to the case 11. The piston 54e is fitted into the servo cylinder 54d composed of the above, and the above-described fastening chamber 54a and release chamber 54b are formed on both sides thereof. In addition, the piston 54e
A band fastening stem 54f is attached to the stem, and the stem 54f is engaged with one end of a brake band 54g wound around a member to be braked (not shown), and a case is attached to the other end. 11 is engaged with the fixing stem 54h.
4b, a spring 54i for urging the piston 54e toward the fastening chamber 54a, that is, toward the loosening side of the brake band 54g.
Is stored.

Operating pressure is supplied from a control valve unit constituting a hydraulic control circuit to the fastening chamber 54a and the release chamber 54b through an oil hole (not shown). By tightening or loosening, the 2-4 brake 54 is engaged or released. In particular, in this hydraulic actuator, the pressure receiving areas of the piston 54e on the fastening chamber 54a side and the release chamber 54b side are substantially equal. Therefore, for example, when an equal operating pressure is supplied to both chambers 54a and 54b, these pressures cancel each other, and only the urging force of the spring 54i acts on the release side.

(1) Overall Configuration As shown in FIG. 5, the hydraulic control circuit 100 includes, as main components, a regulator valve 101 for generating line pressure, and a manual valve for manually switching a range. 102, a low reverse valve 103, a bypass valve 104, a 3-4 shift valve 105, and a lock-up control valve 106, which operate at the time of gear shifting to switch oil passages leading to the friction elements 51 to 55, and these valves 103 to 106 First and second ON-OFF solenoid valves (hereinafter, referred to as “first and second SVs”) 111 and 112 for operation, and a solenoid relay valve (hereinafter referred to as “relay”) for switching the supply destination of the operating pressure from the first SV 111 "Valve") 1
07, and first to third duty solenoid valves (hereinafter, referred to as “first to third DSVs”) that control generation, adjustment, discharge, and the like of the operating pressure supplied to the hydraulic chambers of the friction elements 51 to 55.
121, 122, 123, etc. are provided.

Here, the first and second SVs 111, 11
Each of the second and first to third DSVs 121 to 123 is a three-way valve, and can obtain a state in which the upper and downstream oil paths are communicated and a state in which the downstream oil path is drained. ing. In the latter case, the oil passage on the upstream side is shut off, so that the operating oil from the upstream side is not drained out in a drain state, and the drive loss of the oil pump 12 is reduced.

Note that the first and second SVs 111 and 112 are O
At the time of N, the upper and downstream oil passages are communicated. Also, the first
When the third DSVs 121 to 123 are OFF, that is, when the duty ratio (the ratio of the ON time in one ON-OFF cycle) is 0%, the third DSVs 121 to 123 are fully opened to completely communicate the upper and downstream oil passages, and When the duty ratio is 10
At 0%, the oil path on the upstream side is shut off to cause the oil path on the downstream side to be in a drain state. At an intermediate duty ratio, the hydraulic pressure on the upstream side is used as the original pressure, and the duty ratio on the downstream side is reduced. An oil pressure adjusted to a corresponding value is generated.

The regulator valve 101 adjusts the pressure of the hydraulic oil discharged from the oil pump 12 to a predetermined line pressure. This line pressure is supplied to the manual valve 102 via the main line 200, and is supplied to a solenoid reducing valve (hereinafter, referred to as a solenoid reducing valve).
(Reducing valve) 108) and the 3-4 shift valve 105.

The line pressure supplied to the reducing valve 108 is reduced to a constant pressure by the valve 108, and then supplied to the first and second SVs 111 and 112 via lines 201 and 202. .

When the first SV 111 is ON, the constant pressure is supplied to the relay valve 107 via the line 203, and when the spool of the relay valve 107 is located on the right side in the drawing (the same applies hereinafter). Is
Further, a pilot pressure is supplied to a control port at one end of the bypass valve 104 via the line 204 to urge the spool of the bypass valve 104 to the left. When the spool of the relay valve 107 is located on the left side, the 3-4 shift valve 105
Is supplied as pilot pressure to the control port at one end of
The spool of the 3-4 shift valve 105 is biased to the right.

When the second SV 112 is ON,
The constant pressure from the reducing valve 108 is supplied to the bypass valve 104 via the line 106. When the spool of the bypass valve 104 is located on the right side, the lock-up control valve is further connected via the line 207. The control valve 106 is supplied as pilot pressure to a control port at one end of the control valve 106.
Bias the spool to the left. Also, bypass valve 1
When the spool No. 04 is located on the left side, the line 208
Is supplied as pilot pressure to the control port at one end of the low reverse valve 103 to urge the spool of the low reverse valve 103 to the left.

