Patent Publication Number: US-8543299-B2

Title: Control device for automatic transmission

Description:
INCORPORATION BY REFERENCE 
     This is a Continuation of application Ser. No. 12/314,814 filed Dec. 17, 2008, which claims the benefit of Japanese Application No. 2007-338146 filed on Dec. 27, 2007. The disclosure of the prior applications is hereby incorporated by reference herein in its entirety. 
    
    
     BACKGROUND 
     The present invention relates to a control device, method and storage medium for an automatic transmission. 
     There exists a multi-stage automatic transmission mounted on a vehicle or the like. Hydraulic control is performed to engage a plurality of (for example, two) friction engagement elements (clutch or brake) according to a shift speed in order to form a transmission path of a speed change gear mechanism. However, for example, when a hydraulic pressure is output due to a failure (disconnection, or valve sticking, for example) of a linear solenoid valve to a hydraulic servo as a friction engagement element which should be released, there is a possibility that the friction engagement element which should be released is simultaneously engaged in addition to the friction engagement element engaged in a normal state. 
     Thus, in order to prevent the simultaneous engagement, a configuration has been proposed in which a plurality of so-called cut-off valves which block a hydraulic pressure (source pressure) of an additional friction engagement element when an engagement pressure of friction engagement elements engaged in the normal state is input are provided according to the combination of the friction engagement elements to be engaged in each shift speed to thereby prevent the simultaneous engagement in all shift speeds (for example, see Japanese Patent Application Publication No. JP-A-2003-336731). 
     SUMMARY 
     In recent years, improvements in fuel efficiency of vehicles are required due to environmental issues and the like. For example, multi-speed automatic transmissions are required also for small vehicles. Therefore, a number of friction engagement elements for forming shift speeds become necessary even for the automatic transmissions for which size reduction is required. However, when the cut-off valve prevents the simultaneous engagement, there has been a problem in that the number of the necessary cut-off valves increases according to the increase in shift speeds to not only inhibit the reduction in size of a hydraulic control device but also inhibit the reduction in weight and cost. 
     Thus, it is an object of the present invention to provide a control device for an automatic transmission which can prevent a simultaneous engagement without using a cut-off valve to reduce the size, weight, and cost. The present invention can also achieve various other advantages. 
     An exemplary control device for an automatic transmission that includes an automatic speed change mechanism with an input shaft connected to a driving source and an output shaft connected to a drive wheel and a plurality of friction engagement elements engaged based on an engagement pressure supplied to each hydraulic servo, in which a transmission path between the input shaft and the output shaft is changed based on engaged states of two of the plurality of friction engagement elements to form a plurality of shift speeds; and a regulated pressure supply portion that can freely regulate a line pressure to be supplied as the engagement pressure individually to each of the hydraulic servos, and which selectively engages the friction engagement elements in accordance with the shift speeds, the control device includes an input torque detection unit that detects an input torque input to the input shaft; and a controller that: determines torque distribution of two friction engagement elements that form the shift speeds; and calculates a transmission torque of the two friction engagement elements based on the input torque and the torque distribution and sets the engagement pressure to obtain a torque capacity that can transmit the transmission torque, wherein the controller sets the engagement pressure such that slippage does not occur in the two friction engagement elements in a state where engagement of the two friction engagement elements forms the shift speeds and such that, even if an additional friction engagement element engages based on the line pressure while the two friction engagement elements are engaged, one of the three friction engagement elements is caused to slip. 
     An exemplary method of operating an automatic transmission that includes an automatic speed change mechanism with an input shaft connected to a driving source and an output shaft connected to a drive wheel and a plurality of friction engagement elements engaged based on an engagement pressure supplied to each hydraulic servo, in which a transmission path between the input shaft and the output shaft is changed based on engaged states of two of the plurality of friction engagement elements to form a plurality of shift speeds; and a regulated pressure supply portion that can freely regulate a line pressure to be supplied as the engagement pressure individually to each of the hydraulic servos, and that selectively engages the friction engagement elements in accordance with the shift speeds, the method includes detecting an input torque input to the input shaft with a detector; determining torque distribution of two friction engagement elements that form the shift speeds with a controller; and calculating a transmission torque of the two friction engagement elements based on the input torque and the torque distribution and sets the engagement pressure to obtain a torque capacity that can transmit the transmission torque with a controller, wherein the engagement pressure is set such that slippage does not occur in the two friction engagement elements in a state where engagement of the two friction engagement elements forms the shift speeds and such that, even if an additional friction engagement element engages based on the line pressure while the two friction engagement elements are engaged, one of the three friction engagement elements is caused to slip. 
     An exemplary computer readable storage medium storing a set of program instructions for operating an automatic transmission that includes an automatic speed change mechanism with an input shaft connected to a driving source and an output shaft connected to a drive wheel and a plurality of friction engagement elements engaged based on an engagement pressure supplied to each hydraulic servo, in which a transmission path between the input shaft and the output shaft is changed based on engaged states of two of the plurality of friction engagement elements to form a plurality of shift speeds; and a regulated pressure supply portion that can freely regulate a line pressure to be supplied as the engagement pressure individually to each of the hydraulic servos, and which selectively engages the friction engagement elements in accordance with the shift speeds, the program including instructions for: detecting an input torque input to the input shaft; determining torque distribution of two friction engagement elements that form the shift speeds; and calculating a transmission torque of the two friction engagement elements based on the input torque and the torque distribution and sets the engagement pressure to obtain a torque capacity that can transmit the transmission torque, wherein the engagement pressure is set such that slippage does not occur in the two friction engagement elements in a state where engagement of the two friction engagement elements forms the shift speeds and such that, even if an additional friction engagement element engages based on the line pressure while the two friction engagement elements are engaged, one of the three friction engagement elements is caused to slip. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Various exemplary aspects of the invention will be described with reference to the drawings, wherein: 
         FIG. 1  is a block diagram showing a control device for an automatic transmission according to the present invention; 
         FIG. 2  is a skeleton view showing an automatic transmission to which the present invention can be applied; 
         FIG. 3  is an engagement table of an automatic speed change mechanism; 
         FIG. 4  is a velocity diagram of the automatic speed change mechanism; 
         FIG. 5  is a circuit diagram showing a hydraulic control device for the automatic transmission; 
         FIG. 6  is a flowchart showing a calculation method of an engagement pressure in a normal state; 
         FIG. 7  is a view showing a torque capacity of each friction engagement element with respect to a drive wheel in a failure state; and 
         FIGS. 8A and 8B  are illustrative views of one example of torque application in the normal state and in the failure state,  FIG. 8A  showing the torque application in a fourth forward speed in the normal state, and  FIG. 8B  showing the torque application in the failure state where a clutch C- 3  is engaged in the fourth forward speed. 
     
    
    
