JPH0477181B2 - - Google Patents

Description

[Detailed description of the invention]

(Industrial Application Field) The present invention relates to a control device for an automatic transmission. (Prior Art) An automatic transmission has a plurality of gear selection friction members built therein, and a predetermined gear is set by appropriately controlling the operation of each gear selection friction member. It's becoming like that. In an automatic transmission with such an operating configuration, especially when downshifting is performed by simultaneous reverse operation of a plurality of gear selection friction members, for example, the D (drive) range (1st gear←→ 4
3 speed), as when shifting from 3rd to 2nd speed.
If the front clutch is engaged and the second brake is released in 2nd gear, but the front clutch is released and the second brake is engaged in 2nd gear, this may occur due to the following reasons. There is a risk of gear shift shock occurring. That is, when shifting down from 3rd gear to 2nd gear, the front clutch is released from the engaged state,
In addition, the second brake is engaged from the released state, but in this case, as the front clutch is released, the turbine rotation speed of the torque converter changes from the rotation speed N ₃ in 3rd gear to the immediate rotation speed N 2 as shown in Figure 4. ₂
The rotational speed increases by an amount corresponding to the gear ratio in 3rd and 2nd gears. Therefore, after releasing the front clutch,
If the engagement of the second brake is completed when the turbine rotation reaches the rotation speed _N2 in second gear, a smooth gear change will be performed as shown in the rotation speed characteristics shown in Fig. 4.
This means fewer gear shifting shocks. However, the rate of change in the turbine rotational speed is not constant, but changes depending on various operating conditions that change from moment to moment, such as the amount of accelerator pedal depression (i.e., throttle opening), vehicle speed, engine rotational speed, etc. It is. Therefore, for example, if the second brake is engaged too early, a double lock state will occur in which both the front clutch and the second brake are temporarily engaged at the same time.
As shown in Fig. 5, the turbine rotational speed drops further than the rotational speed _N3 at 3rd gear at the beginning of the shift, and conversely, if the timing of engaging the second brake is too late, the front clutch will temporarily disengage. Both second brakes are released and become free, and as shown in Fig. 6, at the end of the shift, the turbine rotational speed is reduced to the rotational speed N ₂ at 2nd gear.
This causes the so-called dry blow phenomenon to occur. Such a drop or excessive rise in the turbine rotational speed during a shift appears as a shift shock. In addition, in an automatic transmission, when switching from D range to 2 range, the 3-2 timing valve is operated on the side where the release pressure is released quickly, that is, on the side where the friction member is fastened quickly. An example of a device that quickly releases the release pressure of the friction member to improve responsiveness during downshifting (that is, a device that allows the operating speed of the friction member to be adjusted) is disclosed in, for example, Japanese Patent Application Laid-open No. 193955/1983. There are some that are disclosed in the official gazette. On the other hand, when shifting from 2nd to 3rd gear,
Contrary to when downshifting, the clutch is engaged while releasing the brake, but since engaging the clutch during such a shift involves considerable shock, it is necessary to suppress this as much as possible. It is. Therefore, conventionally, when shifting up, the brake release operation and the clutch engagement operation are performed at the same time as described above, and the hydraulic line for releasing the brake servo mechanism and the hydraulic line for engaging the clutch are integrated. A so-called servo shelf (described later) is created by branching off from the main hydraulic line and using the volume increase effect of the servo mechanism to keep the working oil pressure almost constant (described later), and the clutch engagement is completed within the period of this servo shelf. That's what I do. That is, for example, when line pressure is supplied from the shift valve side to the release side of the clutch and brake servo mechanisms in conjunction with selection from 2nd gear to 3rd gear,
The operating pressure increases as shown by points 0 to A in FIG. 7 until it reaches a predetermined pressure P ₁ (that is, the pressure at which the servo mechanism starts operating); however, when the operating pressure reaches P ₁ , the servo mechanism As the piston begins to move, the operating pressure remains almost constant (points A to E), and when the piston reaches the end of its stroke (point E), the operating pressure increases again. This period from point A to point E is a period during which the pressure is maintained approximately constant due to the effect of increasing the volume of the hydraulic circuit as the piston of the servo engine moves, and this period is called a servo shelf. By engaging the clutch during such a servo shelf period, the shift shock can be reduced as much as possible. By the way, when the clutch is engaged during such a servo shelf period, in the low opening range where the opening of the throttle valve is small, the line pressure of the hydraulic circuit is set to the low pressure side by the vacuum throttle valve. Because of this, the flow rate of hydraulic oil is low and the movement speed of the servo piston is slow. Therefore,