Further, the constant pressure from the reducing valve 108 is also supplied to the control port 101a of the regulator valve 101 via the line 209. In this case, the constant pressure is adjusted by the linear solenoid valve 131 provided on the line 209 according to, for example, the throttle opening of the engine. Therefore, the line pressure is adjusted by the regulator valve 101 according to the throttle opening and the like. Will be adjusted.

The main line 200 led to the 3-4 shift valve 105 is connected to the first line via the line 210 when the spool of the valve 105 is located on the right side.
The accumulator 141 communicates with the accumulator 141.
To introduce line pressure.

On the other hand, the line pressure supplied from the main line 200 to the manual valve 102 is D, S, L
Are introduced into the first output line 211 and the second output line 212 in each forward range, into the first output line 211 and the third output line 213 in the R range, and into the third output line 213 in the N range.

The first output line 211 is connected to the first output line 211.
It is led to the DSV 121 and supplies the first DSV 121 with a line pressure as a control source pressure. The downstream side of the first DSV 121 is connected to a low reverse valve 1 via a line 214.
03, and when the spool of the valve 103 is located on the right side, a further line (servo apply line)
It is guided to the engagement chamber 54a of the 2-4 brake 54 via 215. When the spool of the low reverse valve 103 is located on the left side, the spool is further guided to the hydraulic chamber of the low reverse brake 55 via a line (low reverse brake line) 216.

Here, from the line 214, the line 2
17 is branched and led to the second accumulator 142.

The second output line 212 is connected to the second output line 212.
The line pressure is supplied to the DSV 122 and the third DSV 123, the line pressure is supplied to these DSVs 122 and 123 as the control source pressure, and the line pressure is also supplied to the 3-4 shift valve 105.

The line 212 led to the 3-4 shift valve 105 is led to the lock-up control valve 106 via the line 218 when the spool of the valve 105 is located on the left side. Is located on the left side, and is further led to the hydraulic chamber of the forward clutch 51 via a line (forward clutch line) 219.

Here, the forward clutch line 2
The line 220 branched from 19 is led to the 3-4 shift valve 105. When the spool of the valve 105 is located on the left side, the line 220 communicates with the first accumulator 141 via the aforementioned line 210, and the valve 105 When the spool is located on the right side, it communicates with the release chamber 54b of the 2-4 brake 54 via the line (servo release line) 221.

The downstream side of the second DSV 122 to which the control source pressure is supplied from the second output line 212 is connected to the line 22.
The pilot pressure is supplied to the control port at one end of the relay valve 107 via the control port 2 to supply the pilot pressure to the port, thereby urging the spool of the relay valve 107 to the left. The line 22 branched from the line 222
3 is led to a low reverse valve 103, and the valve 10
When spool 3 is on the right, line 2
Leads to 24.

From this line 224, the orifice 15
1, the line 225 is branched, and the branched line 225 is led to the 3-4 shift valve 105. When the spool of the 3-4 shift valve 105 is located on the left side, the servo Release line 221
Through the release chamber 54b of the 2-4 brake 54.

Also, the orifice 1
A line 226 is further branched from a line 225 branched through the line 51, and the line 226 is further branched.
Is guided to the bypass valve 104, and is guided to the hydraulic chamber of the 3-4 clutch 53 via the line (3-4 clutch line) 227 when the spool of the valve 104 is located on the right side.

Further, the line 224 is directly led to the bypass valve 104, and communicates with the line 225 via the line 226 when the spool of the valve 104 is located on the left side. That is, the line 224 and the line 225
Are connected to bypass the orifice 151.

The downstream side of the third DSV 123 to which the control source pressure is supplied from the second output line 212 is connected to the line 22.
8 and is led to the lock-up control valve 106, and when the spool of the valve 106 is located on the right side, it communicates with the forward clutch line 219.
When the spool of the lock-up control valve 106 is located on the left side, it communicates with the front chamber 26 b of the lock-up clutch 26 via the line 229.

Further, a third output line 213 from the manual valve 102 is led to the low reverse valve 103 to supply a line pressure to the valve 103. When the spool of the valve 103 is located on the left side, it is guided to the hydraulic chamber of the reverse clutch 52 via a line (reverse clutch line) 230.

A line 231 branched from the third output line 213 is led to the bypass valve 104, and when the spool of the valve 104 is located on the right side, the control of the low reverse valve 103 via the line 208 is performed. The line pressure is supplied to the port as the pilot pressure, and the spool of the low reverse valve 103 is urged to the left.

In addition to the above configuration, the hydraulic control circuit 1
00 is provided with a converter relief valve 109. This valve 109 is a regulator valve 10
After the working pressure supplied from 1 through the line 232 is regulated to a constant pressure, this constant pressure is supplied to the lock-up control valve 106 via the line 233. This constant pressure is applied to the lock-up control valve 1
When the spool of the valve 106 is located on the right side, it is supplied to the front chamber 26b of the lock-up clutch 26 via the aforementioned line 229. When the spool of the valve 106 is located on the left side, the rear chamber is provided via the line 234. 26a.

