JP5211945B2 - Friction fastening element control device - Google Patents

Description

  The present invention relates to a control device for a frictional engagement element that generates a fastening force by pressing a friction plate with a piston that is operated by hydraulic pressure.

A technique described in Patent Document 1 is known regarding a control device for a frictional engagement element. In this technique, a command in which a rectangular wave is superimposed on a target clutch pressure that presses the piston is output, and the stroke position of the piston is learned based on a change in the amplitude of the actual clutch pressure at this time.
JP 2006-336806 A

  Here, with respect to a system that employs a hydraulic circuit configuration including a pressure reducing valve or the like on the upstream side (high pressure side) of the engagement pressure control valve that controls the clutch pressure, the stroke position is applied by applying the technique of Patent Document 1 above. Even if this determination was executed, the problem was found that the amplitude change of the actual clutch pressure could not be detected well and proper learning could not be performed.

  The present invention has been made paying attention to the above-mentioned problem, and a friction engagement element control device capable of learning an appropriate stroke position even in a system including a pressure reducing valve or the like upstream of an engagement pressure control valve. The purpose is to provide.

To achieve the above object, the invention according to claim 1, when learning the stroke position of the frictional engagement elements, and outputs a command signal obtained by superimposing the auxiliary signal and the basic fastening pressure signal to engagement pressure control valve Therefore, a command signal having a higher pressure than normal is output to the pressure reducing valve.

  Therefore, even when a pressure reducing valve is provided on the upstream side, the change in the amplitude of the fastening pressure can be obtained, and appropriate stroke position learning can be performed.

  DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The best mode for realizing a friction engagement element control device according to the present invention will be described below based on an embodiment shown in the drawings.

  FIG. 1 is a system diagram showing a power train of a vehicle equipped with a twin clutch type automatic manual transmission AMT. The engine ENG, which is an internal combustion engine, includes an injector IJ that controls the fuel injection amount, a throttle valve actuator TVA that controls the throttle opening, and an accelerator opening sensor signal as actuators that control engine output torque, engine speed, and the like. And an engine controller CU that outputs a control signal to the injector IJ and the throttle valve actuator TVA based on the engine speed sensor signal and the like. In addition, although Example 1 shows an example in which the engine ENG is mounted, a configuration including other drive sources such as an electric motor may be used.

  A twin-clutch automatic manual transmission AMT is connected to the output side of the engine ENG. The twin-clutch automatic manual transmission AMT is equipped with multiple clutches, shift actuators, transmission gear trains, and so on. In addition, a clutch hydraulic module 46 that controls the first clutch CA and the second clutch CB, an actuator hydraulic module 59 that controls each shift actuator, and an automatic MT that outputs a control signal to each module based on various sensor signals. And a controller 47. Details will be described later.

  Next, the configuration of the automatic manual transmission AMT will be described. FIG. 2 is an overall system diagram showing a twin-clutch automatic manual transmission AMT to which the frictional engagement element control device according to the first embodiment is applied. As shown in FIG. 2, the transmission gear train of the twin-clutch automatic manual transmission AMT according to the first embodiment includes a transmission case 1, a drive input shaft 2, a first clutch CA, a second clutch CB, and a toe. The motor includes a national damper 3, an oil pump 4, a first transmission input shaft 5, and a second transmission input shaft 6.

  The first clutch CA is for odd gears (first speed, third speed, fifth speed, reverse), and the second clutch CB is an even gear speed (second speed, fourth speed, sixth speed). It is for. The drive sides of both clutches CA and CB are connected via a torsional damper 3 to a drive input shaft 2 for inputting drive force from the engine ENG.

  The driven side of the first clutch CA inputs the driving force from the driving source to the first transmission input shaft 5 when the odd-numbered gear stage is selected. The driven side of the second clutch CB inputs the driving force from the driving source to the second transmission input shaft 6 when the even-numbered gear stage is selected.

  The oil pump 4 is always operated by a drive source, and the oil discharged from the oil pump 4 is used as a hydraulic pressure source to execute engagement / release control of both clutches CA and CB, which will be described later, and gear position selection control by a shift actuator. To do.

  The second transmission input shaft 6 is a hollow shaft, the first transmission input shaft 5 is a solid shaft, and is connected to the first transmission input shaft 5 via a front side needle bearing 7 and a rear side needle bearing 8. The second transmission input shaft 6 is rotatably supported in a concentric state.

  The second transmission input shaft 6 is rotatably supported by a ball bearing 9 with respect to the front end wall 1 a of the transmission case 1. The first transmission input shaft 5 protrudes from the rear end of the second transmission input shaft 6, and the protruding rear end portion 5 a of the first transmission input shaft 5 passes through the intermediate wall 1 b of the transmission case 1. At the same time, it is rotatably supported by the ball bearing 10 with respect to the intermediate wall 1b.

  A transmission output shaft 11 is provided on the concentric shaft at the rear end portion 5 a of the first transmission input shaft 5. The transmission output shaft 11 is rotatably supported on the rear end wall 1c of the transmission case 1 by a taper roller bearing 12 and an axial bearing 13, and is connected to the rear of the first transmission input shaft 5 via a needle bearing 14. The end 5a is rotatably supported.

  A counter shaft 15 is provided in parallel with the first transmission input shaft 5, the second transmission input shaft 6 and the transmission output shaft 11. The countershaft 15 is rotatably supported on the front end wall 1a, the intermediate wall 1b, and the rear end wall 1c of the transmission case 1 via roller bearings 16, 17, and 18.

  A counter gear 19 is integrally provided at the rear end of the counter shaft 15. An output gear 20 is provided on the transmission output shaft 11, and the counter gear 19 and the output gear 20 are engaged with each other to drive-couple the counter shaft 15 to the transmission output shaft 11. The counter gear 19 and the output gear 20 constitute a reduction gear set.

  Between the rear end portion 5a of the first transmission input shaft 5 and the countershaft 15, there is a gear set of an odd-numbered gear stage group (first speed, third speed, reverse), that is, first in order from the front side. A speed gear set G1, a reverse gear set GR, and a third speed gear set G3 are arranged.

  The first speed gear set G1 meshes the first speed input gear 21 provided at the rear end portion 5a of the first transmission input shaft 5 and the first speed output gear 22 provided on the counter shaft 15 with each other. Configured.

  The reverse gear set GR includes a reverse input gear 23 provided at the rear end portion 5a of the first transmission input shaft 5, a reverse output gear 24 provided on the countershaft 15, and a reverse idler gear meshing with both gears 23 and 24. 25. The reverse idler gear 25 is rotatably supported with respect to a reverse idler shaft 25a that protrudes from the intermediate wall 1b of the transmission case 1.

  The third speed gear set G3 meshes a third speed input gear 26 provided at the rear end portion 5a of the first transmission input shaft 5 and a third speed output gear 27 provided on the counter shaft 15 with each other. Configured.

  A 1-R synchronous mesh mechanism 28 is provided on the countershaft 15 between the first speed gear set G1 and the reverse gear set GR. The first speed output gear 22 is driven to the countershaft 15 by causing the coupling sleeve 28a of the 1-R synchronous meshing mechanism 28 to stroke to the left from the neutral position shown in the figure and to be splined to the clutch gear 28b. The first speed can be selected by combining. Further, the coupling sleeve 28a of the 1-R synchronous meshing mechanism 28 is stroked to the right from the neutral position shown in the drawing, and the clutch gear 28c is spline-fitted to drive-couple the reverse output gear 24 to the countershaft 15. The reverse speed can be selected.