The period of the servo shelf becomes longer, and the engagement of the clutch can be fully completed within this period of the servo shelf. However, in the high opening range where the throttle valve opening is large, the line pressure is set to a higher pressure side than in the low opening range, so the flow rate of hydraulic oil is large, and the servo piston moves faster. Therefore, the period of the servo shelf will be shortened accordingly. As a result, the engagement of the clutch is not completed within the period of the servo shelf, and the clutch is disengaged from the servo shelf and is carried out at a time when the pressure suddenly changes between points E and F, resulting in a large shift shock. The technical idea of incorporating an accumulator into a hydraulic servo circuit to control fluid pressure is conventionally known (see, for example, Japanese Utility Model Application No. 146101/1983). (Objective of the Invention) The present invention is intended to solve the problems pointed out in the above-mentioned prior art section. It is an object of the present invention to provide a control device for an automatic transmission that can prevent shift shocks caused by changes in the servo shelf period and shift shocks caused by fluctuations in the servo shelf period. (Means for Achieving the Object) As a means for achieving the above object, the present invention includes a speed change gear mechanism, a speed change switching means for changing the speed by switching the power transmission path of the speed change gear mechanism,
A fluid type actuator for operating this speed change switching means and a shift valve for controlling the supply of pressure fluid to this fluid type actuator are provided, and a friction member for selecting a low gear of the fluid type actuator is configured to disable the fluid type actuator. In addition to supplying release pressure,
In the automatic transmission, the high speed gear is selected by supplying a fastening pressure for actuating the high speed gear selection friction member, and the low speed gear is selected by releasing the release pressure and the fastening pressure. The solenoid valve for timing adjustment is energized on the side of the friction member for low speed selection, rather than the position where the hydraulic circuit that communicates the valve and the release side of the friction member for low speed selection is branched to the friction member for high speed selection. A control valve whose throttle amount is controlled depending on time is provided, and a throttle is provided closer to the shift valve than the branch position and acts to throttle the fluid flowing from the shift valve toward the branch position,
The present invention is characterized in that an accumulator is provided at a position downstream of the throttle with respect to the shift valve. (Function) In the present invention, by the above means, when downshifting from a high speed gear to a low speed gear, the hydraulic oil discharged from the release side of the low gear selection friction member is controlled by the energization time of the timing adjustment solenoid valve. The release timing of the friction member for selecting a low gear can be finely adjusted by appropriately restricting the throttle using a controlled control valve, so the friction member for selecting a high gear can be released regardless of the driving condition of the vehicle. The relative timing of the engagement of the friction member for selecting a low speed gear can be set to an optimal state in accordance with the increase in the rotational speed of the turbine of the torque converter. This ensures that gear shift shocks are prevented. On the other hand, when shifting from a low gear to a high gear, first of all, since the accumulator is provided, in addition to the volume increasing effect of the low gear selection friction member itself, the volume increasing effect of this accumulator can be obtained. A larger volume increase effect can be expected than in the case of only the volume increase effect of the friction member for low speed selection. For example, when the amount of supplied oil is constant, the operating speed of the accumulator and servo mechanism will be increased by the increased volume. will decrease. Further, since the throttle is disposed upstream of the accumulator, the hydraulic fluid supplied to the open side of the accumulator and the brake servo mechanism is appropriately throttled and reduced in pressure. Therefore, the flow rate of hydraulic oil supplied to the accumulator and the servo mechanism decreases, and both the operating speed of the accumulator and the servo piston decrease. As a synergistic effect of the volume increase effect due to the provision of the accumulator and the pressure reduction effect due to the provision of the throttle, the servo shelf period becomes even longer, and as a result, the servo shelf period becomes even longer, and as a result, Even in a region where the line pressure of the hydraulic circuit is set to a relatively high pressure, it is possible to complete the fastening of the high speed stage selection friction member within the servo shelf period. (Embodiments) Hereinafter, preferred embodiments of the present invention will be described with reference to FIGS. 1 to 4. FIG. 1 shows a power transmission section and its hydraulic control circuit of an automatic transmission according to an embodiment of the present invention. Configuration of automatic transmission The automatic transmission includes a torque converter 1, a multi-stage gear transmission 2, and an overdrive planetary gear transmission mechanism 3 disposed between the torque converter 1 and the multi-stage gear transmission 2. are doing. The torque converter 1 includes a pump 5 coupled to an engine output shaft 4, a turbine 6 disposed opposite the pump 5, and a stator 13 disposed between the pump 5 and the turbine 6. Furthermore, a converter output shaft 8 is coupled to the turbine 6 . In addition, this converter output shaft 8 and pump 5
A lock-up clutch 9 is provided between the two. This lock-up clutch 9 is always urged in the fastening direction by hydraulic oil pressure circulating within the torque converter 1, and is released and held by supplying release pressure oil from the outside into its pressure chamber 9a. Ru. The multi-stage gear transmission 2 has a front planetary gear mechanism 10 and a rear planetary gear mechanism 11, and a sun gear 12 of the front planetary gear mechanism 10 and a sun gear 13 of the rear planetary gear mechanism 11 are connected by a connecting shaft 14. . The input shaft 15 of the multi-gear transmission 2 connects to the connecting shaft 1 via a front clutch 16 (which together with a front clutch actuator 41 to be described later constitutes a friction member for high speed gear selection in the claims).