Here, the lock-up clutch 26 is released when the above-mentioned constant pressure is supplied to the front chamber 26b, but the spool of the lock-up control valve 106 is located on the left side, and 3DSV1
When the operating pressure generated at 23 is supplied to the front chamber 26b, a slip state is set, and the slip amount is controlled according to the operating pressure.

(2) Circuit state for each gear stage On the other hand, as shown in FIG. 6, the automatic transmission 10 has the first and second SVs 111,
112, a controller 300 for controlling the first to third DSVs 121 to 123 and the linear solenoid valve 131, and the controller 300 includes:
A vehicle speed sensor 301 for detecting the vehicle speed of the vehicle, a throttle opening sensor 30 for detecting the throttle opening of the engine
2. Engine rotation sensor 30 for detecting engine speed
3. Shift position sensor 304 for detecting a shift position (range) selected by the driver, torque converter 2
0, a turbine rotation sensor 305 for detecting the rotation speed of the turbine 23, and an oil temperature sensor 3 for detecting the oil temperature of the working oil.
06, etc., and these sensors 301 to
Each of the solenoid valves 111, 11 according to the operation state of the vehicle or the engine indicated by the signal from 306, etc.
2, 121 to 123, 131 are controlled. FIG. 2 shows the turbine rotation sensor 305 in an attached state.

Next, the first and second SVs 111, 112
The relationship between the operating states of the first to third DSVs 121 to 123 and the supply and discharge states of the operating pressure of the friction elements 51 to 55 to and from the hydraulic chamber will be described for each shift speed.

Here, the combinations (solenoid patterns) of the operating states of the first and second SVs 111 and 112 and the first to third DSVs 121 to 123 for each gear are set as shown in Table 2 below.

In Table 2, (○) indicates the first and second SV1.
ON for 11 and 112, 1st to 3rd DSV 121
Reference numerals 123 to 123 are OFF, and all indicate a state in which the upstream oil passage is communicated with the downstream oil passage and the original pressure is supplied to the downstream as it is. (×) indicates the first and second
OFF for SV111 and 112, 1st to 3rd DS
V121 to V123 are ON.
This shows a state in which the upstream oil passage is shut off and the downstream oil passage is drained.

[0075]

[Table 2] (2-1) 1st gear First, in 1st gear (excluding 1st gear in L range), Table 2
As shown in FIG. 7, only the third DSV 123 operates to generate an operating pressure using the line pressure from the second output line 212 as a source pressure, and this operating pressure is supplied via a line 228 to the lock-up control valve. 106. At this point, the spool of the lock-up control valve 106 is located on the right side,
The operating pressure further increases the forward clutch line 219
Is supplied to the hydraulic chamber of the forward clutch 51 as a forward clutch pressure, whereby the forward clutch 51 is engaged.

Here, the forward clutch line 2
19 is connected to the first accumulator 1 via the 3-4 shift valve 105 and the line 210.
Due to the communication with 41, the supply of the forward clutch pressure is performed gently.

(2-2) Second speed Next, in the second speed state, as shown in Table 2 and FIG. 8, in addition to the above-mentioned first speed state, the first DSV 121 also operates.
An operating pressure is generated using the line pressure from the first output line 211 as a source pressure. This operating pressure is supplied to the low reverse valve 103 via the line 214. At this time, the servo release line 215 is further provided because the spool of the low reverse valve 103 is located on the right side.
And supplied to the fastening chamber 54a of the 2-4 brake 54 as servo apply pressure. Thus, the 2-4 brake 54 is engaged in addition to the forward clutch 51.

Since the line 214 communicates with the second accumulator 142 via the line 217, the supply of the servo apply pressure or the application of the 2-4 brake 54 is performed gently. When the spool of the low reverse valve 103 moves to the left when shifting to the first speed in the L range, which will be described later, the hydraulic oil stored in the accumulator 142 is transmitted from the low reverse brake line 216 to the hydraulic pressure of the low reverse brake 55. The room is precharged.

(2-3) Third speed In the third speed state, as shown in Table 2 and FIG. 9, in addition to the above-mentioned second speed state, the second DSV 122 also operates, and the second output line 212 The operating pressure is generated by using the line pressure from the pressure source as the original pressure. This operating pressure is supplied to the low reverse valve 103 through the line 222 and the line 223. At this point, the spool of the valve 103 is also located on the right side, so that the line 224 is further moved.
Will be introduced.