  On the rear end portion 5a of the first transmission input shaft 5 between the third speed gear set G3 and the output gear 20, a 3-5 synchronous meshing mechanism 29 is provided. Then, the third-speed input gear 26 is input to the first transmission by causing the coupling sleeve 29a of the 3-5 synchronous mesh mechanism 29 to stroke leftward from the illustrated neutral position and to be splined to the clutch gear 29b. Drive-coupled to the shaft 5 allows the third speed to be selected. Further, the coupling sleeve 29a of the 3-5 synchronous meshing mechanism 29 is stroked to the right from the neutral position shown in the figure and is splined to the clutch gear 29c, whereby the first transmission input shaft 5 and the output gear 20 are Is directly connected and the fifth speed can be selected.

  Between the second transmission input shaft 6 and the countershaft 15, there is a gear group of even-numbered speed group (second speed, fourth speed, sixth speed), that is, the sixth speed gear group in order from the front side. G6, the second speed gear set G2, and the fourth speed gear set G4 are arranged.

  The sixth speed gear set G6 is configured by meshing a sixth speed input gear 30 provided on the second transmission input shaft 6 and a sixth speed output gear 31 provided on the countershaft 15.

  The second speed gear set G2 is configured by meshing a second speed input gear 32 provided on the second transmission input shaft 6 and a second speed output gear 33 provided on the countershaft 15.

  The fourth speed gear set G4 is configured by meshing a fourth speed input gear 34 provided on the second transmission input shaft 6 and a fourth speed output gear 35 provided on the countershaft 15.

  A 6-N synchronous meshing mechanism 37 is provided on the counter shaft 15 on the side of the sixth speed gear set G6. Then, the sixth speed output gear 31 is driven to the countershaft 15 by causing the coupling sleeve 37a of the 6-N synchronous meshing mechanism 37 to stroke leftward from the neutral position shown in the drawing and to be splined to the clutch gear 37b. Combined, the 6th speed can be selected.

  A 2-4 synchronous meshing mechanism 38 is provided on the countershaft 15 between the second speed gear set G2 and the fourth speed gear set G4. Then, the second-speed output gear 33 is driven to the countershaft 15 by causing the coupling sleeve 38a of the 2-4 synchronous meshing mechanism 38 to stroke leftward from the neutral position shown in the figure and by spline fitting with the clutch gear 38b. Combined, the second speed can be selected. Further, the fourth-speed output gear 35 is driven to the countershaft 15 by causing the coupling sleeve 38a of the 2-4 synchronous meshing mechanism 38 to stroke rightward from the neutral position shown in the drawing and to be spline-fitted to the clutch gear 38c. Combined to enable selection of 4th speed.

  Next, a clutch engagement and gear selection control system of the twin clutch type automatic manual transmission AMT according to the first embodiment will be described. FIG. 2 is a system diagram showing a control system of the twin clutch type automatic manual transmission AMT according to the first embodiment. The twin-clutch automatic manual transmission AMT includes a 3-5 shift fork 41, a 1-R shift fork 42, a 6-N shift fork 43, a 2-4 shift fork 44, an actuator unit 45, and a clutch hydraulic module 46. And an automatic MT controller 47.

  The 3-5 shift fork 41 is engaged with the coupling sleeve 29 a of the 3-5 synchronous meshing mechanism 29 and is fixed to the first shift rod 48. The first shift rod 48 is supported so as to be movable in the axial direction with respect to the front end wall 1 a and the intermediate wall 1 b of the transmission case 1. Then, the 3-5 shift bracket 49 is fixed to the first shift rod 48, and the end portion of the 3-5 shift bracket 49 is loosely supported by the spool connecting shaft portion of the 3-5 shift actuator 50. That is, the 3-5 shift fork 41 strokes leftward (when the third speed is selected) or rightward (when the fifth speed is selected) from the illustrated neutral position according to the spool operation of the 3-5 shift actuator 50.

  The 1-R shift fork 42 is engaged with the coupling sleeve 28 a of the 1-R synchronous meshing mechanism 28 and is provided so as to be able to stroke in the axial direction with respect to the second shift rod 51. The second shift rod 51 is provided in a fixed state in the axial direction with respect to the front end wall 1 a and the intermediate wall 1 b of the transmission case 1. The end portion of the bracket arm portion 42b that is integrally formed with the bracket cylindrical portion 42a of the 1-R shift fork 42 is idled and supported by the spool connecting shaft portion of the 1-R shift actuator 52. That is, the 1-R shift fork 42 strokes from the neutral position shown in the figure to the left (when the first speed is selected) or right (when the reverse speed is selected) according to the spool operation of the 1-R shift actuator 52.

  The 6-N shift fork 43 engages with the coupling sleeve 37a of the 6-N synchronous meshing mechanism 37 and is provided so as to be capable of stroke in the axial direction with respect to the second shift rod 51 fixed in the axial direction with respect to the transmission case 1. ing. Then, the end portion of the bracket arm portion 43 b formed integrally with the bracket cylindrical portion 43 a of the 6-N shift fork 43 is idled and supported by the spool connecting shaft portion of the 6-N shift actuator 53. That is, the 6-N shift fork 43 strokes from the neutral position shown in the drawing to the left (when the sixth speed is selected) according to the spool operation of the 6-N shift actuator 53.

  The 2-4 shift fork 44 is engaged with the coupling sleeve 38a of the 2-4 synchronous meshing mechanism 38, and is provided so as to be capable of stroke in the axial direction with respect to the second shift rod 51 fixed in the axial direction with respect to the transmission case 1. ing. The end portion of the bracket arm portion 44b formed integrally with the bracket cylindrical portion 44a of the 2-4 shift fork 44 is loosely supported by the spool connecting shaft portion of the 2-4 shift actuator 54. That is, the 2-4 shift fork 44 strokes from the neutral position shown in the figure to the left (when the second speed is selected) or right (when the fourth speed is selected) according to the spool operation of the 2-4 shift actuator 54.

  The actuator unit 45 is fixed to a lower position, an upper position, a side position, or the like of the transmission case 1, and includes a 3-5 shift actuator 50, a 1-R shift actuator 52, a 6-N shift actuator 53, 2- The 4-shift actuator 54, the 3-5 shift position sensor 55, the 1-R shift position sensor 56, the 6-N shift position sensor 57, the 2-4 shift position sensor 58, and the actuator hydraulic module 59 are integrated. It is a unit to have.

  The actuator hydraulic module 59 generates an even speed step pressure Pe and an odd speed step pressure Po using the line pressure PL adjusted by the clutch hydraulic module 46 as a source pressure, and further shifts each shift according to the selected speed step. Actuator operating pressure is supplied to the transmission pressure oil path to the actuators 50, 52, 53, 54.

  The clutch hydraulic module 46 regulates the line pressure PL based on the oil discharged from the oil pump 4, and creates a clutch engagement pressure for the first clutch CA based on the even speed step pressure Pe from the actuator hydraulic module 59. A clutch engagement pressure for the second clutch CB is generated based on the odd speed step pressure Po.