     DETAILED DESCRIPTION OF EMBODIMENTS 
     Hereinafter, an embodiment of the present invention will be described with reference to  FIGS. 1 to 8B . 
     [Schematic Configuration of Automatic Transmission] 
     First, the schematic configuration of an automatic transmission  3  to which the present invention can be applied will be described with reference to  FIG. 2 . For example, as shown in  FIG. 2 , the automatic transmission  3  suitable for use in a front-engine front-wheel drive (FF) vehicle has an input shaft  8  of the automatic transmission connectable to an engine (driving source)  2  (see  FIG. 1 ), and includes a torque converter  4  and an automatic speed change mechanism  5  with the axis of the input shaft  8  as the center. 
     The torque converter  4  has a pump impeller  4   a  connected to the input shaft  8  of the automatic transmission  3 , and a turbine runner  4   b  to which the rotation of the pump impeller  4   a  is transmitted via a working fluid. The turbine runner  4   b  is connected to an input shaft  10  of the automatic speed change mechanism  5  arranged coaxially with the input shaft  8 . The torque converter  4  includes a lockup clutch  7 . When the lockup clutch  7  is engaged, the rotation of the input shaft  8  of the automatic transmission  3  is directly transmitted to the input shaft  10  of the automatic speed change mechanism  5 . 
     The automatic speed change mechanism  5  includes a planetary gear (deceleration planetary gear) SP and a planetary gear unit (planetary gear set) PU on the input shaft  10 . The planetary gear SP is a so-called single pinion planetary gear including a sun gear S 1 , a carrier CR 1 , and a ring gear R 1 , and having a pinion P 1  that meshes with the sun gear S 1  and the ring gear R 1  in the carrier CR 1 . 
     The planetary gear unit PU is a so-called Ravigneaux-type planetary gear having a sun gear (third rotational element) S 2 , a sun gear (first rotational element) S 3 , a carrier (second rotational element) CR 2 , and a ring gear (fourth rotational element) R 2  as four rotational elements, and having a long pinion PL which meshes with the sun gear S 2  and the ring gear R 2  and a short pinion PS which meshes with the sun gear S 3  to mesh with each other in the carrier CR 2 . 
     The sun gear S 1  of the planetary gear SP is connected to a boss portion fixed integrally with a transmission case  9  such that the rotation is stopped. The ring gear R 1  is subjected to the same rotation (hereinafter called “input rotation”) as the rotation of the input shaft  10 . Further, the carrier CR 1  is subjected to decelerated rotation decelerated from the input rotation by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation, and is connected to a clutch (first clutch as a friction engagement element) C- 1  and a clutch (third clutch as a friction engagement element) C- 3 . 
     The sun gear S 2  of the planetary gear unit PU is connected to a brake (first brake as a friction engagement element) B- 1  formed of a band brake to be fixable with respect to the transmission case, and is connected to the clutch C- 3  such that the decelerated rotation of the carrier CR 1  can be input via the clutch C- 3 . The sun gear S 3  is connected to the clutch C- 1  such that the decelerated rotation of the carrier CR 1  can be input. Note that the brake B- 1  has a brake band  19  provided around a drum-shaped member  18  connected to the clutch C- 3  and the sun gear S 2 . The brake band  19  has one end fixed to the case  9  and the other end drive-connected to a hydraulic servo  44  described later (see  FIG. 5 ), so as to be wound around the drum-shaped member  18  by driving the hydraulic servo  44 . The winding direction of the brake band  19  is arranged to be in the opposite direction of the rotational direction of the drum-shaped member  18  in a second forward speed to a sixth forward speed. That is, the brake band  19  is pulled by the hydraulic servo  44  to be wound in the opposite direction with respect to the rotational direction of the drum-shaped member  18  in the second forward speed to the sixth forward speed. 
     Further, the carrier CR 2  is connected to a clutch (second clutch as a friction engagement element) C- 2  to which the rotation of the input shaft  10  is input such that the input rotation can be input via the clutch C- 2 . Also, the clutch C- 2  is connected to a one-way clutch F- 1  and a brake (second brake as a friction engagement element) B- 2  such that rotation in one direction with respect to the transmission case is restricted via the one-way clutch F- 1  and the rotation is stoppable via the brake B- 2 . The ring gear R 2  is connected to a counter gear (output shaft)  11 , and the counter gear  11  is connected to a drive wheel via a counter shaft and a differential device (not shown). 
     [Operation in Each Shift Speed in Automatic Transmission] 
     Next, based on the configuration described above, the application of the automatic speed change mechanism  5  will be described with reference to  FIGS. 2 ,  3 , and  4 . Note that, in a velocity diagram shown in  FIG. 4 , the ordinate direction shows the rotational speeds of the respective rotational elements (respective gears), and the abscissa direction shows the gear ratios of the respective rotational element. In a portion of the planetary gear SP in the velocity diagram, the ordinates respectively correspond to the sun gear S 1 , the carrier CR 1 , and the ring gear R 1  in order from the left side in  FIG. 4 . Further, in a portion of the planetary gear unit PU in the velocity diagram, the ordinates respectively correspond to the sun gear S 3 , the ring gear R 2 , the carrier CR 2 , and the sun gear S 2  in order from the right side in  FIG. 4 . 
     For example, in a first (1ST) forward speed in a drive (D) range, the clutch C- 1  and the one-way clutch F- 1  are engaged, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation is input to the sun gear S 3  via the clutch C- 1 . The rotation of the carrier CR 2  is restricted to one direction (positive rotation direction), i.e., the carrier CR 2  is prevented from a reverse rotation and fixed. The decelerated rotation input to the sun gear S 3  is output to the ring gear R 2  via the fixed carrier CR 2 , and the positive rotation as the first forward speed is output from the counter gear  11 . 
     At the time of an engine braking (during coasting), the state of the first forward speed is maintained while the brake B- 2  is locked to fix the carrier CR 2  such that the positive rotation of the carrier CR 2  is prevented. In the first forward speed, since the one-way clutch F- 1  prevents the reverse rotation and allows the positive rotation of the carrier CR 2 , the first forward speed when switched from a non-drive range to the drive range, for example, can be achieved smoothly by the automatic engagement of the one-way clutch F- 1 . 
     In the second (2ND) forward speed, the clutch C- 1  is engaged and the brake B- 1  is locked, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation is input to the sun gear S 3  via the clutch C- 1 . The locking of the brake B- 1  stops the rotation of the sun gear S 2 . The carrier CR 2  is subjected to the decelerated rotation lower than that of the sun gear S 3 , the decelerated rotation input to the sun gear S 3  is output to the ring gear R 2  via the carrier CR 2 , and a positive rotation as the second forward speed is output from the counter gear  11 . 
     In the third (3RD) forward speed, the clutch C- 1  and the clutch C- 3  are engaged, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation is input to the sun gear S 3  via the clutch C- 1 . The decelerated rotation of the carrier CR 1  is input to the sun gear S 2  by the engagement of the clutch C- 3 . That is, since the decelerated rotation of the carrier CR 1  is input to the sun gear S 2  and the sun gear S 3 , the planetary gear unit PU is brought to a directly-connected state in the decelerated rotation, such that the decelerated rotation is directly output to the ring gear R 2  and a positive rotation as the third forward speed is output to the counter gear  11 . 
     In the fourth (4TH) forward speed, the clutch C- 1  and the clutch C- 2  are engaged, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation is input to the sun gear S 3  via the clutch C- 1 . The input rotation is input to the carrier CR 2  by the engagement of the clutch C- 2 . By the decelerated rotation input to the sun gear S 3  and the input rotation input to the carrier CR 2 , a decelerated rotation higher than that in the third forward speed is output to the ring gear R 2 , and a positive rotation as the fourth forward speed is output from the counter gear  11 . 
     In the fifth (5TH) forward speed, the clutch C- 2  and the clutch C- 3  are engaged, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected to the input rotation is input to the sun gear S 2  via the clutch C- 3 . The input rotation is input to the carrier CR 2  by the engagement of the clutch C- 2 . By the decelerated rotation input to the sun gear S 2  and the input rotation input to the carrier CR 2 , an increased rotation slightly higher than the input rotation is output to the ring gear R 2 , and a positive rotation as the fifth forward speed is output from the counter gear  11 . 
     In the sixth (6TH) forward speed, the clutch C- 2  is engaged and the brake B- 1  is locked, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the input rotation is input to the carrier CR 2  by the engagement of the clutch C- 2 . The locking of the brake B- 1  stops the rotation of the sun gear S 2 . Due to the fixed sun gear S 2 , the input rotation of the carrier CR 2  becomes an increased rotation higher than that in the fifth forward speed to be output to the ring gear R 2 , and a positive rotation as the sixth forward speed is output from the counter gear  11 . 
     In the first reverse speed (REV), the clutch C- 3  is engaged and the brake B- 2  is locked, as shown in  FIG. 3 . As shown in  FIGS. 2 and 4 , the rotation of the carrier CR 1  decelerated by the fixed sun gear S 1  and the ring gear R 1  subjected the input rotation is input to the sun gear S 2  via the clutch C- 3 . The locking of the brake B- 2  stops the rotation of the carrier CR 2 . The decelerated rotation input to the sun gear S 2  is output to the ring gear R 2  via the fixed carrier CR 2 , and a reverse rotation as the first reverse speed is output from the counter gear  11 . 
     Note that the clutch C- 1 , the clutch C- 2 , and the clutch C- 3  are released in a parking (P) range and a neutral (N) range, for example. The carrier CR 1  and the sun gear S 2  as well as the sun gear S 3 , i.e., the planetary gear SP and the planetary gear unit PU, are brought to a disconnected state, and the input shaft  10  and the carrier CR 2  are brought to a disconnected state. Accordingly, the input shaft  10  and the planetary gear unit PU are brought to a disconnected state regarding power transmission, and the input shaft  10  and the counter gear  11  are brought to a disconnected state regarding power transmission. 
     [Schematic Configuration of Hydraulic Control Device] 
     Next, a hydraulic control device  6  for the automatic transmission according to the present invention will be described. First, portions (not shown) for generating a line pressure, secondary pressure, modulator pressure, range pressure, and the like in the hydraulic control device  6  will be described roughly. Note that the portions for generating the line pressure, secondary pressure, modulator pressure, and range pressure are similar to those of a general hydraulic control device for an automatic transmission and well-known, and therefore will be described briefly. 
     For example, the hydraulic control device  6  includes an oil pump, a manual shift valve, a primary regulator valve, a secondary regulator valve, a solenoid modulator valve, a linear solenoid valve, and the like (not shown). For example, when the engine is started, the oil pump connected to be rotationally driven by the pump impeller  4   a  of the torque converter  4  is driven in combination with the rotation of the engine to generate a hydraulic pressure such that oil is sucked up from an oil pan (not shown) via a strainer. 
     The hydraulic pressure generated by the oil pump is regulated to a line pressure P L  while the discharge is regulated by the primary regulator valve based on a signal pressure P SLT  of the linear solenoid valve, which is regulated and output according to the opening degree of a throttle. The line pressure P L  is supplied to the manual shift valve, the solenoid modulator valve, and more specifically to a linear solenoid valve SLC 3  and the like described later. The line pressure P L  supplied to the solenoid modulator valve is regulated to a modulator pressure P MOD  that is approximately a constant pressure by the valve. The modulator pressure P MOD  is supplied as a source pressure of the linear solenoid valve, and more specifically of solenoid valves S 11 , S 12 , and the like described later. 
     Note that the pressure discharged from the primary regulator valve is regulated to a secondary pressure P SEC  while the discharge is further regulated by the secondary regulator valve, for example. The secondary pressure P SEC  is supplied to, for example, a lubricating oil path and an oil cooler, and is also supplied to the torque converter  4 , and is used for control of the lockup clutch  7 . 
     The manual shift valve (not shown) has a spool driven mechanically (or electrically) by a shift lever provided to a driver&#39;s seat (not shown). By switching the position of the spool according to a shift range (for example, P, R, N, or D) selected by the shift lever, an output state or a non-output state (drainage) of the input line pressure P L  is set. 
     More specifically, when the shift lever is operated to the D range, an input port to which the line pressure P L  is input and a forward range pressure output port are communicated based on the position of the spool, and the line pressure P L  is output from the forward range pressure output port as a forward range pressure (D range pressure) P D . When the shift lever is operated to the reverse (R) range, the input port and a reverse range pressure output port are communicated based on the position of the spool, and the line pressure P L  is output from the reverse range pressure output port as a reverse range pressure (R range pressure) P REV . When the shift lever is operated to the P range or the N range, the input port and the forward range pressure output port as well as the reverse range pressure output port are disconnected by the spool, and the forward range pressure output port and the reverse range pressure output port are communicated with a drain port, i.e., are brought to non-output states where the D range pressure P D  and the R range pressure P REV  are drained (discharged). 
     [Detailed Configuration of Shift Controlling Portion of Hydraulic Control Device] 
     Next, portions for mainly performing a shift control in the hydraulic control device  6  according to the present invention will be described with reference to  FIG. 5 . Note that, in this embodiment, in order to describe the position of the spool, a position in the right half is called a “right half position” and a position in the left half is called a “left half position” in  FIG. 5 . 