This operating pressure is introduced from the line 224 to the line 225 via the orifice 151,
At this time, since the spool of the 3-4 shift valve 105 is located on the left side, the servo release line 221 is further moved.
Is supplied as a servo release pressure to the release chamber 54b of the 2-4 brake 54 via the. Thereby, the 2-4 brake 54 is released.

Also, the orifice 1
From line 225 branched through 51, line 22
6 is branched, the operating pressure is increased by the line 226.
At this time, the spool of the bypass valve 104 is located on the right side at this point, and the 3-4 clutch line 227
Is supplied to the hydraulic chamber of the 3-4 clutch 53 as a 3-4 clutch pressure. Therefore, in the third speed, the forward clutch 51 and the 3-4 clutch 53 are engaged, while the 2-4 brake 54 is released.

In the third speed, the second DSV 122 generates the operating pressure as described above,
The spool of the relay valve 107 is moved to the left by being supplied to the control port 107a of the relay valve 107 via 2.

(2-4) Fourth gear Further, in the fourth gear state, as shown in Table 2 and FIG. 10, the third DSV 123 stops generating the operating pressure and the first SV111 Operates.

By the operation of the first SV 111, a constant pressure from the line 201 is supplied to the relay valve 107 via the line 203. As described above, the spool of the relay valve 107 is moved to the left at the third speed. , The constant pressure is
This is supplied to the control port 105a of the four-shift valve 105, and the spool of the valve 105 moves to the right. Therefore, the servo release line 221 is divided into the line 22 branched from the forward clutch line 219.
0, and the release chamber 54b of the 2-4 brake 54 communicates with the hydraulic chamber of the forward clutch 51.

Then, as described above, the third DSV 123 stops generating the operating pressure and sets the downstream side to the drain state, whereby the servo release pressure in the release chamber 54b of the 2-4 brake 54 and the forward clutch 51 Of the third DSV via the lock-up control valve 106 and the line 228.
It will be drained at 123.
The brake 54 is engaged again, and the forward clutch 51 is released.

Control Operation Next, the characteristic part of the present invention in the specific operation of the control by the controller 300 will be described.

The controller 300 stores a shift control map as conceptually shown in FIG. 11, and performs 1-2 shift, 2 shift from the low vehicle speed side in accordance with the throttle opening and the vehicle speed. 3-3 shift and 3-4 shift upshift lines, and from the high vehicle speed side, 4-3 shift, 3-2 shift
The shift and the downshift lines of the 2-1 shift are preset. In this case, when the throttle opening is substantially fully open (8/8), each shift line has only the vehicle speed,
That is, it is determined only by the engine speed.

Then, the controller 300 applies the vehicle speed detected by the vehicle speed sensor 301 and the throttle opening detected by the throttle opening sensor 302 to the shift map to determine the optimum gear position of the vehicle to be realized. Is determined. In this case, the controller 300 corrects the shift point based on the traveling state of the vehicle so that each shift ends substantially on the shift line. The description will be given later, and an outline of the shift control performed by the controller 300 will be described first, taking a 1-2 upshift as an example.

The shift control of the upshift is basically performed by
As shown in FIG. 12 is performed by the change rate dNt during lowering of the turbine speed Nt to match the change rate dNt ₀ goal, feedback control of the supply of hydraulic pressure to the frictional elements mainly engagement side .

This turbine rotation change rate dNt Figure 13
As shown in the figure, during the inertia phase during the shifting operation,
Of the output torque of the transmission
Corresponding to the height ΔTr, which is the torque before shifting.
The higher the gear, the greater the shift shock and the lower
If too long, the shift time becomes longer. So, as shown
This height ΔTr is set to be substantially equal to the height before shifting.
Target turbine rotation change rate dNt corresponding to₀Set
It is. Specifically, the throttle opening of the engine is large.
The turbine speed at the start of gear shifting to a significantly higher rate of change
The higher the Nt is, the smaller the change rate is set.

(1) 1-2 Shift In the 1-2 shift, the servo apply pressure is generated by the first DSV 121 and is applied to the engagement chamber 54 of the 2-4 brake 54.
a during which the feedback control of the servo apply pressure by the first DSV 121 is performed.

Here, as described above, each DSV 121-
Reference numeral 123 denotes a drain state in which no operating pressure is generated at a duty ratio of 100%, a fully open state in which the operating pressure is equal to the original pressure at 0%, and control of the operating pressure at an intermediate duty ratio.

(1-1) Control of 1st DSV The control of the servo apply pressure by the 1st DSV 121 at the time of this 1-2 shift is performed according to the program shown in FIG. 14, and when the 1-2 shift command is output, First,
In steps S1 to S3, a hydraulic pressure Pb serving as a base portion of the servo apply pressure supplied at the time of controlling the turbine speed,
Similarly, a feedback hydraulic pressure Pfb that is feedback-controlled at the time of controlling the turbine rotational speed and a learning control hydraulic pressure Pad for making the next shift control more precise are calculated.