  The automatic MT controller 47 inputs information from a vehicle speed sensor, an accelerator opening sensor, a range position sensor, and other sensors / switches, and outputs a shift speed selection control command to each solenoid of the actuator hydraulic module 59. A clutch engagement control command (including a line pressure control command) is output to each solenoid of the clutch hydraulic module 46.

  FIG. 3 is a hydraulic circuit diagram showing when the sequence solenoid is off in the actuator hydraulic module 59 and the clutch hydraulic module 46 of the first embodiment. FIG. 4 shows when the sequence solenoid is on in the actuator hydraulic module 59 and the clutch hydraulic module 46 of the first embodiment. It is a hydraulic circuit diagram.

  As shown in FIG. 2, the automatic manual transmission AMT according to the first embodiment includes a first clutch CA that is engaged when an odd-numbered gear group (first speed, third speed, fifth speed, reverse) is selected, and an even speed change. A twin-clutch automatic manual transmission including a second clutch CB that is engaged when a stage group (second speed, fourth speed, and sixth speed) is selected.

  The first embodiment includes the 1-R synchronous meshing mechanism 28, the 1-R shift fork 42, and the 1-R shift actuator 52 that are common to the first speed stage and the reverse stage, and the 1-R shift actuator 52 when starting forward. The first speed output gear 22 (first speed gear) on the countershaft 15 is drive-coupled by the selection operation according to, and the reverse output gear 24 on the countershaft 15 (by the selection operation by the 1-R shift actuator 52 at the time of reverse start. A start control means for drivingly connecting a reverse gear) is provided. The start control means includes an actuator hydraulic module 59 that invalidates the selection of the first speed or the reverse speed immediately after the selection of the first speed or the reverse speed.

  The actuator hydraulic module 59 includes eight oil passages 61, 62, 63, 64, 65, 66, 67, 68 for the four shift actuators 50, 52, 53, 54, four actuator solenoids 71, 72, The actuator hydraulic circuit is opened and closed by 73 and 74 and one sequence solenoid 75.

  The eight oil passages are the third speed pressure oil path 61, the fifth speed pressure oil path 62, the first speed pressure oil path 63, the reverse pressure oil path 64, the second speed pressure oil path 65, and the fourth speed pressure. An oil passage 66, a sixth speed oil passage 67, and a neutral pressure oil passage 68 are configured.

  The four actuator solenoids include a first actuator solenoid 71 and a second actuator solenoid 72 that generate the hydraulic pressure of the even gear group, a third actuator solenoid 73 that generates the hydraulic pressure of the odd gear group, and a fourth actuator solenoid. 74.

  As shown in FIG. 3, the one sequence solenoid 75 is a low-speed gear stage (first speed stage, second speed stage, fourth speed stage, reverse stage) including a first speed stage and a reverse speed stage on the solenoid-off side. ) Can be selected, and as shown in FIG. 4, high speed gears (third speed, fifth speed, sixth speed) can be selected on the solenoid-on side, and the first speed and the reverse speed are selected. A spool 76 that invalidates the selection.

  The actuator hydraulic module 59 uses the line pressure PL generated by the clutch hydraulic module 46 as a source pressure, and upstream of the even-speed shift pressure Pe to the first actuator solenoid 71 and the second actuator solenoid 72 and the engagement pressure of the first clutch CA. The even-numbered gear stage pressure solenoid 77 and the first pressure reducing valve 770 (see FIG. 5) for generating the AXIS pressure, which is the pressure, the odd-numbered gear stage pressure Po to the third actuator solenoid 73 and the fourth actuator solenoid 74 and the second clutch CB An odd-numbered speed step pressure solenoid 78 and a second pressure reducing valve 780 that produce an AXIS pressure that is an upstream pressure of the engagement pressure are provided. The solenoids 77 and 78 and the pressure reducing valves 770 and 780 are pressure reducing valves configured by VBS (variable bleed solenoid).

  The clutch hydraulic module 46 includes a line pressure solenoid (not shown) that regulates the line pressure PL based on the oil discharged from the oil pump 4 and a pressure regulator valve 800 (see FIG. 5) that is operated by the line pressure solenoid. Further, a first clutch engagement pressure solenoid 81 and a first clutch engagement pressure control valve 810 that produce a clutch engagement pressure to the first clutch CA based on the even-numbered shift stage pressure Pe from the actuator hydraulic module 59, and an odd-number shift stage pressure Po. The second clutch engagement pressure solenoid 82 and the second clutch engagement pressure control valve 820 for generating the clutch engagement pressure to the second clutch CB based on the above.

  In addition, a first clutch pressure sensor 83 that detects a first clutch engagement pressure that is downstream of the first clutch engagement pressure control valve 810 and a second clutch engagement pressure that is downstream of the second clutch engagement pressure control valve 820 are used. And a second clutch pressure sensor 84 for detection. The solenoids 81 and 82 and the control valves 810 and 820 are constituted by a VFS (variable force solenoid).

  FIG. 5 is a schematic diagram showing a hydraulic pressure distribution between the oil pump 4 and the clutch. Since the first clutch CA and the second clutch CB are basically equivalent, the first clutch CA will be described as an example.

  The oil pump 4 driven by the engine ENG draws oil from an oil pan or the like (not shown) and discharges a flow rate corresponding to the rotational speed to the line pressure oil path LO. The pressure regulator valve 800 controls the opening amount of the drain port in accordance with the operating position, and an electromagnetic force acting on a line pressure solenoid (not shown) may be a control signal pressure using a pilot pressure as an original pressure, The flow rate is controlled by the balance position of the spring force and the feedback pressure.

  Specifically, the pump discharge flow rate is drained by a predetermined amount and regulated to a desired line pressure. When the actual line pressure is lower than the desired line pressure, the feedback pressure is reduced, the amount of opening of the drain port (x in FIG. 5) is reduced, and the pressure is adjusted so that the desired line pressure corresponding to the electromagnetic force is obtained. On the other hand, when the pump discharge flow rate becomes excessive, the opening amount of the drain port increases, and the pressure is adjusted so as to be a desired line pressure corresponding to the electromagnetic force. The pressure regulator valve 800 is designed to control a very large flow rate. In other words, the pressure regulator valve 800 has high vibration resistance against sudden changes in the flow rate. Therefore, a constant pressure can be supplied relatively stably even when the pump discharge flow rate changes suddenly or the consumption flow rate suddenly increases.

  The first pressure reducing valve 770 is a so-called normal close valve, and controls the flow rate (AXIS pressure) supplied to the oil passage L1 by the balance position of the electromagnetic force acting on the even-speed gear pressure solenoid 77, the spring force, and the feedback pressure. To do. Specifically, the flow rate supplied from the line pressure oil passage LO is branched into a drain port (× in FIG. 5) and the oil passage L1 downstream of the first pressure reducing valve 770 according to the operating position. Supply the oil pressure in the oil passage L1 to a desired pressure.

  Here, a flow path 771 that connects the oil path L0 and the oil path L1 is formed in the first pressure reducing valve 770, and exhibits an orifice function. Therefore, the amount of oil flowing into the oil passage L1 is likely to be limited. In addition, the volume of the oil passage L1 to be controlled is small because it is only the oil passage portion. Therefore, a slight flow rate fluctuation is easily reflected as a pressure, and when the changed pressure acts as a feedback pressure, there is a strong tendency to overshoot. That is, it can be said that the time constant in the response of the pressure to the flow rate change is short.