4 and to an internal gear 18 of a front planetary gear mechanism 10 via a rear clutch 17. A second brake 19 (which together with a second brake actuator 45 to be described later constitutes a friction member for selecting a low speed gear in the claims) is provided between the connecting shaft 14, that is, the sun gears 12 and 13, and the transmission casing. . The planetary carrier 20 of the front planetary gear mechanism 10 and the internal gear 21 of the rear planetary gear mechanism 11 are connected to an output shaft 22 . Further, a low reverse brake 24 and a one-way clutch 25 are provided between the planetary carrier 23 of the rear planetary gear mechanism 11 and the transmission case. This multi-gear transmission 2 has a conventionally known transmission mechanism, and has three forward speeds and one reverse speed, and has a front clutch 16 and a rear clutch 17.
By appropriately operating the second brake 19 and the low reverse brake 24 using a hydraulic actuator, as will be described later, a desired gear position can be obtained. The overdrive planetary gear transmission mechanism 3 includes a planetary carrier 27 that rotatably supports a planetary gear 26, and a sun gear 28 coupled to an internal gear 30 via a direct clutch 29. An overdrive brake 31 is provided between the sun gear 28 and the transmission case.
The internal gear 30 is connected to the input shaft 15 of the multi-gear transmission 2. This overdrive planetary gear transmission mechanism 3 connects the converter output shaft 8 and the input shaft 15 in a direct connection state when the direct clutch 29 is engaged and the overdrive brake 31 is released. When the converter output shaft 8 and the input shaft 15 are engaged and the direct clutch 29 is released, the converter output shaft 8 and the input shaft 15 are coupled in an overdrive manner. This transmission is operated by manually selecting and operating a manual valve 61 of a hydraulic control circuit, which will be described later.
The required gear stage is obtained by appropriately operating each friction member (clutch and brake) of the multi-stage gear transmission 2 and the overdrive planetary gear transmission mechanism 3, and the control pattern of each friction member is conventional. Similarly to the structure, each range is set as follows.

【table】

[Table] Hydraulic control circuit A: Configuration The hydraulic control circuit operates each friction member in the operation pattern shown in Table 1 above in response to the driver's selection operation so that a predetermined gear can be obtained. It has a circuit configuration, which will be explained in detail below. In the hydraulic control circuit, reference numeral 50 denotes a hydraulic pump, and the hydraulic oil discharged from the hydraulic pump 50 is regulated to a predetermined pressure by a pressure regulating valve 62.
Manual valve 6 via pressure line 101
1 will be introduced. The manual valve 61 is connected to the pressure line 101
It has five ports that can be communicated with. That is, port a communicates with the pressure line 101 in the three ranges (D, 2, 1), and (D,
2) Pressure line 101 in the two ranges
Port b communicates with pressure line 101 only in R range, port c communicates with pressure line 101 (P,
It has a port d that communicates with the pressure line 101 in the four ranges R, 2, 1), and a port e that communicates with the pressure line 101 in the two ranges (R, 1). Port a is connected to line 111. This line 111 has three pilot lines at its line ends: a first pilot line 102 with a 1-2 shift solenoid valve 51 controlling the operation of the 1-2 shift valve 63 and an orifice 86; A second pilot line 103 equipped with a 2-3 shift solenoid valve 52 and a throttle 87 that controls the operation of the shift valve 64 (corresponding to the shift valve in the claims), and a second pilot line 103 that controls the operation of the 3-4 shift valve 65. It is branched into a third pilot line 104 equipped with a 3-4 shift solenoid valve 53 and a throttle 88 to control. The solenoid valves 51, 52, and 53 are turned on (energized) to close the drain lines 105, 106, and 107, respectively, and build up pilot pressure in the pilot lines 102, 103, and 104, respectively. This moves each of the shift valves 63, 64, and 65 from the OFF position (rightward movement position) to the ON position (leftward movement position) to open and close the hydraulic circuit of the associated friction element. Operation patterns are established according to each range and gear as shown in Table 2 below.