     The hydraulic control device  6  is structured to include four linear solenoid valves SLC 1 , SLC 2 , SLC 3 , and SLB 1  for directly supplying output pressures regulated as engagement pressures respectively to a total of five hydraulic servos of a hydraulic servo  41  of the clutch C- 1 , a hydraulic servo  42  of the clutch C- 2 , a hydraulic servo  43  of the clutch C- 3 , the hydraulic servo  44  of the brake B- 1 , and a hydraulic servo  45  of the brake B- 2 . The hydraulic control device  6  is also structured to include a solenoid valve S 11 , a solenoid valve S 12 , a first clutch apply relay valve  21 , a second clutch apply relay valve  22 , a C- 2  relay valve (switching portion)  23 , a B- 2  relay valve  24 , and the like as portions for achieving a limp-home function and switching the pressure output of the linear solenoid valve SLC 2  to the hydraulic servo  42  of the clutch C- 2  or the hydraulic servo  45  of the brake B- 2 . 
     An oil path a 1 , an oil path a 4 , and an oil path a 5  shown in  FIG. 5  are structured such that the forward range pressure output port (not shown) of the manual shift valve is connected thereto and the forward range pressure P D  can be input thereto. Also, an oil path  1  is structured such that the reverse range pressure output port (not shown) of the manual shift valve is connected thereto and the reverse range pressure P REV  can be input thereto. Further, the line pressure P L  from the primary regulator valve (not shown) is input to an oil path d, and the modulator pressure P MOD  from the modulator valve (not shown) is input to an oil path g 1 . 
     The oil path a 1  is connected to an input port  21   e  of the first clutch apply relay valve  21  described later in detail via an oil path a 2 , and is arranged with a check valve  50  and an orifice  60 . The oil path a 1  is connected to an accumulator  30  via an oil path a 3 , and is connected to the linear solenoid valve SLC 1 . The accumulator  30  has a case  30   c , a piston  30   b  arranged inside the case  30   c , a spring  30   s  which biases the piston  30   b , and an oil chamber  30   a  formed between the case  30   c  and the piston  30   b.    
     The linear solenoid valve (first solenoid valve as a regulated pressure supply portion) SLC 1  is a normally-closed type which is in the non-output state in the non-energized state, and has an input port SLC 1   a  which inputs the forward range pressure P D  via the oil path a 1 , and an output port SLC 1   b  which regulates the forward range pressure P D  and outputs a control pressure P SLC1  as an engagement pressure P C1  to the hydraulic servo  41 . That is, the linear solenoid valve SLB 1  is structured to disconnect the input port SLB 1   a  and the output port SLB 1   b  in the non-energized state to be in the non-output state, and to be capable of increasing the amount of communication (opening amount) between the input port SLC 1   a  and the output port SLC 1   b  in accordance with the command value, i.e., outputting the engagement pressure P C1  in accordance with the command value, in the energized state based on the command value from a hydraulic pressure command unit  71  (see  FIG. 1 ) of a control unit (ECU)  70  described later in detail. The output port SLC 1   b  of the linear solenoid valve SLC 1  is connected to an input port  22   c  of the second clutch apply relay valve  22  described later via an oil path b 1 . 
     The linear solenoid valve (second solenoid valve as a regulated pressure supply portion) SLC 2  is a normally-opened type which is in the output state in the non-energized state, and has an input port SLC 2   a  which inputs the forward range pressure P D  via the oil path a 4  and the like, and an output port SLC 2   b  which regulates the forward range pressure P D  and outputs a control pressure P SLC2  as an engagement pressure P C2  (or an engagement pressure P B2 ) to the hydraulic servo  42 . That is, the linear solenoid valve SLC 2  is structured to be in the output state in which the input port SLC 2   a  and the output port SLC 2   b  are communicated in the non-energized state, and to be capable of decreasing the amount of communication (i.e., reducing the opening amount) between the input port SLC 2   a  and the output port SLC 2   b  in accordance with the command value, i.e., outputting the engagement pressure P C2 , (or P B2 ) in accordance with the command value, in the energized state based on the command value from the hydraulic pressure command unit  71  of the control unit (ECU)  70  described later in detail. The output port SLC 2   b  of the linear solenoid valve SLC 2  is connected to an input port  22   f  of the second clutch apply relay valve  22  described later via an oil path c 1 . 
     The linear solenoid valve (third solenoid valve as a regulated pressure supply portion) SLC 3  is a normally-opened type which is in the output state in the non-energized state, and has an input port SLC 3   a  which inputs the line pressure P L  via the oil path d and the like, and an output port SLC 3   b  which regulates the line pressure P L , and outputs a control pressure P SLC3  as an engagement pressure P C3  to the hydraulic servo  43 . That is, the linear solenoid valve SLC 3  is structured to be in the output state in which the input port SLC 3   a  and the output port SLC 3   b  are communicated in the non-energized state, and to be capable of decreasing the amount of communication (i.e., reducing the opening amount) between the input port SLC 3   a  and the output port SLC 3   b  in accordance with the command value, i.e., outputting the engagement pressure P C3  in accordance with the command value, in the energized state based on the command value from the hydraulic pressure command unit  71  of the control unit (ECU)  70  described later in detail. The output port SLC 3   b  of the linear solenoid valve SLC 3  is connected to the hydraulic servo  43  of the clutch C- 3  via an oil path e 1 . The oil path e 1  is arranged with a check valve  53  and an orifice  63 , and is connected with an oil chamber  33   a  of a C- 3  damper  33  via an oil path e 2 . Note that, since the C- 3  damper  33  has a similar configuration as the accumulator  30  and is a general damper device, a detailed description thereof will be omitted. 
     The linear solenoid valve (fourth solenoid valve as a regulated pressure supply portion) SLB 1  is a normally-closed type which is in the non-output state in the non-energized state, and has an input port SLB 1   a  which inputs the forward range pressure P D  via the oil path a 5  and the like, and an output port SLB 1   b  which regulates the forward range pressure P D  and outputs a control pressure P SLB1  as an engagement pressure P B1  to the hydraulic servo  44 . That is, the linear solenoid valve SLB 1  is structured to disconnect the input port SLB 1   a  and the output port SLB 1   b  in the non-energized state to be in the non-output state, and to be capable of increasing the amount of communication (opening amount) between the input port SLB 1   a  and the output port SLB 1   b  in accordance with the command value, i.e., outputting the engagement pressure P B1  in accordance with the command value, in the energized state based on the command value from the hydraulic pressure command unit  71  of the control unit (ECU)  70  described later in detail. The output port SLB 1   b  of the linear solenoid valve SLB 1  is connected to the hydraulic servo  44  of the brake B- 1  via an oil path f 1 . The oil path f 1  is arranged with a check valve  54  and an orifice  64 , and is connected with an oil chamber  34   a  of a B- 1  damper  34  via an oil path f 2 . 
     The solenoid valve S 11  is a normally-opened type which is in the output state in the non-energized state, and has an input port S 1   a  which inputs the modulator pressure P MOD  via oil paths g 1  and g 2 , and an output port S 1   b  which outputs the modulator pressure P MOD  in the non-energized state (i.e., off-state) almost directly as a signal pressure P S1 . The output port S 1   b  is connected to an oil chamber  21   a  of the first clutch apply relay valve  21  via oil paths h 1  and h 2 . Further, the output port S 1   b  is connected to an oil chamber  22   a  of the second clutch apply relay valve  22  via the oil path h 1  and an oil path h 3 , and is connected to an input port  24   c  of the B- 2  relay valve  24  via an oil path h 4 . 
     The solenoid valve S 12  is a normally-closed type which is in the non-output state in the non-energized state, and has an input port S 2   a  which inputs the modulator pressure P MOD  via oil paths g 1  and g 3 , and an output port S 2   b  which outputs the modulator pressure P MOD  in the energized state (i.e., on-state) almost directly as a signal pressure P S2 . The output port S 2   b  is connected to an oil chamber  24   a  of the B- 2  relay valve via an oil path i. 
     The first clutch apply relay valve  21  is structured to have two spools  21   p  and  21   q , a spring  21   s  which biases the spool  21   p  upward in the drawing, a spring  21   t  which biases the spools  21   p  and  21   q  in directions to depart from each other, the oil chamber  21   a  above the spool  21   q  in the drawing, an oil chamber  21   d  below the spool  21   p  in the drawing, an oil chamber  21   c  between the two spools  21   p  and  21   q , an oil chamber  21   b  formed by a difference in diameter (difference in pressure receiving area) of a land portion of the spool  21   q , the input port  21   e , an input port  21   f , an input port  21   g , an input port  21   h , an output port  21   i , an output port  21   j , and a drain port EX. 
     The first clutch apply relay valve  21  is structured such that the input port  21   e  and the output port  21   j  are communicated and the input port  21   e  and the output port  21   i  are disconnected when the spools  21   p  and  21   q  are in the left half positions, and the input port  21   e  and the output port  21   i  are communicated and the output port  21   j  and the drain port EX are communicated when the spools  21   p  and  21   q  are in the right half positions. The input port  21   h  is disconnected when the spool  21   p  is in the left half position, and the input port  21   g  is disconnected when the spool  21   q  is in the right half position. 
     As described above, the oil chamber  21   a  is connected to the output port S 1   b  of the solenoid valve S 11  via the oil paths h 1  and h 2 , and the oil chamber  21   b  is connected to an output port  22   i  of the second clutch apply relay valve  22  described later via an oil path b 4  from the input port  21   f . The forward range pressure P D  is input to the input port  21   e  via the oil paths a 1  and a 2 . The output port  21   j  which communicates with the input port  21   e  when the spool  21   p  is in the left half position is connected to an input port  22   h  of the second clutch apply relay valve  22  via an oil path j. The output port  21   i  which communicates with the input port  21   e  when the spool  21   p  is in the right half position is connected to the input port  21   g  via oil paths k 1  and k 2  and to the input port  21   h  via the oil paths k 1 , k 2  and an oil path k 3 . That is, the output port  21   i  is connected to the oil chamber  21   c  regardless of the positions of the spools  21   p  and  21   q . Further, the output port  21   i  is connected to an input port  22   e  of the second clutch apply relay valve  22  described later via the oil path k 1 . The oil chamber  21   d  is connected with an output port  23   c  of the C- 2  relay valve  23  via an oil path c 5 , and the oil path c 5  is provided with a check valve  55  and an orifice  65 . 
     The second clutch apply relay valve  22  is structured to have a spool  22   p , a spring  22   s  which biases the spool  22   p  upward in the drawing, the oil chamber  22   a  above the spool  22   p  in the drawing, an oil chamber  22   b  below the spool  22   p  in the drawing, the input port  22   c , an output port  22   d , the input port  22   e , the input port  22   f , an output port  22   g , the input port  22   h , and the output port  22   i.    
     The second clutch apply relay valve  22  is structured such that the input port  22   c  and the output port  22   d  as well as the output port  22   i  are communicated, the input port  22   f  and the output port  22   g  are communicated, and the input port  22   e  and the input port  22   h  are disconnected when the spool  22   p  is in the left half position, and the input port  22   e  and the output port  22   d  are communicated, the input port  22   h  and the output port  22   g  are communicated, and the input port  22   c , the output port  22   i , and the input port  22   f  are disconnected when the spool  22   p  is in the right half position. 
     As described above, the oil chamber  22   a  is connected to the output port S 1   b  of the solenoid valve S 11  via the oil paths h 1  and h 3 , and is connected to the input port  24   c  of the B- 2  relay valve  24  described later via the oil path h 4 . The input port  22   c  is connected to the output port SLC 1   b  of the linear solenoid valve SLC 1  via the oil path b 1 , and the output port  22   d  which connects with the input port  22   c  when the spool  22   p  is in the left half position is connected to the hydraulic servo  41  of the clutch C- 1  via an oil path b 2 . The oil path b 2  is arranged with a check valve  51  and an orifice  61 , and is connected with an oil chamber  31   a  of a C- 1  damper  31  via an oil path b 3 . In a similar manner, the output port  22   i  which communicates with the input port  22   c  when the spool  22   p  is in the left half position is connected to the input port  21   f  of the first clutch apply relay valve  21  via the oil path b 4 , and is connected to the oil chamber  22   b  via oil paths b 4  and b 5 . The input port  22   f  is connected to the output port SLC 2   b  of the linear solenoid valve SLC 2  via the oil path c 1 , and the input port  22   h  is connected to the output port  21   j  of the first clutch apply relay valve  21  via the oil path j. The output port  22   g  which communicates with the input port  22   f  when the spool  22   p  is in the left half position and communicates with the input port  22   h  when the spool  22   p  is in the right half position is connected to an input port  23   b  of the C- 2  relay valve  23  described later via an oil path c 2 . The oil path c 2  is provided with a check valve  52  and an orifice  62 , and is connected with an oil chamber  32   a  of a C 2 -B 2  damper  32  via an oil path c 4 . 
     The C- 2  relay valve  23  is structured to have a spool  23   p , a spring  23   s  which biases the spool  23   p  upward in the drawing, an oil chamber  23   a  above the spool  23   p  in the drawing, the input port  23   b , the output port  23   c , an output port  23   d , an output port  23   e , and a drain port EX. 
     The C- 2  relay valve  23  is structured such that the input port  23   b  and the output port  23   c  as well as the output port  23   e  are communicated and the output port  23   d  and the drain port EX are communicated when the spool  23   p  is in the left half position, and the input port  23   b  and the output port  23   d  are communicated and the output port  23   c  as well as the output port  23   e  and the drain port EX are communicated when the spool  23   p  is in the right half position. 