  The first clutch engagement pressure control valve 810 controls the flow rate (pressure) supplied to the oil passage L2 according to the balance position of the electromagnetic force acting on the first clutch engagement pressure solenoid 81, the spring force, and the feedback pressure. Specifically, the flow rate supplied from the oil passage L1 is changed according to the operation position, the drain port (X in FIG. 5) and the oil passage L2 (first clutch) downstream of the first clutch engagement pressure control valve 810. CA) and the oil passage L2 (first clutch CA) is adjusted to a desired pressure.

  Here, the volume of the first clutch to be controlled by the first clutch engagement pressure control valve 810 includes the volume of the cylinder chamber in addition to the oil passage L2 during the piston stroke. It is larger than the volume to be controlled. In addition, it has an element that absorbs a change in flow rate during the stroke of the clutch piston c1. Therefore, it can be said that the flow rate fluctuation is not easily reflected as pressure, and the tendency to overshoot is weak even when the feedback pressure is applied, that is, the time constant in the pressure response to the flow rate change is long.

  Thus, the reason why the AXIS pressure is created between the oil passage L0 and the oil passage L2 is to secure a double system in terms of the pressure control configuration. For example, even if the first clutch pressure solenoid 81 breaks down and the even gear side cannot be used, it is desired to continue the shift control using only the odd gear side. At this time, if the even speed stage and the odd speed stage have the same hydraulic pressure path, the engagement pressure is supplied to the first clutch CA and the second clutch CB is engaged. The control used cannot be achieved.

  Therefore, the hydraulic system on the even-numbered speed side and the hydraulic system on the odd-numbered speed side are divided and a pressure reducing valve is provided in each system. Thereby, for example, even if the first actuator solenoid 71 or the first clutch engagement pressure solenoid 81 fails, the engagement pressure is not supplied to the first clutch CA by closing the first pressure reducing valve 770. On the other hand, by opening the second pressure reducing valve 780, shifting can be continued on the odd-numbered speed side.

  Note that the AXIS pressure has no particular problem in control even if it is set to the same pressure as the line pressure if it is normal during normal control. However, in the first embodiment, a normally closed valve is used as the pressure reducing valve. When the pressure reducing valve is at the maximum opening, a high current value continues to flow through the gear stage pressure solenoid, and the durability of the solenoid decreases. Therefore, the AXIS pressure is not a line pressure, but a value higher than the clutch engagement pressure by a predetermined pressure (a pressure lower than the line pressure), thereby suppressing the amount of current supplied to the shift speed pressure solenoid. Although it is most advantageous to improve the durability to make the AXIS pressure the same as the clutch engagement pressure without adding the predetermined pressure, considering the stability of hydraulic control, the AXIS pressure is higher than the clutch engagement pressure. It is desirable to set a higher pressure.

  As shown in the lower part of FIG. 5, when each actuator is normal and in a steady state, the oil pressure in the oil passage L0 is a line pressure and is adjusted to the highest pressure. The oil pressure in the oil passage L1 is AXIS pressure, and is adjusted to a pressure lower than the line pressure and higher than the clutch engagement pressure. In the first embodiment, a pressure obtained by adding a predetermined pressure to the first clutch engagement pressure is controlled so as to ensure controllability while suppressing the energization amount of the solenoid and improving durability. The oil pressure in the oil passage L2 is a clutch engagement pressure, and is naturally regulated to a pressure lower than the AXIS pressure.

  In the first embodiment, as described above, the hydraulic control by the even-numbered shift pressure solenoid 77 and the first clutch engagement pressure solenoid 81 is controlled by the magnitude of the electromagnetic force, and both solenoids have a high hydraulic pressure when the electromagnetic force is large. When the electromagnetic force is small, control to a low oil pressure. Hereinafter, for the sake of explanation, when the same command signal is output to the even-numbered shift stage pressure solenoid 77 and the first clutch engagement pressure solenoid 81, the same hydraulic pressure is controlled. These may be configured to be inversely proportional to the electromagnetic force depending on the design specifications, and it goes without saying that the relationship between the electromagnetic force and the hydraulic pressure may be different for each solenoid by changing the spring set load or land area. .

[Hydraulic control processing]
Next, an outline of the hydraulic pressure control process for the actuator solenoids 71 to 74, the shift speed solenoids 77 and 78, and the clutch engagement pressure solenoids 81 and 82 will be described. In the following description, it is assumed that any one of the actuator solenoid, one of the shift speed pressure solenoids, and one of the clutch engagement pressure solenoids is appropriately selected and executed according to the state of the automatic manual transmission AMT.

FIG. 6 is a flowchart showing a basic control process of the hydraulic control process. Hereinafter, the first clutch CA and the second clutch CB are collectively referred to as a clutch, and each shift actuator is collectively referred to as a shift.
In step S1, it is determined whether or not the conditions for the clutch learning control process are satisfied. If the conditions are satisfied, the process proceeds to step S2 to output the learning request pressure to the gear position pressure solenoid and the clutch engagement pressure solenoid. Details of the clutch learning control process will be described later.

  In step S3, it is determined whether or not the request for clutch and shift is at rest, that is, in a neutral state. If in the neutral state, the process proceeds to step S4 to output standby pressure to all solenoids. The standby pressure means that a preliminary operation is performed so as to be able to respond immediately when a command signal is output. On the other hand, when it is determined that it is not in the neutral state, the process proceeds to step S5.

  In step S5, it is determined whether or not there is a shift operation. If it is determined that there is a shift operation, the process proceeds to step S6 to output a shift required pressure to the actuator solenoid so as to operate the shift actuator. On the other hand, when there is no shift operation, the current shift speed is maintained, so that the actuator solenoid is left as it is, and the clutch required pressure is output to the shift speed pressure solenoid and the clutch engagement pressure solenoid. Thus, when torque is transmitted in the clutch, a value obtained by adding a predetermined pressure to the clutch engagement pressure is set as the target value of the AXIS pressure, and a command signal corresponding to the value is output (normal control state).

[Clutch learning control processing]
Next, the clutch learning control process will be described. Here, in describing the logical configuration of the clutch learning control, a schematic system diagram of the clutch, the hydraulic circuit, and the automatic MT controller 47 will be described. FIG. 7 is a schematic system diagram of the clutch learning control, and FIG. 8 is a diagram showing command signals output during the learning control process. Although the configuration on the first clutch CA side will be described, the same applies to the second clutch CB.

  The first clutch CA transmits the driving force between the drive side rotating member CA1 to which the driving force from the engine ENG is input and the driven side rotating member CA2 connected to the first transmission input shaft 5, or the This is a wet multi-plate clutch type clutch mechanism that cuts off, and includes a plurality of drive plates a2, a plurality of driven plates b2, a friction material b3, a clutch piston c1, and a return spring c2. In FIG. 7, only the upper half of the rotation axis CL is shown.

  The clutch piston c1 is provided so as to be slidable in the direction of the rotation axis CL with respect to the drive side rotation member CA1. Also, a hydraulic chamber e is formed on one end surface of the clutch piston c1, and the clutch piston c1 is pressed by the hydraulic pressure of the hydraulic oil supplied to the hydraulic chamber e via the hydraulic oil passage f, so that the direction of the rotation axis CL To slide. The clutch of the first embodiment is configured such that the higher the hydraulic pressure supplied to the hydraulic chamber e, the more the clutch piston c1 moves to the A side in FIG.