Then, in step S4, these hydraulic pressures P
b, Pfb, and Pad are added to calculate the calculated oil pressure Ps, and in step S5, the value of the precharge flag Fp is used to determine whether or not the servo apply pressure precharge control period performed at the time of shifting command output is being performed. Judgment based on

As described later, this precharge control is to quickly fill the oil passage to the engagement chamber 54a of the 2-4 brake 54 and the hydraulic oil in the engagement chamber 54a at the start of the gear shift, thereby improving the responsiveness of the gear shifting operation. To improve
When p = 1, that is, during a precharge period set according to a program shown in FIG. 16 described later, a signal having a duty ratio of 0% is output to the first DSV 121 in step S6. Further, when Fp = 0, that is, when the precharge period ends, further, step S
At 7, it is determined whether or not the 1-2 shift has been completed. The end of the shift is determined when the turbine rotation change rate dNt changes from minus to plus, and when the turbine rotation change rate dNt
Either the absolute value of the engine speed has decreased to half or less of the value during the shift, or the turbine speed Nt has decreased to the speed at the end of the shift calculated from the speed at the start of the shift. Done by

Before the shift is completed, that is, after the end of the precharge period and before the end of the shift, in step S8,
The duty ratio of the first DSV 121 is determined so that the calculated hydraulic pressure Ps determined as described above is obtained, and a signal of the duty ratio is output, and the servo apply pressure or the engagement force of the 2-4 brake 54 is controlled. . Further, after the shift is completed, in steps S9 and S10, the duty ratio is output while being subtracted at a constant rate until the duty ratio becomes 0%.

[0097] By the above control, obtained servo apply pressure changes as shown in FIG. 15, 1-2 as the turbine speed change rate dNt during the shift coincides with the target turbine speed change rate dNt _0, 2- The engagement force of the four brakes 54 is controlled.

(1-2) Precharge Period Setting Control Next, the setting of the precharge flag Fp whose value is determined in step S5 of the program in FIG. 14, that is, the control of setting the precharge period, will be described.

As described above, in this 1-2 shift,
The hydraulic oil is supplied to the engagement chamber 54a of the 4 brake 54 to
The second shift stage is achieved by engaging the -4 brake 54. At this time, the turbine rotation change rate is controlled by feedback controlling the servo apply pressure supplied to the engagement chamber 54a.

However, in this 1-2 shift, the fastening chamber 54a of the 2-4 brake 54 and the oil passages 214 and 215 leading to the fastening chamber 54a are provided with the first gear before shifting.
In the fast state, there is no hydraulic oil and therefore
Immediately after the output of the shift command, it is not possible to shift to the feedback control of the servo apply pressure. As a result, 2-
The engagement operation of the fourth brake 54 is delayed, and the responsiveness of the 1-2 shift is reduced.

Therefore, when the 1-2 shift command is output, the controller 300 fully opens the first DSV 121 for controlling the supply of the operating pressure to the 2-4 brake 54 for a predetermined time in order to eliminate the shift response delay. In this state, precharge control is performed to quickly fill the hydraulic path to the hydraulic chamber 54a of the 2-4 brake 54 and the hydraulic oil into the hydraulic chamber 54a.

The setting control of the precharge period in the precharge control is performed according to the program shown in FIG. This program is executed when the shift command is output as shown in FIG.
4 is executed in parallel with the program No. 4. That is, first, in step S11, the total flow rate Qt is set to 0 as initialization, and then in step S12
Then, based on the map set as shown in FIG.
The base flow rate Q passing through the DSV 121 when the first DSV 121 is fully opened (duty ratio 0%) is obtained from the line pressure at that time. In that case, the map above
The base flow rate Q is set to increase as the line pressure increases. This is because even when the first DSV 121 is fully opened, the flow rate Q of the hydraulic oil passing therethrough changes according to the line pressure at that time, This is because the flow rate Q increases as the line pressure increases.

Next, in step S13, the oil temperature correction coefficient K is read from the map set as shown in FIG.
In the oil temperature correction coefficient map, the correction coefficient K is set to be smaller than 1 as the temperature of the hydraulic oil decreases.

Then, in step S14, a base flow rate Q 'whose temperature has been corrected by multiplying the base flow rate Q by a correction coefficient K is obtained. Thus, when the temperature of the hydraulic oil is low, and thus the viscosity is high, the flow rate through the valve is reduced from the standard state even at the same line pressure.
The flow rate calculated in accordance with the actual situation is also reduced, and the base flow rate Q 'suitable for the actual flow rate is always calculated.

Further, in step S15, the corrected base flow rate Q 'is integrated for each cycle of the program, and the total flow rate Qt from the start of control to the present time is calculated.