     The oil chamber  23   a  is connected to an output port  24   b  of the B- 2  relay valve  24  described later via an oil path h 5 . The input port  23   b  is connected to the output port  22   g  of the second clutch apply relay valve  22  via the oil path c 2 , and the output port  23   e  which communicates with the input port  23   b  when the spool  23   p  is in the left half position is connected to the hydraulic servo  42  of the clutch C- 2  via an oil path c 3 . In a similar manner, the output port  23   c  which communicates with the input port  23   b  when the spool  23   p  is in the left half position is connected to the oil chamber  21   d  of the first clutch apply relay valve  21  via the oil path c 5 , and the oil path c 5  is provided with the check valve  55  and the orifice  65 . The output port  23   d  which communicates with the input port  23   b  when the spool  23   p  is in the right half position is connected to an input port  24   e  of the B- 2  relay valve  24  via an oil path m. 
     The B- 2  relay valve  24  is structured to have a spool  24   p , a spring  24   s  which biases the spool  24   p  upward in the drawing, the oil chamber  24   a  above the spool  24   p  in the drawing, the output port  24   b , the input port  24   c , an input port  24   d , the input port  24   e , an output port  24   f , an output port  24   g , and a drain port EX. 
     The B- 2  relay valve  24  is structured such that the input port  24   d  and the output port  24   f  as well as the output port  24   g  are communicated, the output port  24   b  and the drain port EX are communicated, and the input port  24   c  is disconnected when the spool  24   p  is in the left half position, and the input port  24   c  and the output port  24   b  are communicated, the input port  24   e  and the output port  24   g  are connected, and the input port  24   d  and the drain port EX are disconnected when the spool  24   p  is in the right half position. 
     The oil chamber  24   a  is connected to the output port S 2   b  of the solenoid valve S 12  via the oil path i. The input port  24   d  is connected to the reverse range pressure output port (not shown) of the manual shift valve from which the reverse range pressure P REV  is output via the oil path  1 , and the input port  24   e  is connected to the output port  23   d  of the C- 2  relay valve  23  via the oil path m. The output port  24   g  which communicates with the input port  24   d  when the spool  24   p  is in the left half position and communicates with the input port  24   e  when the spool  24   p  is in the right half position is connected to the hydraulic servo  45  of the brake B- 2  via an oil path n. That is, the hydraulic servo  45  of the brake B- 2  is connected to the reverse range pressure output port (not shown) of the manual shift valve or the output port SLC 2   b  of the linear solenoid valve SLC 2 . As described above, the input port  24   c  is connected to the output port S 1   b  of the solenoid valve S 11  via the oil path h 4 , the oil chamber  22   a  of the second clutch apply relay valve  22 , and the oil paths h 1  and h 3 , and the output port  24   b  which communicates with the input port  24   c  when the spool  24   p  is in the right half position is connected to the oil chamber  23   a  of the C- 2  relay valve  23  via the oil path h 5 . Note that the output port  24   f  which communicates with the input port  24   d  when the spool  24   p  is in the left half position is connected to an oil chamber of the primary regulator valve via an oil path (not shown), and is structured such that the reverse range pressure P REV  is applied to the primary regulator valve so as to increase the line pressure P L  when moving in reverse. 
     [Operation of Hydraulic Control Device] 
     Next, the application of the hydraulic control device  6  according to this embodiment will be described. 
     For example, when a driver turns on the ignition, the hydraulic control of the hydraulic control device  6  is started. First, for example, when the selected position of the shift lever is in the P range or the N range, the normally-opened type linear solenoid valve SLC 2 , linear solenoid valve SLC 3 , and solenoid valve S 11  are energized by an electric command of the hydraulic pressure command unit  71  of the control unit  70 , and the respective input ports and the output ports are disconnected. Next, for example, when the engine is started, a hydraulic pressure is generated by the rotation of the oil pump (not shown) based on the engine rotation, and the hydraulic pressure is regulated and output as the line pressure P L  or the modulator pressure P MOD  by the primary regulator valve or the solenoid modulator valve as described above. The line pressure P L  is input to the input port of the manual shift valve (not shown) and to the input port SLC 3   a  of the linear solenoid valve SLC 3  via the oil path d, and the modulator pressure P MOD  is input to the input ports S 1   a  and S 2   a  of the solenoid valves S 11  and S 12  via the oil paths g 1 , g 2 , and g 3 . 
     [Operation in N-D State (First Forward Speed)] 
     Next, for example, when the driver brings the shift lever from the N range position to the D range position, the forward range pressure P D  is output from the forward range pressure output port of the manual shift valve to the oil paths a 1 , a 4 , and a 5 . The forward range pressure P D  is input to the linear solenoid valve SLC 1  via the oil path a 1 , to the linear solenoid valve SLC 2  via the oil path a 4 , to the linear solenoid valve SLB 1  via the oil path a 5 , and to the first clutch apply relay valve  21  via the oil paths a 1  and a 2 . 
     The oil path a 1  is provided with the check valve  50  and the orifice  60 . Since the forward range pressure PD opens the check valve  50 , the supply of the forward range pressure P D  to the linear solenoid valve SLC 1  is more rapid compared to the time of discharge. The forward range pressure P D  supplied to the oil path a 1  is input to the oil chamber  30   a  of the accumulator  30  via the oil path a 3 , and a pressure accumulation of the forward range pressure P D  supplied to the linear solenoid valve SLC 1  is performed by the accumulator  30 . 
     The first clutch apply relay valve  21  in which the forward range pressure P D  is input to the input port  21   e  from the oil path a 2  is in the left half position due to the biasing force of the spring  21   s  immediately after the switch to the D range (immediately after an N-D shift) since the solenoid valve S 11  is turned on and the signal pressure P S1  is not output. Thus, the forward range pressure P D  is output to the oil path j from the output port  21   j . In a similar manner, since the solenoid valve S 11  is turned on and the signal pressure P S1  is not output, the second clutch apply relay valve  22  is in the left half position due to the biasing force of the spring  22   s , and the input port  22   h  is disconnected. 
     Next, for example, when the first forward speed is confirmed by a shift determination unit  75  (see  FIG. 1 ) of the control unit  70  described later, the electric control of the control unit  70  causes the linear solenoid valve SLC 1  to be turned on, the forward range pressure P D  input to the input port SLC 1   a  to be subjected to pressure regulation control, the control pressure P SLC1  to be output as the engagement pressure P C1  from the output port SLC 1   b  to gradually increase, and the control pressure P SLC1  (engagement pressure P C1 ) to be input to the input port  22   c  of the second clutch apply relay valve  22  via the oil path b 1 . 
     The second clutch apply relay valve  22  in the left half position outputs the control pressure P SLC1 , which is input to the input port  22   c , from the output port  22   i  and also from the output port  22   d . The control pressure P SLC1  output from the output port  22   i  is input to the oil chamber  22   b  via the oil paths b 4  and b 5  to lock the second clutch apply relay valve  22  in the left half position, and is input to the oil chamber  21   b  of the first clutch apply relay valve  21  via the oil path b 4  to press the spools  21   p  and  21   q  downward in the drawing against the biasing force of the spring  21   s  to switch the first clutch apply relay valve  21  to the right half position. 
     In the first clutch apply relay valve  21  in which the spools  21   p  and  21   q  are switched to the right half positions, the spool  21   q  is pressed downward in the drawing against the biasing force of the spring  21   t  by the control pressure P SLC1  output from the output port  22   i  of the second clutch apply relay valve  22 . However, since the forward range pressure P D  input through the input port  21   e  is output from the output port  21   i  and input to the oil chamber  21   c  via the oil paths k 1 , k 2 , and k 3  and the input port  21   h , the spool  21   q  is switched to the upper side in the drawing by the hydraulic pressure applied to the oil chamber  21   c  and by the biasing force of the spring  21   t . That is, the spool  21   p  and the spool  21   q  are locked in a distant state. Note that the forward range pressure P D  input from the oil path k 1  to the input port  22   e  of the second clutch apply relay valve  22  is blocked in the input port  22   e.    
     The control pressure P SLC1  input to the input port  22   c  of the second clutch apply relay valve  22  from the linear solenoid valve SLC 1  as described above is output as the engagement pressure P C1  to the hydraulic servo  41  via the oil path b 2  from the output port  22   d  to engage the clutch C- 1 . Accordingly, by a combination with the lock of the one-way clutch F- 1 , the first forward speed is achieved. 
     The oil path b 2  is provided with the check valve  51  and the orifice  61 . The check valve  51  is closed when the engagement pressure P C1  (control pressure P SLC1 ) is supplied to the hydraulic servo  41  such that the hydraulic pressure is supplied moderately via only the orifice  61 , and the engagement pressure P C1  is discharged from the hydraulic servo  41  more rapidly compared to a case where the supply is made with the check valve  51  opened. The engagement pressure P C1  supplied to the oil path b 2  is input to the oil chamber  31   a  of the C- 1  damper  31  via the oil path b 3 , and the pulsation of the engagement pressure P C1  supplied and discharged to and from the hydraulic servo  41  is prevented, and suction of a surge pressure (rapid changing pressure) is performed by the C- 1  damper  31 . 
     [Operation of Engine Braking in First forward Speed] 
     For example, when an engine braking in the first forward speed is confirmed by the shift determination unit  75  of the control unit  70 , the solenoid valve S 12  is turned on, the solenoid valve S 11  is turned off, and the linear solenoid valve SLC 2  is subjected to pressure regulation control by the electric command from the control unit  70 . When the solenoid valve S 12  is turned on, the modulator pressure P MOD  input to the input port S 2   a  via the oil paths g 1  and g 3  is output from the output port S 2   b  as the signal pressure P S2  and input to the oil chamber  24   a  of the B- 2  relay valve  24  via the oil path i, and the spool  24   p  is switched to the lower side in the drawing against the biasing force of the spring  24   s  such that the B- 2  relay valve  24  is brought to the right half position. 
     When the solenoid valve S 11  is turned off, the modulator pressure P MOD  input to the input port S 1   a  via the oil paths g 1  and g 2  is output from the output port S 1   b  as the signal pressure P S1 , and input to the oil chamber  21   a  of the first clutch apply relay valve  21  via the oil paths h 1  and h 2 , to the oil chamber  22   a  of the second clutch apply relay valve  22  via the oil paths h 1  and h 3 , to the input port  24   c  of the B- 2  relay valve  24  via the oil path h 4 , and to the oil chamber  23   a  of the C- 2  relay valve  23  via the oil path h 5  from the output port  24   b  of the B- 2  relay valve  24  in the right half position. 
     In the C- 2  relay valve  23 , the spool  23   p  is switched to the lower side in the drawing against the biasing force of the spring  23   s  by the signal pressure P S1  input to the oil chamber  23   a  to be in the right half position. Note that, in the first clutch apply relay valve  21 , the spool  21   q  is switched to the lower side in the drawing to be in the right half position since the signal pressure P S1  is input to the oil chamber  21   a , but the spool  21   p  is in the same right half position as that in the first forward speed and is not particularly influenced. In the second clutch apply relay valve  22 , although the signal pressure P S1  is input to the oil chamber  22   a , the spool  22   p  stays locked in the left half position since the engagement pressure P C1  of the oil chamber  22   b  and the biasing force of the spring  22   s  are stronger. 
     When the linear solenoid valve SLC 2  is subjected to pressure regulation control and the control pressure P SLC2  is output from the output port SLC 2   b , the control pressure P SLC2  is input to the input port  22   f  of the second clutch apply relay valve  22  locked in the left half position via the oil path c 1 , and is output to the oil path c 2  from the output port  22   g  as the engagement pressure P B2 . 
     The engagement pressure P B2  output to the oil path c 2  is input to the input port  23   b  of the C- 2  relay valve  23  in the right half position, and is output from the output port  23   d . Further, the engagement pressure P B2  is input to the input port  24   e  of the B- 2  relay valve  24  in the right half position via the oil path m, output from the output port  24   g , and input to the hydraulic servo  45  via the oil path n to lock the brake B- 2 . Accordingly, in combination with the engagement of the clutch C- 1 , the engine braking in the first forward speed is achieved. 
     Note that the oil path c 2  is provided with the check valve  52  and the orifice  62 . The check valve  52  is closed when the engagement pressure P B2  is supplied to the hydraulic servo  45  of the brake B- 2  such that the hydraulic pressure is supplied moderately via only the orifice  62 , and the check valve  52  is opened to discharge the hydraulic pressure in the oil path c 2  rapidly at the time of discharge described later. Further, the engagement pressure P B2  supplied to the oil path c 2  is input to the oil chamber  32   a  of the C 2 -B 2  damper  32  via the oil path c 4 , and the pulsation of the engagement pressure P B2  supplied and discharged to and from the hydraulic servo  45  is prevented, and suction of a surge pressure (rapid changing pressure) is performed by the C 2 -B 2  damper  32 . 
     For example, when a positive drive in the first forward speed is confirmed, i.e., the release of the engine braking state is confirmed, by the shift determination unit  75  of the control unit  70 , the solenoid valve S 12  is turned off, the solenoid valve S 11  is turned on, and the linear solenoid valve SLC 2  is turned on (energized) and closed such that the control pressure P SLC2  as the engagement pressure P B2  is reduced to zero and drained. Since the B- 2  relay valve  24  is switched to the left half position by the solenoid valve S 12  being turned off, the engagement pressure P B2  of the hydraulic servo  45  of the brake B- 2  is discharged from the drain port of the manual shift valve via the input port  24   d , the oil path  1 , and the reverse range pressure output port (not shown) of the manual shift valve. Accordingly, a quick drain quicker than the drainage via the linear solenoid valve SLC 2  is performed to quickly release the brake B- 2 . Note that the hydraulic pressure in the oil path m is discharged from the drain port EX of the C- 2  relay valve  23  switched to the left half position, and the hydraulic pressures in the oil paths c 1  and c 2  are discharged from the drain port EX of the linear solenoid valve SLC 2 . 