  When the pressure acting on the hydraulic chamber e is smaller than the set load of the return spring c2, the clutch piston c1 is positioned at the end on the B side in FIG. 7, and at this time, the plates a2 and b2 have a predetermined distance d. Ensure clearance. This prevents the occurrence of drag torque when the clutch is released.

  Here, the movement of the clutch piston c1 to the A side in FIG. 7 with the position where the clutch piston c1 holds a gap of a predetermined distance d as a clearance with respect to the drive plate a2 as a reference position is hereinafter referred to as a stroke. When the stroke of the clutch piston c1 is 0, the clearance with respect to the drive plate a2 is a predetermined distance d. When the clutch piston c1 moves in the A direction, the clearance with respect to the drive plate a2 is a distance obtained by subtracting the stroke from the predetermined distance d. When becomes equal to d, the clutch piston c1 comes into contact with the drive plate a2 to enter a so-called half-clutch state (clutch slip state), and power transmission is started.

Within automatic MT controller 47, a first command signal output unit 471 for outputting a command signal to the first clutch engagement pressure solenoid 81, a second command signal for outputting a command signal for the even-shift-stage pressure solenoid 77 An output unit 472, a basic engagement pressure signal output unit 473 that outputs a basic engagement pressure signal to the first command signal output unit 471 and the second command signal output unit 472, a first command signal output unit 471, and a second command A rectangular wave auxiliary signal output unit 474 that outputs a rectangular wave auxiliary signal to the signal output unit 472. The first and second command signal output units 471 and 472 output a signal obtained by superimposing a basic fastening pressure signal and a rectangular wave auxiliary signal, which will be described later, to each solenoid (see FIG. 8C).

  Here, in normal fastening pressure control other than during learning control processing, the signal output from the second command signal output unit 472 is a signal that is higher than the signal output from the first command signal output unit 471. Is done. This predetermined pressure is a value at which the pressure in the oil passage L1 does not become excessive (avoidance of deterioration in fuel consumption) and the durability of the gear stage pressure solenoid can be improved as described above. In the first embodiment, a value obtained by adding a predetermined pressure to the first command signal is output as the second command signal. At the time of learning control, the predetermined pressure is set to a higher value than that at the time of normal fastening pressure control. Details will be described later.

The basic engagement pressure signal is one of hydraulic pressure signals for detecting the stroke of the clutch piston c1, and in the first embodiment, as shown in FIG. The magnitude of the signal value is set so that the current value, voltage value, etc.) increase . With such a setting, the position of the clutch piston c1 is defined by the magnitude of the hydraulic signal of the hydraulic oil.

  When the basic engagement pressure signal increases, the pressure acting on the hydraulic chamber e gradually increases, and when the pressure exceeds the specified pressure, the clutch piston c1 that has been stopped starts its stroke against the set load of the return spring c2. To do. That is, the basic engagement pressure signal is a command signal for causing the clutch piston c1 in a non-stroke state to stroke A in FIG. 7, and the gradient of the signal in FIG. That is, it is set to correspond to the stroke speed of the clutch piston c1. The initial value of the basic engagement pressure signal is set to a value at which the thrust generated in the clutch piston c1 is smaller than the set load.

  As shown in FIG. 8B, the rectangular wave auxiliary signal is a pulse signal applied to detect the stroke state, and the amplitude of the pulse signal is the minimum detectable by the first clutch pressure sensor 83. The amplitude is set to be larger than the magnitude of the hydraulic pressure signal corresponding to the magnitude of the hydraulic pressure fluctuation, but not so large as to change the stroke of the clutch piston c1 only by the pulse signal. Specifically, the amplitude is such that the clutch piston c1 starts and does not end with a single pulse signal.

  Further, in the first embodiment, the basic pressure corresponding to the case where the pressure in the hydraulic chamber e is controlled by the basic engagement pressure signal corresponding to the state where the stroke of the clutch piston c1 is 0 and the case where the stroke of the clutch piston c1 is zero. The amplitude of the pulse signal is set so that the influence on the driver does not increase when compared with the case where the pressure in the hydraulic chamber e is controlled by applying a pulse signal to the fastening pressure signal.

  Specifically, the first clutch pressure sensor 83 can detect and the magnitude of the torque output from the clutch immediately after the end of the stroke of the clutch piston c1 is not perceived by the driver as uncomfortable (that is, The amplitude is set such that it is not sensed at all, or even if it is sensed, it is slight and does not cause the driver to feel uncomfortable.

  In the clutch learning control process in the first embodiment, as will be described later, the clutch piston c1 is gradually stroked, and the start and end of the stroke are grasped by the fluctuation of the actual hydraulic pressure amplitude. That is, the command signal is stopped immediately after the clutch meet, and torque is slightly transmitted via the clutch immediately before the command signal is stopped. Therefore, the amplitude that does not give a sense of incongruity is set as described above. For example, with respect to the initial value of the hydraulic pressure controlled by the basic fastening pressure signal (about 180 kPa), the fluctuation of the hydraulic pressure controlled by the rectangular wave auxiliary signal is about 1/10 (about 20 kPa). Is set. Although it depends on the vehicle weight and the like, the amplitude is set so that the acceleration generated in the vehicle body by the minute torque transmission immediately after the clutch meet is less than 0.03G.

  Here, a logical configuration in which a stroke can be detected by the first clutch pressure sensor 83 by superimposing the basic engagement pressure signal and the rectangular wave auxiliary signal will be described. FIG. 9 is a graph showing temporal changes in the command signal and the actual clutch pressure during the clutch pressure learning control process, and FIG. 10 is an enlarged view in which the non-stroke region and the stroke region are extracted from the graph of FIG.

  When a minute hydraulic pressure change is applied when the clutch piston c1 is not moving, that is, when the clearance is a predetermined distance d, the hydraulic pressure change (hydraulic pressure fluctuation range h) is caused by a reaction force caused by the return spring c2 or friction. It is stored substantially as it is and is detected from the first clutch pressure sensor 83 (see FIG. 10A).

  However, during the stroke, a minute change in hydraulic pressure is absorbed by the volume change accompanying the stroke of the clutch piston c1. That is, a part of the energy given by the minute hydraulic pressure change is converted into the elastic energy of the return spring c2, the inertia energy of the clutch piston c1, the turbulent oil flow, and the thermal energy, and is detected by the first clutch pressure sensor 83. The actual hydraulic pressure fluctuation (hydraulic fluctuation width h) is reduced, and the fluctuation amount detected by the first clutch pressure sensor 83 is reduced (see FIG. 10B).

  Further, when the stroke amount increases and the clearance becomes zero, the hydraulic pressure fluctuation is substantially preserved by the reaction force of the drive plate a2 against a minute hydraulic pressure change and is detected from the first clutch pressure sensor 83. . That is, when the stroke of the clutch piston c1 ends, the actual hydraulic pressure variation detected by the first clutch pressure sensor 83 begins to be detected again.

  That is, a change between the hydraulic pressure fluctuation amplitude when the clutch piston c1 is not in a stroke and the hydraulic pressure fluctuation amplitude during the stroke is detected, thereby determining the start and end of the stroke.