[Table] In this case, the 3-4 shift solenoid valve 53 is in the OFF position during 1st and 2nd speeds in the D range, but when in 1st and 2nd speeds in the 1st and 2nd ranges ( ) As shown, each is set to the ON position. This is to set the 3-4 shift solenoid valve in the ON position in the 1st and 2nd ranges and apply pilot pressure to the backup control valve 70, which will be described later. The first pilot line 102 is the above-mentioned 1-2
On the downstream side of the shift solenoid valve 51, a first branch line 10 communicates with the right end portion (pilot pressure load portion) of the 1-2 shift valve 63.
2a and a second branch line 102b communicating with the right end portion (pilot pressure loading portion) of the cut valve 66.
It is branched into two lines. Line pressure is supplied to both ends of the 1-2 shift valve 63 via a line 112 branching from the line 111 and a line 113 further branching from the line 112.
Line pressure is introduced into the intermediate portion of each of the valves through a line 122 that communicates with port e of the manual valve 61. Note that this line 122 is connected to the 1-2 shift valve 63.
When it is in the OFF position, that is, when the gear shift position is in the first gear position, it is communicated with the line 123. Further, this line 123 is connected to a low reverse brake actuator 44. In contrast, line 113 is 1-
When the 2-shift valve 63 is in the ON position, that is, when the shift position is at a shift position other than 1st speed, it is communicated with the line 161. This line 161 is connected to the engagement side 45A of the second brake actuator 45. or,
This line 161 is provided with an accumulator 79 whose back pressure is controlled by a reducing valve 68 and a one-way orifice 82 . The 2-3 shift valve 64 is ON-OFF controlled by pilot pressure introduced from a line 103 connected to its right end. This 2-3 shift valve 64 has a line 121 communicating with port b of the manual valve 61 and a line 1 communicating with port c.
31 are connected to each other. Of these lines 121 and 131, the line 121 is connected to the line 132 when the 2-3 shift valve 64 is in the ON position (i.e., the 3rd or 4th gear shift position), and the line 131 is connected to the OFF position of the 2-3 shift valve 64. (i.e., the first speed or second speed shift position) are selectively connected to the line 132. Note that a reducing valve 67 and a one-way orifice 85 are each connected in parallel to the line 131. The line 132 is branched on its downstream side into a line 136 connected to the front clutch actuator 41 and a line 138 connected to the release side 45B of the second brake actuator 45. This line 13
A one-way orifice 74 is installed on the upstream side of the branch part 2. The one-way orifice 74 acts to throttle the hydraulic fluid flowing from the 2-3 shift valve 64 toward the branch part. In addition, a one-way orifice 75 (as defined in the claims) is located immediately downstream of the branch part of the line 138, which acts to throttle the hydraulic fluid from the second brake actuator 45 side to the 2-3 shift valve 64 side. (corresponding to the aperture) is installed. Further, a line 137 equipped with a 3-2 timing valve 73 is connected between an intermediate position between the branch part of the line 132 and the one-way orifice 74 and a position downstream of the one-way orifice 75 of the line 138. . This 3-2 timing valve 73 has a spool that is moved forward and backward by the pilot pressure applied to its left end portion, thereby making the passage area of the line 137 variable. It is connected to the line 112 mentioned above. This line 163 and line 1
12 is provided with a throttle 72, and a drain line 117 provided on the line 163 side from the throttle 72 is provided with a timing adjustment solenoid valve 47, and when the solenoid valve 47 is turned on, line 163
A predetermined pilot pressure is generated within the cylinder. Also, depending on the ON operation time of this timing adjustment solenoid valve 47, the above 3-2
The amount of spool displacement of the timing valve 73, that is,
The aperture amount is adjusted. The timing adjustment solenoid valve 47 is controlled by a control signal _A5 output from a controller (see FIG. 2), as will be described later. Further, the one-way orifice 75 provided in the line 138 is designed to throttle the hydraulic fluid discharged from the release side 45B of the second brake actuator 45, but the amount of throttle of this one-way orifice 75 is fixed. . Also, the above line 137,1 of line 138
A one-way orifice 83 is provided further downstream from the branch part 38. Further, in the line 136, there is an accumulator 78 whose back pressure is controlled by line pressure supplied via a line 139 branching from the line 132.
They are connected via a one-way orifice 81 that acts to throttle the hydraulic fluid flowing out.