     [Operation in Second Forward Speed] 
     Next, for example, when the second forward speed is confirmed by the shift determination unit  75  of the control unit  70  from the state of the first forward speed, the linear solenoid valve SLB 1  is subjected to pressure regulation control while the pressure regulated state of the linear solenoid valve SLC 1  is maintained in a state where the solenoid valve S 11  is turned on and the solenoid valve S 12  is turned off by the electric command from the control unit  70  in a similar manner as in the first forward speed (excluding the time of the engine braking). 
     That is, when the linear solenoid valve SLB 1  is subjected to the pressure regulation control, the control pressure P SLB1  is output from the output port SLB 1   b  as the engagement pressure P B1 , and input to the hydraulic servo  44  via the oil path f 1  to lock the brake B- 1 . Accordingly, in combination with the engagement of the clutch C- 1 , the second forward speed is achieved. 
     The oil path f 1  is provided with the check valve  54  and the orifice  64 . The check valve  54  is closed when the engagement pressure P B1  is supplied to the hydraulic servo  44  of the brake B- 1  such that the hydraulic pressure is supplied moderately via only the orifice  64 , and the engagement pressure P B1  is discharged from the hydraulic servo  44  more rapidly compared to a case where the supply is made with the check valve  54  opened. Further, the engagement pressure P B1  supplied to the oil path f 1  is input to the oil chamber  34   a  of the B- 1  damper  34  via the oil path f 2 , and the pulsation of the engagement pressure P B1  supplied and discharged to and from the hydraulic servo  44  is prevented, and suction of a surge pressure (rapid changing pressure) is performed by the B- 1  damper  34 . 
     [Operation in Third Forward Speed] 
     Next, for example, when the third forward speed is confirmed by the shift determination unit  75  of the control unit  70  from the state of the second forward speed, the linear solenoid valve SLB 1  is turned off to be closed and the pressure regulation control of the linear solenoid valve SLC 3  is performed while the pressure regulated state of the linear solenoid valve SLC 1  is maintained in the state where the solenoid valve S 11  is turned on and the solenoid valve S 12  is turned off in a similar manner by the electric command from the control unit  70 . 
     That is, the release control of the brake B- 1  is performed by the pressure regulation control of the linear solenoid valve SLB 1 , i.e., the engagement pressure P B1  (control pressure P SLB1 ) of the hydraulic servo  44  of the brake B- 1  is subjected to discharge control by the drain port EX of the linear solenoid valve SLB 1  via the oil path f 1 , to release the brake B- 1 . The pressure regulation control is performed from the closed state where the linear solenoid valve SLC 3  is turned on (energized) to bring the control pressure P SLC3  to zero. The control pressure P SLC3  is output from the output port SLC 3   b  as the engagement pressure P C3  and input to the hydraulic servo  43  via the oil path e 1  to engage the clutch C- 3 . Accordingly, in combination with the engagement of the clutch C- 1 , the third forward speed is achieved. 
     The oil path e 1  is provided with the check valve  53  and the orifice  63 . The check valve  53  is closed when the engagement pressure P C3  is supplied to the hydraulic servo  43  of the clutch C- 3  such that the hydraulic pressure is supplied moderately via only the orifice  63 , and the engagement pressure P C3  is discharged from the hydraulic servo  43  more rapidly compared to a case where the supply is made with the check valve  53  opened. Further, the engagement pressure P C3  supplied to the oil path e 1  is input to the oil chamber  33   a  of the C- 3  damper  33  via the oil path e 2 , and the pulsation of the engagement pressure P C3  supplied and discharged to and from the hydraulic servo  43  is prevented, and suction of a surge pressure (rapid changing pressure) is performed by the C- 3  damper  33 . 
     [Operation in Fourth Forward Speed] 
     Next, for example, when the fourth forward speed is confirmed by the shift determination unit  75  of the control unit  70  from the state of the third forward speed, the linear solenoid valve SLC 3  is turned off to be closed and the pressure regulation control of the linear solenoid valve SLC 2  is performed while the pressure regulated state of the linear solenoid valve SLC 1  is maintained in the state where the solenoid valve S 11  is turned on and the solenoid valve S 12  is turned off in a similar manner by the electric command from the control unit  70 . 
     That is, the release control of the clutch C- 3  is performed by the pressure regulation control of the linear solenoid valve SLC 3 , i.e., the engagement pressure P C3  (control pressure P SLC3 ) of the hydraulic servo  43  of the clutch C- 3  is subjected to discharge control by the drain port EX of the linear solenoid valve SLC 3  via the oil path e 1 , to release the clutch C- 3 . The pressure regulation control is performed from the closed state where the linear solenoid valve SLC 2  is turned on (energized) to bring the control pressure P SLC2  to zero. The control pressure P SLC2  is output from the output port SLC 2   b  as the engagement pressure P C2  and input to the input port  22   f  of the second clutch apply relay valve  22  via the oil path c 1 . 
     Since the signal pressure P S1  is not input to the oil chamber  22   a  due to the solenoid valve S 11  being turned on and the second clutch apply relay valve  22  is locked to the left half position by the engagement pressure P C1  input to the oil chamber  22   b  as described above, the control pressure P SLC2  (engagement pressure P C2 ) input to the input port  22   f  is output as the engagement pressure P C2  from the output port  22   g . The engagement pressure P C2  output from the output port  22   g  is input to the input port  23   b  of the C- 2  relay valve  23  via the oil path c 2 . 
     Further, since the solenoid valve S 12  is turned off, the B- 2  relay valve  24  is in the left half position, the oil chamber  23   a  and the oil path h 5  are in drain states, and the C- 2  relay valve  23  is in the left half position by the biasing force of the spring  23   s , the engagement pressure P C2  input to the input port  23   b  is output from the output port  23   c  and also output from the output port  23   e . The engagement pressure P C2  output from the output port  23   c  is input to the oil chamber  21   d  of the first clutch apply relay valve  21  via the oil path c 5 . The engagement pressure P C2  combined with the biasing force of the spring  21   s  causes the spool  21   p  of the first clutch apply relay valve  21  to be switched and locked to the left half position. At this time, the forward range pressure P D  input to the input port  22   e  via the oil path k 1  is switched to the output port  21   j  from the output port  21   i  and output to the oil path j, but is blocked by the input port  22   h  of the second clutch apply relay valve  22 . Since the forward range pressure P D  supplied to the oil path k 1  is blocked, the supply of the forward range pressure P D  as a lock pressure with respect to the oil chamber  21   c  via the oil paths k 2  and k 3  is released. 
     Note that the oil path c 5  is provided with the check valve  55  and the orifice  65 . The check valve  55  is closed when the engagement pressure P C2  is supplied to the oil chamber  21   d  of the first clutch apply relay valve  21  such that the hydraulic pressure is supplied moderately via only the orifice  65 , and the engagement pressure P C2  is discharged from the oil chamber  21   d  more rapidly compared to a case where the supply is made with the check valve  55  opened. 
     The engagement pressure P C2  output from the output port  23   e  of the C- 2  relay valve  23  is input to the hydraulic servo  42  via the oil path c 3  to engage the clutch C- 2 . Accordingly, in combination with the engagement of the clutch C- 1 , the fourth forward speed is achieved. 
     As described above, the oil path c 2  is provided with the check valve  52  and the orifice  62 . In a similar manner as the engine braking in the first forward speed, the check valve  52  is closed when the engagement pressure P C2  is supplied to the hydraulic servo  42  of the clutch C- 2  such that the hydraulic pressure is supplied moderately via only the orifice  62 , and the engagement pressure P C2  is discharged from the hydraulic servo  42  more rapidly compared to a case where the supply is made with the check valve  52  opened. Further, the engagement pressure P C2  supplied to the oil path c 2  is input to the oil chamber  32   a  of the C 2 -B 2  damper  32  via the oil path c 4 , and the pulsation of the engagement pressure P C2  supplied/discharged with respect to the hydraulic servo  42  is prevented and suction or the like of a surge pressure (rapid changing pressure) is performed by the C 2 -B 2  damper  32 . 
     [Operation in Fifth Forward Speed] 
     Next, for example, when the fifth forward speed is confirmed by the shift determination unit  75  of the control unit  70  from the state of the fourth forward speed, the linear solenoid valve SLC 1  is turned off to be closed and the pressure regulation control of the linear solenoid valve SLC 3  is performed while the pressure regulated state of the linear solenoid valve SLC 2  is maintained in the state where the solenoid valve S 11  is turned on and the solenoid valve S 12  is turned off in a similar manner by the electric command from the control unit  70 . 
     That is, the release control of the clutch C- 1  is performed by the pressure regulation control of the linear solenoid valve SLC 1 , i.e., the engagement pressure P C1  (control pressure P SLC1 ) of the hydraulic servo  41  of the clutch C- 1  is controlled to be discharged from the drain port EX of the linear solenoid valve SLC 1  via the oil paths b 1  and b 2 , to release the clutch C- 1 . In a similar manner as in the third forward speed, the pressure regulation control is performed from the closed state where the linear solenoid valve SLC 3  is turned on (energized) to bring the control pressure P SLC3  to zero. The control pressure P SLC3  is output from the output port SLC 3   b  as the engagement pressure P C3  and input to the hydraulic servo  43  via the oil path e 1  to engage the clutch C- 3 . Accordingly, in combination with the engagement of the clutch C- 2 , the fifth forward speed is achieved. 
     [Operation in Sixth Forward Speed] 
     For example, when the sixth forward speed is confirmed by the shift determination unit  75  of the control unit  70  from the state of the fifth forward speed, the linear solenoid valve SLC 3  is turned on (energized) to be closed and the pressure regulation control of the linear solenoid valve SLB 1  is performed while the pressure regulated state of the linear solenoid valve SLC 2  is maintained in the state where the solenoid valve S 11  is turned on and the solenoid valve S 12  is turned off in a similar manner by the electric command from the control unit  70 . 
     That is, the release control of the clutch C- 3  is performed by the pressure regulation control of the linear solenoid valve SLC 3 , i.e., the engagement pressure P C3  (control pressure P SLC3 ) of the hydraulic servo  43  of the clutch C- 3  is controlled to be discharged from the drain port EX of the linear solenoid valve SLC 3  via the oil path e 1 , to release the clutch C- 3 . In a similar manner as in the second forward speed, the linear solenoid valve SLB 1  is turned on (energized) to perform the pressure regulation control from the closed state where the linear solenoid valve SLB 1  is turned off to bring the control pressure P SLB1  to zero. The control pressure P SLB1  is output from the output port SLB 1   b  as the engagement pressure P B1  and input to the hydraulic servo  44  via the oil path f 1  to engage the brake B- 1 . Accordingly, in combination with the engagement of the clutch C- 2 , the sixth forward speed is achieved. 
     [Operation in D-N State] 
     Then, for example, when the shift lever is brought to the N range position from the D range position after the driver has decelerated the vehicle to cause a downshift according to the vehicle speed and stopped the vehicle in the first forward speed, the forward range pressure output port of the manual shift valve is disconnected from the input port and communicated with the drain port. That is, the forward range pressure P D  is drained. 
     Simultaneously, when a shift lever sensor (not shown) detects that the shift lever is in the N range position and the control unit  70  confirms the N range based on the shift lever position, the linear solenoid valve SLC 2  and the linear solenoid valve SLC 3  are turned on (energized), and the linear solenoid valve SLB 1  is turned off. The control pressures P SLC2 , P SLC3 , and P SLB1  are drained to zero pressures (non-output states), i.e., the hydraulic pressures of the respective hydraulic servos  42 ,  43 ,  44 , and  45  are drained, to release the clutch C- 2 , the clutch C- 3 , the brake B- 1 , and the brake B- 2 . Note that the solenoid valve S 11  is maintained in the on-state (energized state) and the solenoid valve S 12  is maintained in the off-state. That is, the signal pressures P S1 , and P S2  are not output from the two solenoid valves S 11  and S 12 . 
     For example, since a release shock occurs when the clutch C- 1  is suddenly released, the linear solenoid valve SLC 1  performs the pressure regulation control to gradually reduce the control pressure P SLC1  and finally drains the control pressure P SLC1  to zero pressure (non-output state) to moderately release the clutch C- 1 . When the clutch C- 1  is released, all clutches and brakes are released to bring the automatic transmission  3  to a neutral state. 
     During the release control by the linear solenoid valve SLC 1 , since the accumulator  30  connected to the input port SLC 1   a  of the linear solenoid valve SLC 1  via the oil path a 3  and the like releases the hydraulic pressure accumulated while in the D range to perform pressure maintenance with respect to the oil paths a 1  and a 3  located closer to the linear solenoid valve SLC 1  than the orifice  60 , a moderate release control of the clutch C- 1  by the linear solenoid valve SLC 1  is possible. Accordingly, the occurrence of the release shock is prevented in a D-N shift operation from the state of the first forward speed. 
     [Operation in First Reverse Speed] 
     For example, when the shift lever is brought to the R range position by the operation of the shift lever by the driver, the reverse range pressure P REV  is output from the reverse range pressure output port of the manual shift valve as described above, and the reverse range pressure P REV  is input to the input port  24   d  of the B- 2  relay valve  24  via the oil path  1 . 