  The stroke determination unit 475 detects the fluctuation range h of the actual hydraulic pressure input from the first clutch pressure sensor 83, and when the hydraulic pressure fluctuation range h changes from a predetermined value h2 or more to less than a predetermined value h1 (<h2). When it is determined that the stroke starts and the hydraulic pressure fluctuation range changes from less than the predetermined value h1 to the predetermined value h2 or more, it is determined that the stroke ends. Here, the predetermined value h1 <h2 is set to avoid an erroneous determination when the hydraulic pressure fluctuation range h is accidentally reduced or increased due to a disturbance or the like.

  The learning unit 476 learns the control hydraulic pressure of the clutch piston c1 at the start and end of the stroke from the determination result of the stroke determination unit 475. Here, the learned hydraulic pressure is the magnitude of the pressure corresponding to the basic engagement pressure signal output from the basic engagement pressure signal output unit 473 (or the first clutch pressure sensor 83 at the start and end of the stroke). The actual hydraulic pressure detected). The stroke start / end positions learned here are stored in the storage unit 477, which will be described later, and are used as a reference when start control and shift control are performed.

  A series of control from the output of a command signal to learning performed in the automatic MT controller 47 as described above is called learning control. In the learning unit 476, learning conditions (learning conditions) are set in order to improve the learning accuracy of learning control. Here, an indispensable condition for performing the learning control is that the first clutch CA is not transmitting torque. Other conditions include, for example, that the engine oil temperature is stable within a predetermined range, that there is no fluctuation in the engine speed (centrifugal force acting on the first clutch CA, hydraulic chamber e For example, the centrifugal hydraulic pressure in the engine is not fluctuating), the vehicle speed is 0, and the brake pedal is depressed (the vehicle is stopped and neutral idle control is being performed). When the learning condition is not satisfied, the learning control is immediately terminated.

[Problems in learning control processing]
In performing the learning control, a problem that occurs when a pressure reducing valve (shift pressure solenoid) is provided upstream of the clutch engagement pressure control valve (clutch engagement pressure solenoid) as in the first embodiment will be described. FIG. 11 is a time chart showing the time change of the hydraulic pressure fluctuation range h. As shown by the solid line in FIG. 11, when the hydraulic pressure fluctuation range h is equal to or greater than the predetermined value h2, it is determined that the stroke is not performed, and when the hydraulic pressure fluctuation range h is less than the predetermined value h1, it is determined that the stroke is performed. Therefore, the hydraulic pressure fluctuation range h needs to change by a predetermined amount Δh (= h2−h1) between the stroke and the non-stroke.

  However, when the learning control process is applied to a unit that actually includes a pressure reducing valve on the upstream side, depending on the unit, as shown by the one-dot chain line in FIG. The problem was found. If the predetermined amount Δh is not obtained, even if the learning control process is executed, it is not possible to determine the appropriate start / end of the stroke, and it cannot be used effectively for other controls. On the other hand, if the predetermined amount Δh is set small, the stroke can be determined. However, even if the hydraulic pressure fluctuation range h changes due to a slight disturbance or the like, the stroke is immediately determined, and the predetermined amount Δh is reduced. The method is not valid.

  Thus, as a result of intensive studies to investigate the cause of the above problem, it has been found that the presence of the pressure reducing valve is the cause of the small change Δh. Hereinafter, the influence of the pressure reducing valve will be described.

  FIG. 12 is a time chart showing the relationship between the first and second command signals in the unit where the problem occurs and the pressure actually generated. The first command signal corresponds to the actual clutch pressure, and the second command signal corresponds to the actual pressure in the oil passage L1. When verifying based on this time chart, there is a problem that the hydraulic pressure fluctuation range h detected at the time of non-stroke is sufficiently large and the hydraulic pressure fluctuation width h detected at the time of stroke becomes large. I understood.

  That is, when the waveform during the stroke is enlarged, the clutch pressure control valve operates in the valve opening direction to increase the flow rate when the rectangular wave rises, but the clutch includes the hydraulic chamber e and strokes. Therefore, the flow consumption is very large. At this time, when the pressure reducing valve is activated for the second command signal at the same time, the oil pressure in the oil passage L1 once rises, and then the flow rate consumed by the clutch is large. Will drop significantly. Along with this, the clutch pressure control valve operates in the valve opening direction to further increase the flow rate and tries to raise the actual clutch pressure.

  Then, the pressure reducing valve and the clutch pressure control valve are excessively opened to compensate for the shortage, and when the hydraulic pressure in the oil passage L1 on the upstream side starts to be secured, the actual clutch pressure on the downstream side becomes excessive. It tends to increase. At the falling edge of the rectangular wave, both the pressure reducing valve and the clutch pressure control valve are operated in the valve closing direction. Then, the excessive flow rate in the oil passage L1 suddenly loses its place, and since the volume of the oil passage L1 is small, the pressure is affected with a small flow rate (the time constant is short), so the overshoot in the oil passage L1. Will be caused. This also affects the clutch side, and a higher clutch actual pressure is generated.

  Thus, it has been found that when the pressure reducing valve is provided upstream of the clutch pressure control valve, the fluctuation range of the hydraulic pressure due to the rectangular wave is larger than the target fluctuation range (the predetermined amount Δh is reduced). Now, since the second command signal is controlled so that a pressure higher than the first command signal by a predetermined pressure is output (same as in normal control), the magnitude of the predetermined pressure is the hydraulic pressure fluctuation range. We decided to verify the impact. Note that, as described above, this predetermined pressure is a value that takes into consideration fuel efficiency and controllability during normal control, and is a value that does not cause a problem in other controls than learning control.

  FIG. 13 is a characteristic diagram showing the relationship of the hydraulic pressure fluctuation range h with respect to the magnitude of the predetermined pressure. When a predetermined pressure is applied to the first command signal and the second command signal is output, the relationship between the magnitude of the predetermined pressure and the hydraulic pressure fluctuation range h shows that the characteristic during the stroke is a region where the predetermined pressure is low. Then, it was found that the fluctuation range of the hydraulic pressure became larger according to the predetermined pressure, and then gradually decreased according to the predetermined pressure when the predetermined pressure became higher than the value used during normal control. On the other hand, the non-stroke characteristics, like the characteristics at the time of stroke, increase and decrease in hydraulic pressure, and then decrease. It turned out not to decline.

  Based on this characteristic, paying attention to the predetermined pressure that causes current problems, the hydraulic pressure fluctuation range h itself was sufficiently large, but the difference Δh between the stroke and the non-stroke is insufficient. found.

  Therefore, in the first embodiment, at the time of learning control, a predetermined pressure larger than that at the time of normal control is added to the first command signal, and the second command signal for learning control (third time) different from the normal second command signal. (Refer to the arrow in FIG. 13).

  FIG. 14 is a time chart showing the relationship between the first command signal for first and learning control of the first embodiment and the pressure actually generated. As shown in FIG. 14, the actual pressure in the oil passage L1 does not cause an overshoot, and accordingly the clutch actual pressure is not output high. This is because the second command signal is output higher, so the flow rate from the oil passage L0 to the oil passage L1 is smooth, and even if the flow rate is consumed on the clutch side due to the stroke, the oil flow in the oil passage L1 is excessive. This is considered to be due to the fact that a stable pressure is obtained without causing a decrease in oil pressure.

  As a result, an appropriate value can be obtained as the change amount Δh of the hydraulic pressure fluctuation range h, and the stroke determination can be appropriately performed in the learning control process.