The back pressure of the accumulator 78 is controlled by the reducing valve 67. The third pilot line 104 is the above-mentioned 3-4
On the downstream side of the shift solenoid valve 53, there are two branch lines: a first branch line 104a that communicates with the right end of the 3-4 shift valve 65, and a second branch line 104b that leads pilot pressure to a backup control valve 70, which will be described later. It is branched into. A line 141 branching from the pressure line 101 is connected to an intermediate portion of the 3-4 shift valve 65. This line 141 is 3-4
It is communicated with the line 142 at the OFF position of the shift register 65 (ie, at a shift position other than 4th speed). This line 142 is branched into two lines on its downstream side: a line 143 that communicates with the direct clutch actuator 42 and a line 144 that communicates with the release side 46B of the overdrive brake 46.
Further, a hydraulic switch 90 is installed at a position upstream of the branch. Further, an accumulator 77 is provided in the line 143.
In addition, the engagement side 46 of the overdrive brake 46
A is connected to the pressure line 1 via line 148.
01. On the other hand, a line 151 communicating with the manual valve 61 and port d is connected to the left end of the 3-4 shift valve 65.
The spool of the 4-shift valve 65 is forced by the line pressure introduced through the line 151 at select positions other than the D range.
Positioned and held in the OFF position. or,
A line 152 branches from this line 151 and communicates with the vacuum throttle valve 69.
The throttle backup valve 71 and the backup control valve 70 are connected in series, and the backup control valve 70
is installed so as to be located downstream of the throttle backup valve 71.
This throttle back up valve 71 has two
The vacuum throttle valve 69 controls the line pressure in the line 152 in the range and 1 range.
and act on the vacuum throttle valve 6.
9, the pressure regulating valve 62 is driven in the pressure increasing direction to increase the line pressure. Also, backup control valve 70
is located between the throttle back up valve 71 and the vacuum throttle valve 69, and when pilot pressure is built up in the branch line 104b, that is, when the 3-4 shift solenoid valve 53 is turned on, the line 152 is closed. It operates to open the throttle back-up valve 71 and enable the line pressure increasing action by the throttle backup valve 71. Further, the line 112 is connected to a line 116 communicating with the rear clutch actuator 43 on the upstream side of the throttle 72. This line 116 is provided with an accumulator 80 and a one-way orifice 84, respectively. Further, a line 114 branching from this line 112 has a regulating valve 6 for adjusting the back pressure of the accumulator 79.
8 is provided. Furthermore, the line 149 is a line 1 equipped with a lock-up valve 76 and an aperture 91 in the middle thereof.
46, it is communicated with the working chamber 9a of the lock-up platform 9. The pilot line 145 of this lock-up valve 76 includes a throttle 89 and a lock-up solenoid valve 54.
The lock-up clutch 9 is provided with a lock-up solenoid valve 54.
When activated, pilot pressure is built up in line 145, which is released when line 146 and line 149 communicate with each other due to the spool. In this embodiment, the lock-up clutch 9 is engaged only between the first and third speeds of the D range. B: Operation and Effect The automatic transmission Z equipped with the power transmission part and the hydraulic control circuit as described above has three shift solenoid valves 51 of the hydraulic control circuit as shown in FIG.
52, 53 and lock-up solenoid valve 5
4 and timing adjustment solenoid valve 47
, control signals A ₁ , A ₂ , which are output from the controller based on various engine state detection elements such as a throttle opening signal and an engine speed signal.
By performing ON-OFF control using A ₃ , A ₄ , and A ₅ according to the control patterns shown in Table 2, each friction member operates according to the operation pattern shown in Table 1, and a predetermined gear is achieved. It is something that can be done. Furthermore, in this embodiment, the timing adjustment solenoid valve 47 is controlled by the control signal output from the controller by applying the present invention.
_A5 adjusts the engagement timing of the second brake 19 when shifting from 3rd gear to 2nd gear in the D range, thereby preventing a shift shock caused by inappropriate engagement timing. The operation of the hydraulic control circuit and its effects will be explained below.
A description will be given of the time of downshifting from the third speed driving state to the second speed driving state and the time of shifting up from the second speed driving state to the third speed driving state as examples. a: Downshifting from 3rd speed to 2nd speed a-1: When driving in 3rd speed In this case, port a and port b of the manual valve 61 are connected to the pressure line 101, and the three shift Among the solenoid valves 51, 52, and 53, the 1-2 shift solenoid valve 51 and the 2-3 shift solenoid valve 52 are both in the ON position, and the 3-2 shift solenoid valve 51 is in the ON position.