     Simultaneously, when the shift lever sensor (not shown) detects that the shift lever is in the R range position and the control unit  70  confirms the R range based on the shift lever position, the solenoid valve S 11  is maintained in the on-state (energized state) and the solenoid valve S 12  is maintained in the off-state. That is, the signal pressure P S2  is not output. Therefore, the B- 2  relay valve  24  is maintained in the left half position by the biasing force of the spring  24   s . Accordingly, the reverse range pressure P REV  input to the input port  24   d  is supplied to the hydraulic servo  45  of the brake B- 2  via the output port  24   g  and the oil path n to engage the brake B- 2 . 
     Further, the control unit  70  performs the pressure regulation control such that the control pressure P SLC3  is gradually output by the linear solenoid valve SLC 3  as the engagement pressure P C3  from the output port SLC 3   b  and input to the hydraulic servo  43  via the oil path e 1 . That is, the clutch C- 3  is moderately engaged. Accordingly, in combination with the locking of brake B- 2 , the first reverse speed is achieved. 
     Note that, when switched from the R range to the N range, control is performed in a similar manner to the N range. That is, the engagement pressure P B2  of the hydraulic servo  45  of the brake B- 2  is drained via the oil path n, the B- 2  relay valve  24 , the oil path  1 , and the manual shift valve, and the engagement pressure P C3  of the hydraulic servo  43  of the clutch C- 3  is drained from the linear solenoid valve SLC 3 . 
     For example, when the driver operates the shift lever to the R range position and a vehicle speed of a predetermined speed or greater in the forward direction is detected, a so-called reverse inhibit function is performed in which the control unit  70  causes the solenoid valve S 12  to be turned on and the on-state (energized state) of the linear solenoid valve SLC 3  to be maintained, i.e., the R range pressure P REV  to be blocked by the B- 2  relay valve  24  so as not to be supplied to the hydraulic servo  45  of the brake B- 2  and the engagement pressure P C3  (control pressure P SLC3 ) to be not supplied to the hydraulic servo  43  of the clutch C- 3 , to prevent the achievement of the first reverse speed. 
     [Operation in Solenoid All-Off Failure State] 
     Next, the operation in a solenoid all-off failure state in the hydraulic control device  6  will be described. In the case where all of the solenoid valves (the linear solenoid valve SLC 1 , the linear solenoid valve SLC 2 , the linear solenoid valve SLC 3 , the linear solenoid valve SLB 1 , the solenoid valve S 11 , and the solenoid valve S 12 ) fail (hereinafter called an “all-off failure”) due to, for example, a short circuit or disconnection of a battery during a normal driving with the shift lever position in the D range, the normally-closed type linear solenoid valve SLC 1 , linear solenoid valve SLB 1 , and solenoid valve S 12  do not output hydraulic pressures, and the normally-opened type linear solenoid valve SLC 2 , linear solenoid valve SLC 3 , and solenoid valve S 11  respectively output hydraulic pressures. 
     In the first clutch apply relay valve  21  when driving in the first forward speed to the third forward speed in the normal state, the spool  21   p  is locked to the right half position by the forward range pressure P D  input to the oil chamber  21   c  as described above. Therefore, the forward range pressure P D  output from the output port  21   i  is input to the input port  22   e  of the second clutch apply relay valve  22  via the oil path k 1  and blocked by the second clutch apply relay valve  22  in the left half position. 
     When the all-off failure occurs in this state, the second clutch apply relay valve  22  is switched to the right half position by the signal pressure P S1  output from the solenoid valve S 11  being input to the oil chamber  22   a  via the oil paths h 1  and h 3 , and the forward range pressure P D  input to the input port  22   e  is output from the output port  22   d  and input to the hydraulic servo  41  via the oil path b 2  to engage the clutch C- 1 . The control pressure P SLC2  (engagement pressure P C2 ) output from the normally opened linear solenoid valve SLC 2  is blocked by the input port  22   f  of the second clutch apply relay valve  22  switched to the right half position. Further, in the normally opened linear solenoid valve SLC 3 , the line pressure P L  input to the input port SLC 3   a  is almost directly output as the engagement pressure P C3  from the output port SLC 3   b  and input to the hydraulic servo  43  via the oil path e 1  to engage the clutch C- 3 . Accordingly, the clutch C- 1  and the clutch C- 3  are engaged to achieve the third forward speed (see  FIG. 3 ). That is, when the all-off failure occurs while driving in the first forward speed to the third forward speed, the driving state in the third forward speed is ensured. 
     When driving in the fourth forward speed to the sixth forward speed in the normal state, since the engagement pressure P C2  of the clutch C- 2  is input to the oil chamber  21   d  of the first clutch apply relay valve  21  via the oil path c 1 , the second clutch apply relay valve  22 , the oil path c 2 , the C- 2  relay valve  23 , and the oil path c 5  as described above to lock the spools  21   p  and  21   q  in the left half positions, the forward range pressure P D  output from the output port  21   j  is input to the input port  22   h  of the second clutch apply relay valve  22  via the oil path j and blocked by the second clutch apply relay valve  22  in the left half position. 
     When the all-off failure occurs in this state, since the second clutch apply relay valve  22  is switched to the right half position by the signal pressure P S1  output from the solenoid valve S 11  being input to the oil chamber  22   a  via the oil paths h 1  and h 3 , and the B- 2  relay valve  24  is not switched and maintained in the left half position by the solenoid valve S 12  being turned off, the oil path h 4  is disconnected and the signal pressure P S1  of the solenoid valve S 11  is not output to the oil path h 5 . Thus, the C- 2  relay valve  23  is also not switched and maintained in the left half position. Therefore, the forward range pressure P D  input to the input port  22   h  of the second clutch apply relay valve  22  is output from the output port  22   g  and input to the hydraulic servo  42  via the oil path c 2 , the C- 2  relay valve  23 , and the oil path c 3  to engage the clutch C- 2 . The control pressure P SLC2  (engagement pressure P C2 ) output from the normally opened linear solenoid valve SLC 2  is blocked by the input port  22   f  of the second clutch apply relay valve  22  switched to the right half position, but the forward range pressure P D  output to the oil path c 2  is output also to the oil path c 5  via the C- 2  relay valve  23  and input to the oil chamber  21   d  of the first clutch apply relay valve  21 . Therefore, the first clutch apply relay valve  21  continues to be locked in the left half position. In the normally opened linear solenoid valve SLC 3 , the line pressure P L  input to the input port SLC 3   a  is almost directly output as the engagement pressure P C3  from the output port SLC 3   b  and input to the hydraulic servo  43  via the oil path e 1  to engage the clutch C- 3 . Accordingly, the clutch C- 2  and the clutch C- 3  are engaged to achieve the fifth forward speed (see  FIG. 3 ). That is, when the all-off failure occurs while driving in the fourth forward speed to the sixth forward speed, the driving state in the fifth forward speed is ensured. 
     When the vehicle is stopped and the shift lever is temporarily brought to the N range position in the case where the all-off failure has occurred in the normal driving state in the fourth forward speed to the sixth forward speed, the manual shift valve (not shown) stops the output and drains the forward range pressure P D . Particularly, the forward range pressure P D  for the normally opened linear solenoid valve SLC 2  and the input port  21   e  of the first clutch apply relay valve  21  is drained. Thus, the forward range pressure P D  which has been input to the oil chamber  21   d  via the oil paths j, c 2 , and c 5  is drained to release the lock by the forward range pressure P D . Since the signal pressure P S1  from the normally opened solenoid valve S 11  continues to be output, the spools  21   p  and  21   q  are switched to the right half positions by the signal pressure P S1  input to the oil chamber  21   a  in the first clutch apply relay valve  21 . 
     Note that, since the line pressure P L  is the source pressure in the N range state at the time of the all-off failure and the control pressure P SLC3  (engagement pressure P C3 ) which is approximately the same as the line pressure P L  is output from the normally opened linear solenoid valve SLC 3 , the clutch C- 3  is in the engaged state. Since the clutches C- 1  and C- 2  and the brakes B- 1  and B- 2  are in the released states even though the clutch C- 3  is engaged, and the sun gear S 3  and the carrier CR 2  run idle even if the decelerated rotation is input to the sun gear S 2 , it is approximately in the neutral state between the input shaft  10  and the counter gear  11  (see  FIG. 2 ). 
     For example, when the driver returns the shift lever to the D range position, the forward range pressure P D  is output from the manual shift valve. The forward range pressure P D  is input to the input port  21   e  of the first clutch apply relay valve  21  switched to the right half position, output to the oil path k 1  from the output port  21   i , and input to the hydraulic servo  41  of the clutch C- 1  via the input port  22   e  of the second clutch apply relay valve  22  in the right half position, the output port  22   d , and the oil path b 2  to engage the clutch C- 1 . That is, it is brought to a state similar to the all-off failure state while driving in the first forward speed to the third forward speed to ensure the third forward speed. Accordingly, the vehicle can be restarted even after the all-off failure and after the vehicle is temporarily stopped, whereby the limp-home function is ensured. 
     [Description of the Present Invention] 
     Next, a control device  1  for the automatic transmission according to the present invention will be described mainly with reference to  FIGS. 1 and 6  to  8 B. 
     As shown in  FIG. 1 , the control device  1  for the automatic transmission has the control unit (ECU)  70 . The control unit  70  is connected with an accelerator opening degree sensor  81 , an output shaft rotation speed (vehicle speed) sensor  82 , and the like, and is connected to the respective linear solenoid valves SLC 1 , SLC 2 , SLC 3 , and SLB 1 , and the solenoid valves S 11  and S 12 , for example, of the hydraulic control device  6 . The control unit  70  includes the hydraulic pressure command unit  71  having a normal state hydraulic pressure setting unit  72 , an input torque detection unit  73 , a torque distribution determination unit  74 , the shift determination unit  75 , and a shift map map. 
     The shift determination unit  75  references the shift map map based on the accelerator opening degree detected by the accelerator opening degree sensor  81  and the vehicle speed detected by the output shaft rotation speed sensor  82  to determine the first forward speed to the sixth forward speed. That is, an upshift line and a downshift line (shift points) corresponding to the accelerator opening degree and the vehicle speed are recorded in the shift map map, and the shift is confirmed by the shift determination unit  75  when the accelerator opening degree and the vehicle speed at that point exceed the shift lines (see steps S 1  and S 2  of  FIG. 6 ). The shift speed determined by the shift determination unit  75  is output to the hydraulic pressure command unit  71  and the torque distribution determination unit  74 . 
     The input torque detection unit  73  measures an engine torque (see step S 3  of  FIG. 6 ) by inputting an engine torque signal from the engine  2 , and detects an input torque currently input to the input shaft  10  of the automatic speed change mechanism  5 . The torque distribution determination unit  74  determines (calculates) the torque distribution (see step S 4  of  FIG. 6 ) of the clutch and brake (see  FIG. 3 ) engaged in the automatic speed change mechanism  5  based on the shift speed determined by the shift determination unit  75 , i.e., the ratio to the input torque necessary for the clutch and brake based on the respective gear ratios (λ 1 , λ 2 , λ 3 , λ 4 , and λ 5  of  FIGS. 8A and 8B  described later). 
     Next, the normal state hydraulic pressure setting unit  72  multiplies the torque distribution of the clutch or brake engaged in accordance with the shift speed determined by the torque distribution determination unit  74  by a safety ratio (for example, 1.1 to 1.3 times set in accordance with the variation and the like of each part) (step S 5  of  FIG. 6 ), further multiplies the value of the torque distribution multiplied by the safety ratio by the input torque detected by the input torque detection unit  73  to calculate the torque capacity (transmission torque) of the engaged clutch or brake, and calculates the engagement pressure (control pressure) supplied to the hydraulic servo of the engaged clutch or brake from the number of friction plates, the area, the pressure receiving area of the hydraulic servo, and the like of each clutch or brake (step S 6  of  FIG. 6 ). 
     Based on the engagement pressure set by the normal state hydraulic pressure setting unit  72 , the hydraulic pressure command unit  71  gives the electric command to the linear solenoid valves SLC 1 , SLC 2 , SLC 3 , and SLB 1  to supply the engagement pressure to the hydraulic servo of the engaged clutch of brake. That is, while driving in the normal state, the clutch or brake is engaged such that the safety ratio is taken into consideration in addition to the input torque for the torque capacity, and such that the clutch or brake is prevented from slipping even if the engine torque of the engine  2  fluctuates or the torque fluctuation is received from the drive wheel due to the road situation or the like in particular. 
     Next, the change in torque distribution which occurs by the simultaneous engagement of three friction engagement elements in the ease where a failure has occurred in a state where one of the linear solenoid valves SLC 1 , SLC 2 , SLC 3 , and SLB 1  supplying the engagement pressure to the hydraulic servo of the released clutch or brake while driving in the normal state outputs the maximum pressure, i.e., in a state where the same pressure as the line pressure P L  is output, will be described as an example taking a case (fifth failure case Fa 5  of  FIG. 7 ) where the clutch C- 3  is engaged in the state of the fourth forward speed. 
     For example, the clutch C- 1  and the clutch C- 2  are engaged while driving in the fourth forward speed in the normal state, as shown in  FIG. 3 , and the state of torque application in the automatic speed change mechanism  5  is as shown in  FIG. 8A . That is, the relational expression of balanced force in the planetary gear unit PU is as follows.
 