As described above, the effects listed below can be obtained in the first embodiment.
(1) Friction engagement elements (clutch CA, CB) that press and fasten the friction plates (a2, b2) by the piston c1 stroked by the engagement pressure, and the engagement pressure that operates based on the command signal and controls the engagement pressure It operates based on a control valve (clutch pressure control valves 810, 820) and a second command signal (second command signal) that is higher than the pressure corresponding to the command signal, and controls the upstream pressure of the engagement pressure control valve. Pressure reducing valves (770, 780) to be controlled, pressure detecting means (clutch pressure sensors 83, 84) for detecting an actual engagement pressure that is downstream pressure of the engagement pressure control valve, and a basic engagement pressure with respect to the engagement pressure control valve A command signal in which the signal (473) and the rectangular wave auxiliary signal (474) are superimposed is output (first command signal output unit 471), and at the same time, a pressure higher than the pressure corresponding to the second command signal is applied to the pressure reducing valve. Outputs the third command signal (second command signal for learning control) , Equipped with a stroke position learning means (stroke determining unit 475, learning unit 476) for learning the stroke position of the piston based on the amplitude change of the actual pressure detected by the pressure detecting means.

  Therefore, even when a pressure reducing valve is provided on the upstream side, the change in the amplitude of the fastening pressure can be obtained, and appropriate stroke position learning can be performed. The second command signal represents a command signal during normal control and is a signal output during torque transmission of the clutch.

  (2) The third command signal (second command signal for learning control) is a command signal that is a pressure obtained by adding a predetermined pressure to a pressure corresponding to the command signal (first command signal). Therefore, it is possible to prevent the hydraulic pressure fluctuation range from becoming larger than expected without applying excessive pressure to the oil passage L1. Moreover, since only the predetermined pressure is changed, learning control can be executed without changing the basic logic of normal control.

  Next, Example 2 will be described. Since the basic configuration is the same as that of the first embodiment, only different points will be described. FIG. 15 is a time chart showing the relationship between the first command signal for first and learning control of the second embodiment and the pressure actually generated. The second embodiment is different from the first embodiment in that a second command signal for learning control, which is a constant value lower than the line pressure, is given instead of increasing the predetermined pressure.

  As shown in FIG. 15, the actual pressure in the oil passage L1 does not cause an overshoot, and accordingly the clutch actual pressure is not output high. This is because the second command signal for learning control is output higher, so the flow from the oil passage L0 to the oil passage L1 is smooth, and even if the clutch consumes the flow rate due to the stroke, excessive oil is supplied. This is considered to be due to the fact that a stable pressure is obtained without causing a decrease in the hydraulic pressure in the path L1.

  As a result, an appropriate value can be obtained as the change amount Δh of the hydraulic pressure fluctuation range h, and the stroke determination can be appropriately performed in the learning control process.

  Next, Example 3 will be described. Since the basic configuration is the same as that of the first embodiment, only different points will be described. FIG. 16 is a time chart showing the relationship between the first command signal for first and learning control of the third embodiment and the actually generated pressure. The third embodiment is different from the first embodiment in that instead of increasing the predetermined pressure as in the first embodiment, a second command signal for learning control which is a constant value higher than the line pressure is given.

As shown in FIG. 16 , the actual pressure in the oil passage L1 does not cause an overshoot, and accordingly the clutch actual pressure is not output high. This is because the second command signal for learning control is output at a higher level, so that the flow rate supply from the oil passage L0 to the oil passage L1 is equal to the state where there is no pressure reducing valve (the orifice effect component exists). It is considered that even if the flow rate is consumed on the clutch side due to the stroke, the oil pressure in the oil passage L1 is not excessively reduced and a stable pressure is obtained.

  As a result, an appropriate value can be obtained as the change amount Δh of the hydraulic pressure fluctuation range h, and the stroke determination can be appropriately performed in the learning control process.

  Next, Example 4 will be described. Since the basic configuration is the same as that of the first embodiment, only different points will be described. FIG. 17 is a time chart showing the relationship between the first command signal for first and learning control of the fourth embodiment and the actually generated pressure. The fourth embodiment is different from the first embodiment in that, instead of increasing the predetermined pressure as in the first embodiment, a command signal that minimizes the pressure reduction amount of the pressure reducing valve is given as a second command signal for learning control. In other words, the opening of the pressure reducing valve is maximized and the pressure in the oil passage L1 is controlled to be the same as the line pressure. However, as described in the first embodiment, since the pressure reducing valve has an orifice effect, it is not certain whether or not it completely matches the line pressure, but the pressure in the oil passage L1 remains stable.

  As shown in FIG. 17, the actual pressure in the oil passage L1 does not cause an overshoot, and accordingly, the clutch actual pressure is not output high. This is because the second command signal for learning control is output at a high level, so even if the flow rate is consumed on the clutch side due to the stroke, the oil pressure in the oil passage L1 is not excessively reduced and stable pressure is maintained. This is considered to be due to the fact that

  As a result, an appropriate value can be obtained as the change amount Δh of the hydraulic pressure fluctuation range h, and the stroke determination can be appropriately performed in the learning control process.

  Next, Example 5 will be described. Since the basic configuration is the same as that of the first embodiment, only different points will be described. FIG. 18 is a characteristic diagram showing the relationship between the hydraulic pressure fluctuation width h and the magnitude of the predetermined pressure (the characteristics themselves are the same as those in FIG. 13). When a predetermined pressure is applied to the first command signal and the second command signal is output, the relationship between the magnitude of the predetermined pressure and the hydraulic pressure fluctuation range h shows that the predetermined pressure is low as a characteristic during stroke. It was found that when the pressure was gradually increased, the hydraulic pressure fluctuation range increased and then decreased. On the other hand, it was found that when the same change was made as the characteristic at the time of non-stroke, the fluctuation range of hydraulic pressure became larger and then decreased, but not as much as at the time of stroke.

  Based on this characteristic, paying attention to the predetermined pressure that causes current problems, the hydraulic pressure fluctuation range h itself was sufficiently large, but the difference Δh between the stroke and the non-stroke is insufficient. found. Here, the overshoot in the oil passage L1 can be considered as a reason why the difference Δh cannot be secured. This overshoot is caused by the time constant in the oil passage L1 being smaller than the time constant of the oil passage L2 (including the clutch). That is, when a rectangular wave auxiliary signal is given to the pressure reducing valve, the time constant is short, and therefore it is considered that amplitude overshoot is caused by vibration of the rectangular wave auxiliary signal.

  Therefore, in the fifth embodiment, at the time of learning control, the same predetermined pressure as that at the time of normal control is added to the basic fastening pressure signal without adding the rectangular wave auxiliary signal to the second command signal output to the pressure reducing valve. The second command signal for learning control (corresponding to the third command signal) different from the normal second command signal is output (see the arrow in FIG. 18).

  The one-dot chain line in FIG. 18 is a characteristic diagram showing the relationship between the hydraulic pressure fluctuation width h and the predetermined pressure when the rectangular wave auxiliary signal is not given. It can be seen that the hydraulic pressure fluctuation range h is suppressed to a small value even when the same predetermined pressure as that in the normal control is added to the basic fastening pressure signal without giving the rectangular wave auxiliary signal.