4-shift solenoid valve 53 is in the OFF position. For this reason, the 1-2 shift valve 6
3, the line 113 and the line 161 communicate with each other, and line pressure is applied to the fastening side 45A of the second brake actuator 45.
In addition, in the 2-3 shift valve 64, the line 121 and the line 132 communicate with each other, and the release side 45 of the front clutch actuator 41 and the second brake actuator 45 is connected to each other.
B is loaded with line pressure, respectively. Therefore, the front clutch 16 is engaged and the second brake 19 is released. On the other hand, since the line 123 is drained, the low reverse brake 24 is released. That is, the engine brake is in a non-acting state. Moreover, in the 3-4 shift valve 65,
Line 141 and line 142 are in communication. Therefore, the direct clutch 29 is engaged. On the other hand, overdrive brake 46
, line pressure is introduced to the release side 46B via the pressure line 142,
Therefore, overdrive brake 31 is in a released state. Furthermore, line pressure is introduced into the line 112, so that the rear clutch 17 is in the engaged state. By operating each friction member in this manner, third speed running in the D range is realized. On the other hand, in this state, line 151
is being drained, the 3-4 shift solenoid valve 53 is turned on and pilot pressure is generated in the branch line 104b, and even though the vacuum control valve 70 is set to the ON position, the vacuum throttle valve 69 does not have throttle pressure. does not work. Therefore, the line pressure is maintained on the low pressure side. a-2: Downshifting to 2nd speed When downshifting to 2nd speed, the 1-2 shift solenoid valve 51 remains in the ON position.
3-4 shift solenoid valve 53 is OFF
Each position is maintained, but the 2-3 shift solenoid valve 52 is shifted from the ON position.
Switched to OFF position. Therefore, the line 121 and the line 161 communicate with each other in the 1-2 shift valve 63 portion, the line 123 is drained, and the low reverse brake 24 is released. Also, 2-3
Line 13 in the shift valve 64 part
2 is drained. Therefore, the engagement pressure of the front clutch 16 and the second brake 1
9 release pressure is released, and the front clutch 16
is released and the second brake 19 is engaged. Further, the rear clutch 17 is also tightened from the point where this release pressure is released. On the other hand, engagement pressure is applied to both the direct clutch 29 and the rear clutch 17, so that the direct clutch 29 and the rear clutch 17 are brought into a engaged state. By operating each friction member in this manner, two-speed running is realized. By the way, as mentioned above, in the D range, 3
When a downshift operation is performed from speed to second speed, the front clutch 16 is released from the engaged state, and in response, the second brake 19 is
In this case, if the release timing of the front clutch 16 and the engagement timing of the second brake 19 are inappropriate, the turbine rotational speed may drop at the beginning of the shift or may become excessive at the end of the shift. It has already been mentioned that this may cause a large shift shock (see the section on prior art, Figures 5 and 6).
However, in this embodiment, a 3-2 timing valve 73 is provided in the line 137 by applying the present invention, and this 3-2 timing valve 73 is controlled by the timing adjustment solenoid valve 47. Efforts are being made to prevent such problems from occurring. In other words, when downshifting from 3rd gear to 2nd gear, the turbine rotational speed increases from 3rd gear rotational speed N ₃ to 2nd gear rotational speed N ₂ as shown in Figure 4, but the rate of increase (in other words, For example, the time during which the rotational speed fluctuates) changes in a complicated manner depending on various vehicle conditions (e.g., stroke opening, vehicle speed, engine rotational speed, etc.).
Therefore, if the time lag between the release of the front clutch 16 and the engagement of the second brake 19 is set to a constant value, it is impossible to set the time lag between the release of the front clutch 16 and the engagement of the second brake 19 to be the same as when the turbine speed has finished rising and when the second brake 19 has been engaged, as shown in FIG. It is impossible to match. Therefore, in this embodiment, the hydraulic fluid discharged from the release side 45B of the second brake actuator 45 is used as a throttle mechanism that acts to delay the engagement timing of the second brake 19 from the release timing of the front clutch 16. One-way orifice 75 with fixed aperture amount
3. The amount of throttle can be changed and adjusted by changing the energization time of the timing adjustment solenoid valve 47.
-2 timing valve 73 is provided, the basic timing deviation amount is set by a one-way orifice 75, and a control valve 7 is provided.