 Tout =T   C1   +T   C2    (1)
 
     The relational expression of balanced moment in the planetary gear unit PU is as follows.
 
 T   C2 ·λ3 +T   C1 ·(λ3+λ4+λ5)= Tout ·(λ3+λ4)   (2)
 
     Note that, as shown in  FIGS. 8A and 8B , λ 1  is the gear ratio of the sun gear S 1  and the carrier CR 1 , λ 2  is the gear ratio of the carrier CR 1  and the ring gear R 1 , λ 3  is the gear ratio of the sun gear S 2  and the carrier CR 2 , λ 4  is the gear ratio of the carrier CR 2  and the ring gear R 2 , and λ 5  is the gear ratio of the ring gear R 2  and the sun gear S 3 . 
     For example, when a failure occurs in a state where the linear solenoid valve SLC 3  outputs the control pressure P SLC3  at the line pressure P L , the simultaneous engagement of the clutch C- 1 , the clutch C- 2 , and the clutch C- 3  occurs. At this time, a force which attempts to stall the automatic speed change mechanism  5  is generated, and a force which attempts to rotate the automatic speed change mechanism  5  with the driving force of the engine  2 , and a force which attempts to rotate the automatic speed change mechanism  5  with the grip force of the drive wheel (inertia force of the vehicle) are generated. The state of the torque application when the driving force (i.e., input torque) of the engine  2  is zero and the automatic speed change mechanism  5  is attempted to be rotated only by the grip force from the drive wheel as the worst condition is shown in  FIG. 8B . That is, the relational expression of balanced force is as follows.
 