Next, in step S16, it is determined whether or not the total flow rate Qt has exceeded a predetermined value V. Until the total flow rate Qt exceeds the predetermined value V, the precharge flag Fp is set to 1 in step S17. When the value exceeds the predetermined value V, the flag Fp is set to 0 in step S18.

In this case, the predetermined value V is the capacity of the hydraulic oil required to make a stroke until the friction element is in a state immediately before the engagement, that is, in the case of this 1-2 shift, FIG. 2-4 Servo cylinder 5 of brake 54 shown
It is set to a value substantially corresponding to the volume of hydraulic oil required to stroke the piston 54e in 4d toward the release chamber 54b against the urging force of the spring 54i until immediately before the brake band 54g starts to work. Therefore, Q> V
At this time, the turbine speed control enters the standby state, and at this time, the flag Fp is set to 0 to end the precharge control.

Using the value of the precharge flag Fp set in this way, during Fp = 1, that is, during the precharge period, step S6 of the program of FIG.
Then, the control for setting the duty ratio of the first DSV 121 to 0% is performed, whereby the engagement chamber 5 of the 2-4 brake 54 is controlled.
The oil passages 214 and 215 leading to 4a and the fastening chamber 54a are quickly filled with hydraulic oil.

(2) Other shifts In the case where the precharge control is performed on the engagement-side friction element during the shift, such as the above-mentioned 1-2 upshift,
In addition, a 2-3 upshift with a 3-4 clutch 53 engaged, a 3-4 upshift with a 2-4 brake 54 engaged, and a 4 with the forward clutch 51 engaged 4
-3 downshift, 2-4 brake 54 is engaged, and 3-2 downshift.

In this embodiment, the controller 300 operates the engagement side frictional element in the case of a manual downshift or coast downshift in which the throttle opening is substantially fully closed during the downshift. Pressure control is performed. On the other hand, in the case of a torque demand downshift that is performed with an increase in the throttle opening, the operating pressure control for the release-side friction element is performed. Is a downshift that is performed when the throttle opening is substantially fully closed.

Here, 3-4 upshifting and 3
For -2 downshifts, 2-4 brakes 5
The engagement operation is performed by discharging hydraulic oil in the release chamber 54b from a state in which hydraulic pressure is supplied to both the engagement chamber 54a and the release chamber 54b of the 2-4 brake 54. Piston 5 in servo cylinder 54d
4e against the urging force of the spring 54i.
It is to make a stroke to the side. Therefore, 2-4
Oil passages 214 and 21 communicating with the fastening chamber 54a of the brake 54
5 is filled with hydraulic oil from the beginning, and the predetermined value V in step S16 in the program of FIG.
The value is set to a value substantially corresponding to the capacity of 4a.

It is to be noted that only the 2-1 downshift has a
-4 Since the hydraulic oil is only discharged from the engagement chamber 54a of the brake 54 and there is no engagement operation of the friction element, the shift operation for performing the precharge control is not started.

(3) Shift Point Correction Control Next, the shift point correction control performed by the controller 300 will be described. In this control, when it is predicted that a shift in which the above-described precharge control is performed will occur next, the shift is always a constant vehicle speed, specifically, FIG.
The shift point is corrected so as to end on the shift line in the shift map shown in FIG.

This correction control is constantly performed during running of the vehicle according to the program shown in FIGS. 19 and 20. First, in step S21, for example, the current vehicle speed is calculated from the rate of change of the vehicle speed read every cycle of this program. Is calculated.

Next, in step S22, it is determined whether or not the shift control is being performed. If not, the process proceeds to step S23 to update the vehicle acceleration DVSP. That is, according to the equation in the figure, a so-called smoothing process is performed from the current vehicle acceleration DVSP1 and the previous vehicle acceleration value DVSP to obtain the current vehicle acceleration value DVSP. On the other hand, when the shift control is being performed, such vehicle acceleration DVS
The process proceeds to step S24 without updating P. Note that the smoothing coefficient α in the equation in the figure is a positive value less than 1.

In step S24, the vehicle acceleration DV
It is determined whether SP is a positive value. And if yes,
That is, when the vehicle speed is increasing, the process proceeds from step S25 to S30. In the case of NO, that is, when the vehicle speed is decreasing, step S3 is performed.
The process proceeds from S1 to S34 to predict an upshift or a downshift that is supposed to occur next based on the current gear. In the case of the downshift, the reason why the 2-1 shift is not predicted is as described above.
This is because only the 2-1 downshift has no engagement operation of the friction element and does not perform the precharge control.

Next, in step S35, the predicted line pressure P0 at the time of the shift is predicted from the throttle opening using the map shown in FIG. This is because the gearshift operation has not yet been performed and the line pressure at the time of gearshift cannot be directly detected.