3 sets the amount of timing deviation according to the vehicle condition, and by the cooperation of these two, it is possible to control the timing when the turbine rotation speed reaches the rotation speed N ₂ in 2nd gear and the timing when the second brake 19 is engaged. 4, thereby obtaining an ideal fluctuation characteristic of the number of fluctuations in the turbine rotational speed as shown in FIG. In order to realize such a function, in this embodiment, in advance, the rotation speed of the turbine is changed from N ₃ to the rotation speed of the turbine when downshifting from 3rd gear to 2nd gear under various vehicle conditions.
The time required for the change to _N2 is determined and stored in the controller's map as a characteristic diagram of the change in turbine rotation speed for each control element. In actual control, as shown in the flowchart in Figure 3, first, data on various actual control elements (e.g., throttle opening, vehicle speed, engine speed, etc.) when downshifting from 3rd gear to 2nd gear is collected. input (step _S1 ).
Next, it is determined whether a downshift operation from 3rd speed to 2nd speed has been performed based on the operating state of each shift solenoid valve 51, 52, 53 (step _S2 ). If the result of the determination is that no downshift operation has been performed, no control operation is performed. On the other hand, when a downshift operation is performed, each control element input this time is checked and compared with the above-mentioned operating characteristic diagram to determine the optimum energization time (fast, 3-2 The throttle amount of the timing valve 78 is set (step _S3 ), and its control signal is output (step _S4 ). In this way, by controlling the energization time of the timing adjustment solenoid valve 47,
The ideal turbine rotation speed characteristics as shown in FIG. 4 can be obtained. Therefore, a downshift operation with less gear shift shock can be realized. b: Shifting from 2nd speed to 3rd speed Shifting from 2nd speed to 3rd speed is as follows:
This is done by switching the 2-3 shift solenoid valve 52 from OFF to ON, keeping the 1-2 shift solenoid valve 51 in the ON position, and keeping the 3-4 shift solenoid valve 53 in the OFF position. Therefore, in this case, the 1-2 shift valve 63 and the 4-3 shift valve 65 are the same as when running in the 2nd speed, but the 2-3 shift valve 64 has a line 121 and a line 132.
The line pressure is introduced into the release side 45B of the front clutch actuator 41 and the second brake actuator 45, respectively. Therefore, the front clutch 16 is engaged from the open state, and the second brake 19 is also engaged.
is released from the fastened state. Further, the rear clutch 17, lock-up clutch 9, and direct clutch 29 remain engaged, and the low reverse brake 24 and overdrive brake 31 remain open. By the way, at this time, if the throttle opening is large, that is, in a high-load operating state, the line pressure is set high due to the action of the vacuum throttle valve 69, so the movement of the servo piston of the second brake actuator 45 is reduced. As already mentioned, there is a risk that the front clutch 16 will not be fully engaged within the shelf pressure period and a large shift shock will occur as the speed increases (see the section on the prior art). However, in this embodiment, the present invention is applied to provide the accurator 48 in the line 136 and the one-way orifice 74 in the line 132, so that the occurrence of such a shift shock can be prevented as much as possible. . That is, first, due to the throttle effect of the one-way orifice 74, the release side 45B of the second brake actuator 45, the front clutch actuator 41, and the above-mentioned accumulator 78 are
The hydraulic pressure introduced into the line is lower than the line pressure. As a result, the moving speeds of the servo piston of the second brake actuator 45 and the piston of the accumulator 78 are suppressed (achieving a pressure reducing effect). Furthermore, line 13 communicating with line 138
6 is provided with an accumulator 78,
The volume of the second brake actuator 45 on the release side 45B side is substantially increased by the volume of the accumulator 78. Therefore, the moving speed of the servo piston of the second brake actuator 45 and the piston of the accumulator 78 is suppressed (realization of the volume increase effect). As a result of the volume increase effect of In the figure, the curve OA
-E-B-C-D, conventionally (i.e., when one-way orifice 74 and accumulator 79 are not provided), the range a from point A to point E is
Although the servo shelf only occurs during this period, the servo shelf period is connected to range b from point A to point B (range b>range a). Therefore, even during high load operation of the engine where the line pressure is high and the moving speed of the servo piston tends to be accelerated, the engagement of the front clutch actuator 41 can be reliably completed within the period of the servo shelf. As a result, the front clutch actuator 41 is engaged more smoothly and softly compared to the case where the front clutch actuator 41 deviates beyond the servo shelf into the pressure sudden change area, and the speed change is made that much faster. This is to prevent shocks as much as possible. (Effects of the Invention) The control device for an automatic transmission of the present invention includes a speed change gear mechanism, a speed change means for changing the speed by switching the power transmission path of the speed change gear mechanism, and a fluid actuator for operating the speed change change means. and a shift valve