  FIG. 19 is a time chart showing the relationship between the first command signal for first and learning control of the fifth embodiment and the actually generated pressure. As shown in FIG. 19, the actual pressure in the oil passage L1 does not cause an overshoot, and accordingly the clutch actual pressure is not output high. This is considered to be due to the fact that since the rectangular wave auxiliary signal is not given to the pressure reducing valve, overshoot is avoided and a stable pressure is obtained even when the time constant of the oil passage L1 is small.

  As a result, an appropriate value can be obtained as the change amount Δh of the hydraulic pressure fluctuation range h, and the stroke determination can be appropriately performed in the learning control process.

  Although the first embodiment has been described above, the present invention is not limited to the above-described configuration, and the effects of the present invention may be achieved by other configurations. For example, although applied to the clutch control of the twin clutch type automatic manual transmission in the first embodiment, the present invention may be applied to a fastening element of a normal planetary gear type automatic transmission. Further, the present control may be applied to a brake in which only one of the fastening side and the fastened side rotates like a clutch, and only one of them rotates.

1 is a system diagram showing a power train of a vehicle equipped with a twin clutch type automatic manual transmission according to Embodiment 1. FIG. BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is an overall system diagram showing a twin-clutch automatic manual transmission to which a control device for a frictional engagement element of Example 1 is applied. FIG. 3 is a hydraulic circuit diagram illustrating a sequence solenoid OFF in the actuator hydraulic module and the clutch hydraulic module according to the first embodiment. FIG. 3 is a hydraulic circuit diagram illustrating a sequence solenoid ON in the actuator hydraulic module and the clutch hydraulic module according to the first embodiment. It is the schematic showing the oil pressure distribution between the oil pump of Example 1, and a clutch. 3 is a flowchart illustrating basic control processing of hydraulic control processing according to the first embodiment. FIG. 2 is a schematic system diagram of clutch learning control according to the first embodiment. It is a figure showing the command signal output at the time of the learning control process of Example 1. FIG. It is a graph showing the time change of the command signal at the time of the clutch pressure learning control process of Example 1, and an actual clutch pressure. It is the enlarged view which extracted the area | region at the time of a non-stroke and the area | region at the time of a stroke from the graph of FIG. 3 is a time chart showing a time change of a hydraulic pressure fluctuation range according to the first embodiment. It is a time chart showing the relationship between the 1st and 2nd command signal in the unit in which the subject has arisen, and the pressure which has actually occurred. It is a characteristic view which shows the relationship of the oil pressure fluctuation range with respect to the magnitude | size of a predetermined pressure. 3 is a time chart illustrating a relationship between a first command signal for first and learning control according to the first embodiment and an actually generated pressure. 6 is a time chart showing the relationship between the first command signal for first and learning control of Example 2 and the pressure actually generated. 12 is a time chart showing the relationship between the first command signal for first and learning control in Example 3 and the pressure that is actually generated. 10 is a time chart showing the relationship between the first command signal for first and learning control in Example 4 and the pressure actually generated. It is a characteristic view which shows the relationship of the oil pressure fluctuation range with respect to the magnitude | size of a predetermined pressure. FIG. 10 is a time chart illustrating a relationship between a first command signal for first and learning control in Example 5 and a pressure that is actually generated.

Explanation of symbols

45 Actuator unit 46 Clutch hydraulic module 47 Automatic MT controller 59 Actuator hydraulic module 77 Even gear step pressure solenoid 78 Odd gear step pressure solenoid 81 First clutch engagement pressure solenoid 82 Second clutch engagement pressure solenoid 83 First clutch pressure sensor 84 Second Clutch pressure sensor
471 First command signal output section
472 Second command signal output section
473 Basic fastening pressure signal output section
474 Square wave auxiliary signal output section
475 Stroke judgment unit
476 Learning Department
477 Memory
770 First pressure reducing valve
780 Second pressure reducing valve
810 First clutch engagement pressure control valve
820 Second clutch engagement pressure control valve
AMT twin clutch automatic manual transmission
c1 clutch piston
CA 1st clutch
CB 2nd clutch

Claims (5)

A friction fastening element that presses and fastens the friction plate with a piston that strokes due to the fastening pressure;
A fastening pressure control valve for controlling the fastening pressure based on a first command signal;
A pressure reducing valve for controlling the upstream pressure before Symbol fastening pressure control valve based on the second command signal,
A line pressure control valve that controls a line pressure that is an upstream pressure of the pressure reducing valve based on a line pressure command signal;
Pressure detecting means for detecting an actual fastening pressure which is a downstream pressure of the fastening pressure control valve;
And a stroke position learning means for learning the stroke position before Symbol piston,
During normal engagement pressure control in which the stroke position learning means does not perform the learning, the second command signal is obtained by setting the upstream pressure of the engagement pressure control valve to be higher than the pressure corresponding to the first command signal by a predetermined pressure. A normal control signal commanding to be lower than the line pressure,
The stroke position learning means includes
The first command signal is a signal obtained by applying an auxiliary signal to a basic fastening pressure signal that defines a stroke position of the piston,
The second command signal is a learning control signal that commands the upstream pressure of the engagement pressure control valve to be higher than that during the normal engagement pressure control,
The friction engagement element control device according to claim 1, wherein the stroke position of the piston is learned based on an amount of variation caused by the auxiliary signal in the actual engagement pressure detected by the pressure detection means . The control apparatus for a frictional engagement element according to claim 1,
The stroke position learning means includes
2. The friction engagement element control device according to claim 1, wherein the second command signal is a signal for commanding an upstream pressure of the engagement pressure control valve to be higher than the line pressure . The control apparatus for a frictional engagement element according to claim 1,
The stroke position learning means includes
The second command signal is set such that the upstream pressure of the fastening pressure control valve is higher than the pressure corresponding to the first command signal by a second predetermined pressure higher than the predetermined pressure and lower than the line pressure. A control device for a frictional engagement element, characterized in that a signal for instructing to become is used. In the control apparatus of the frictional engagement element according to claim 1 or 2,
The stroke position learning means includes
2. The friction engagement element control device according to claim 1, wherein the second command signal is a signal for commanding the upstream pressure of the engagement pressure control valve to be constant. A friction fastening element that presses and fastens the friction plate with a piston that strokes due to the fastening pressure;
A fastening pressure control valve for controlling the fastening pressure based on a first command signal;
A pressure reducing valve for controlling the upstream pressure before Symbol fastening pressure control valve based on the second command signal,
A line pressure control valve that controls a line pressure that is an upstream pressure of the pressure reducing valve based on a line pressure command signal;
Pressure detecting means for detecting an actual fastening pressure which is a downstream pressure of the fastening pressure control valve;
And a stroke position learning means for learning the stroke position before Symbol piston,
During normal engagement pressure control in which the stroke position learning means does not perform the learning, the second command signal is obtained by setting the upstream pressure of the engagement pressure control valve to be higher than the pressure corresponding to the first command signal by a predetermined pressure. A normal control signal commanding to be lower than the line pressure,
The stroke position learning means includes
The first command signal is a signal obtained by applying an auxiliary signal to a basic fastening pressure signal that defines a stroke position of the piston,
The second command signal is a learning control signal that commands the upstream pressure of the fastening pressure control valve to be higher than the pressure corresponding to the basic fastening pressure signal by the predetermined pressure,
The friction engagement element control device according to claim 1, wherein the stroke position of the piston is learned based on an amount of variation caused by the auxiliary signal in the actual engagement pressure detected by the pressure detection means .

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