The next steps S36 to S37 are the same as those in the precharge control shown in FIG.
Based on the maps set as shown in FIG. 22 or FIG. 23, respectively, the precharge base flow rate Q0 is calculated from the predicted line pressure P0 at the time of shifting, and the temperature correction coefficient K1 is further determined. The read and the temperature-corrected precharge base flow rate Q1 is obtained. At this time, depending on the type of shift, the duty solenoid valves 121, 12 used when precharging are used.
In some cases, the precharge base flow rate Q0 determined in step S36 is always fully open (duty solenoid valves 121, 122, 123). (The rate is 0%). The flow rate map according to the type of shift is stored in the controller 300.

Next, in step S38, the controller 300 obtains the predicted stroke time T0 of the friction element at the time of the shift. That is, the time T0 required for the friction element to make a stroke until the state immediately before the engagement is set to:
The capacity of the oil passage and the hydraulic chamber communicating with the friction element (in the case of 3-4 upshift and 3-2 downshift, the capacity of the engagement chamber 54a of the 2-4 brake 54) and the above temperature correction were performed. It is determined from the precharge base flow rate Q1 according to the equation in the figure.

Next, in step S39, the controller 300 multiplies the stroke time T0 thus obtained by the vehicle acceleration DVSP to obtain a change amount of the vehicle speed, and sets this as a vehicle speed correction amount VSP '. Step S
At 40, the vehicle speed correction amount VSP 'is changed to the vehicle speed F for shifting determination.
By subtracting from the VSP, a corrected shift determination vehicle speed FVSP 'is obtained. In this case, the shift determination vehicle speed F
The VSP is a function of the throttle opening for each shift so that each shift line in the shift map shown in FIG. 11 is obtained.

The controller 300 determines the shift speed based on the shift determination vehicle speed FVSP 'thus corrected.

According to the shift point correction control described above, the absolute value of the vehicle speed correction amount VSP 'is calculated to be a larger value in step S39 as the temperature is lower and the absolute value of the acceleration DVSP is larger. In S40, the shift determination vehicle speed FVSP is corrected to a lower vehicle speed side during an upshift, and the shift determination vehicle speed F is corrected during a downshift.
The VSP is corrected to a higher vehicle speed side, and in any case, the shift command is output earlier, so that the shift ends on the shift line and the vehicle speed or the engine speed over at the end. In addition, the feeling of strangeness in shifting is suppressed.

In this case, when it is determined in step S22 that the shift control is being performed, the vehicle acceleration DVSP is not updated and the shift point correction amount VSP 'is obtained using the previous value of the vehicle acceleration DVSP. Thus, the same shift determination vehicle speed FVSP 'as the previous time is calculated. Therefore, the output of another new shift command during the shift control is suppressed, and the hunting problem of the correction control is solved.

[0124]

As described above, according to the first aspect of the present invention, when the gear position is determined based on the shift point predetermined by the throttle opening and the vehicle speed, the shift ends at the shift point. So that at least according to the vehicle acceleration,
Since the shift point is corrected, the shift can be ended at a fixed shift point, and the problems of over rotation of the engine and poor shift feeling can be solved.

In this case, according to the second aspect, the shift point is corrected in accordance with the vehicle acceleration and the time required for the shift, so that the amount of change in the vehicle speed until the end of the shift is detected.
As a result, the corrected shift point can be determined by back calculation from the target vehicle speed at the end of the shift.

According to the third aspect of the invention, the second aspect of the invention is embodied in more detail, and the type of shift that occurs next is detected by the first detector, and the time required for the detected shift is detected by the second detector. Is detected by the third detection unit, the amount of change in the vehicle speed occurring during the shift is detected as a vehicle speed correction amount, and the correction unit corrects the shift point using the vehicle speed correction amount, and executes these processes. After that, the shift ends at a certain shift point.

In this case, according to the fourth and fifth aspects, the correction amount of the shift point is increased as the absolute value of the vehicle acceleration is increased or as the time required for the shift is increased. In any case, when the amount of change in vehicle speed that occurs during shifting is large,
The shift command will be output earlier,
The problem that the vehicle speed or the engine speed at the end of the shift is excessively increased and the sense of incongruity due to a shift in shift feeling are suppressed.

On the other hand, according to the sixth aspect, during the shifting operation,
By suppressing the detection of the shift point correction amount, so-called hunting of the shift point correction control is prevented, and the smoothness of the control is ensured.

According to the seventh aspect, the shift determination is not made based on the shift point predetermined by the throttle opening and the vehicle speed, but based on the shift point predetermined by the engine speed. When performed, the same effects as those of the above-described inventions are obtained, and in particular, the problem of excessive engine speed is directly solved.