for controlling the supply of pressure fluid to the hydraulic actuator, which supplies a release pressure to the friction member for selecting a low gear of the hydraulic actuator to deactivate it, and also for selecting a high gear. In an automatic transmission, a high speed gear is selected by supplying a fastening pressure that operates the friction member, and a low speed gear is selected by releasing the release pressure and the fastening pressure. The hydraulic circuit that communicates with the release side of the friction member for gear selection is arranged closer to the friction member for low gear selection than the position where it branches to the friction member for high gear selection, depending on the energization time of the timing adjustment solenoid valve. A control valve for controlling the amount is provided, and a throttle is provided closer to the shift valve than the branch position and acts to throttle the fluid flowing from the shift valve toward the branch position, and the throttle is connected to the shift valve. The feature is that an accumulator is provided at a downstream position. Therefore, according to the automatic transmission control device of the present invention, when downshifting from a high gear to a low gear,
By controlling the hydraulic oil discharged from the release side of the low speed selection friction member with a control valve whose throttle amount is controlled by the energization time of the timing adjustment solenoid valve, the low speed selection friction member It is possible to fine-tune the release timing of the torque converter according to the increase in the turbine rotational speed and set it to the optimum timing, thereby preventing shift shocks caused by inappropriate engagement timing of the friction member for low gear selection. This has the effect of preventing Furthermore, when shifting from a low gear to a high gear, the synergistic effect of the pressure reduction effect of the throttle on the hydraulic pressure introduced to the release side of the low gear selection friction member and the volume increase effect of the hydraulic circuit by the accumulator causes the low speed to change. Since the moving speed of the servo piston of the stage selection friction member is reduced, the shelf period with almost no oil pressure change is made as long as possible, and the line pressure of the hydraulic circuit is set to a relatively high pressure, such as in the throttle high opening area. It is possible to complete the fastening of the friction member for high-speed gear selection within the shelf period even in the range where the throttle is opened, and it is possible to more effectively prevent shift shock during upshifting, especially in the high throttle opening range. is obtained.

[Brief explanation of drawings]

Fig. 1 is a system diagram of an automatic transmission equipped with a control device according to an embodiment of the present invention, Fig. 2 is a control system diagram of the automatic transmission shown in Fig. 1, and Fig. 3 is a control flowchart. 4 to 6 are state changes in the turbine rotational speed during gear shifting, and FIG. 7 is a characteristic diagram of the working oil pressure. DESCRIPTION OF SYMBOLS 1... Torque converter, 2... Multi-stage gear transmission, 3... Planetary gear transmission mechanism, 10... Front stage planetary gear mechanism, 11... Rear stage planetary gear mechanism, 16...
Front clutch, 17...Rear clutch, 19
...Second brake, 24...Low reverse brake, 29...Direct clutch, 31...
Overdrive brake, 41... actuator for front clutch, 42... actuator for direct clutch, 43... actuator for rear clutch, 44... actuator for low reverse brake, 45... actuator for second brake, 46... overdrive brake , 51...1-2 shift solenoid valve, 5
2...2-3 shift solenoid valve, 53...
3-4 shift solenoid valve, 54... lock-up solenoid valve, 61... manual valve, 62... pressure regulating valve, 63...1-2 shift valve, 64...2-3 shift valve, 65...3
-4 shift valve, 69...vacuum throttle valve, 70...backup control valve, 71...throttle backup valve,
73...3-2 timing valve, 74...One-way orifice (throttle), 78...Accumulator.

Claims (1)

[Claims] 1 a speed change gear mechanism, a speed change switching means for changing the speed by switching the power transmission path of the speed change gear mechanism;
A fluid type actuator for operating this speed change switching means and a shift valve for controlling the supply of pressure fluid to this fluid type actuator are provided, and a friction member for selecting a low gear of the fluid type actuator is configured to disable the fluid type actuator. In addition to supplying release pressure,
In the automatic transmission, the high speed gear is selected by supplying a fastening pressure for actuating the high speed gear selection friction member, and the low speed gear is selected by releasing the release pressure and the fastening pressure. A hydraulic circuit that communicates between the valve and the release side of the low speed selection friction member is located closer to the low speed selection friction member than the position where it branches to the high speed selection friction member side, during the energization time of the timing adjustment solenoid valve. Therefore, a control valve is provided whose throttle amount is controlled, and a throttle is provided closer to the shift valve than the branch position and acts to throttle the fluid flowing from the shift valve toward the branch position. A control device for an automatic transmission, characterized in that an accumulator is provided at a position downstream of the throttle with respect to the valve.