[Brief description of the drawings]

FIG. 1 is a skeleton diagram showing a mechanical configuration of an automatic transmission according to an embodiment of the present invention.

FIG. 2 is a sectional view showing a configuration of a transmission gear mechanism of the automatic transmission.

FIG. 3 is a cross-sectional view illustrating a configuration of a torque converter unit.

FIG. 4 is a sectional view showing a configuration of a hydraulic actuator of a 2-4 brake.

FIG. 5 is a circuit diagram of a hydraulic control circuit.

FIG. 6 is a control system diagram of the automatic transmission.

FIG. 7 is an enlarged circuit diagram of a main part showing a state of a first speed of the hydraulic control circuit of FIG. 5;

FIG. 8 is an enlarged circuit diagram of a main part showing a state of the second speed in the same manner.

FIG. 9 is an enlarged circuit diagram of a main part showing a state of the third speed in the same manner.

FIG. 10 is a main part enlarged circuit diagram showing a state of the fourth speed in the same manner.

FIG. 11 is a conceptual diagram of a map used for shift determination.

FIG. 12 is an explanatory diagram of a turbine rotation change rate as a control target at the time of an upshift.

FIG. 13 is an explanatory diagram of a torque waveform during an upshift.

FIG. 14 is a flowchart showing an operation of a first DSV at the time of a 1-2 shift.

FIG. 15 is a time chart showing changes in data at the time of 1-2 shift.

FIG. 16 is a flowchart showing an operation of precharge control at the time of a 1-2 shift.

FIG. 17 is a map of a base flow rate used in the control operation.

FIG. 18 is also a map of an oil temperature coefficient.

FIG. 19 is a flowchart showing an operation of shift point correction control.

FIG. 20 is the same flowchart.

FIG. 21 is a map of a line pressure prediction used in the control operation.

FIG. 22 is a map of a base flow rate used in the control operation.

FIG. 23 is also a map of an oil temperature coefficient.

FIG. 24 is an explanatory diagram of a problem to be solved by the present invention.

[Explanation of symbols]

 Reference Signs List 10 automatic transmission 30 transmission gear mechanism 300 controller 301 vehicle speed sensor 302 throttle opening sensor

──────────────────────────────────────────────────の Continued on front page (51) Int.Cl. ⁶ Identification code FI F16H 63:12

Claims (7)

[Claims] A vehicle speed detecting means for detecting a value relating to a vehicle speed and a throttle opening detecting means for detecting a value relating to a throttle opening are provided. A control device for an automatic transmission that determines a shift speed based on a detection result of a detecting unit and executes a shift to the shift speed, wherein at least the shift point of the vehicle is adjusted so that the shift ends at the shift point. An automatic transmission, comprising: a correction unit that corrects according to the acceleration of the vehicle; and a shift execution unit that executes a shift based on the corrected shift point and the detection results of the two detection units. Control device. 2. An acceleration detecting means for detecting a value relating to the acceleration of the vehicle, and a shift time calculating means for obtaining a value relating to a time required for shifting, wherein the correcting means includes a detection result of the acceleration detecting means and a shift time calculating. 2. The control device for an automatic transmission according to claim 1, wherein the shift point is corrected according to a calculation result of the means. 3. A first detecting means for detecting whether the next shift is an upshift or a downshift from a vehicle acceleration.
Detecting unit, based on the line pressure during shifting predicted from the throttle opening, the capacity of hydraulic oil to be supplied or discharged to or from the friction element to be engaged or released in the next shift, and the engagement or connection of the friction element from the oil temperature A second detection unit that detects a time required for release,
It has a third detection unit that detects a vehicle speed correction amount at a shift point from the time required for the engagement or release and the vehicle acceleration, and a correction unit that corrects the shift point using the vehicle speed correction amount. The control device for an automatic transmission according to claim 2. 4. The control device for an automatic transmission according to claim 3, wherein the correction means increases the correction amount of the shift point when the absolute value of the vehicle acceleration is large as compared with when the absolute value of the vehicle acceleration is small. . 5. The automatic transmission according to claim 3, wherein the correction means increases the correction amount of the shift point when the time required to fasten or release the friction element is longer than when the time is short. Machine control device. 6. The control device for an automatic transmission according to claim 3, wherein the correction means suppresses detection of a correction amount of a shift point during a shift operation. 7. An engine speed detecting means for detecting a value relating to the engine speed, wherein a shift speed is determined based on a shift point predetermined by the engine speed and a detection result of the detecting means, and What is claimed is: 1. A control device for an automatic transmission for performing a shift to a shift speed, comprising: a correction unit that corrects a shift point according to at least an acceleration of a vehicle so that the shift ends at the shift point; A control device for an automatic transmission, comprising: a shift execution unit that executes a shift based on a point and a detection result of the detection unit.