 Tout+T   C2   =T   C1   +T   C3    (3)
 
     The relational expression of balanced moment in the planetary gear SP is as follows.
 
( T   C1   +T   C3 )·λ1 =T   C2 ·(λ1+λ2)   (4)
 
     The relational expression of balanced moment in the planetary gear unit PU is as follows.
 
 T   C3 ·λ3= Tout· λ4 −T   C1 ·(λ4+λ5)   (5)
 
     Note that, since the input torque is assumed to be zero, the torque application of the clutch C- 2  is in the opposite direction as a reaction force with respect to the rotational force from the drive wheel. 
     From the expressions (3), (4), and (5), a value in which the torque capacity T C1  of the clutch C- 1  is converted to the output shaft torque Tout, a value in which the torque capacity T C2  of the clutch C- 2  is converted to the output shaft torque Tout, and a value in which the torque capacity T C3  of the clutch C- 3  is converted to the output shaft torque Tout can be obtained respectively. That is, the following is obtained by substitution of the expression (3) into the expression (4).
 
( Tout+T   c2 )·λ1= T   C2 ·(λ1+λ2)
 
Then, the following is obtained.
 
 T   C2 =(λ1/λ2)· Tout    (6)
 
The expression (5) is as follows.
 
 T   C1 ·(λ4+λ5)= Tout ·λ4− T   C3 ·λ3  T   C1 =( Tout ·λ4 −T   C3 ·λ3)/(λ4+λ5)   (5′)
 
The following is obtained by substitution of the expression (5′) and the expression (6) into the expression (3).
 
 Tout+ (λ1/λ2)· Tout =( Tout ·λ4 −T   C3 λ3)/(λ4+λ5)+ T   C3    T   C3 =[(λ1·5+λ2·λ5+λ1·4)/{λ2−(λ4+λ5−λ3)}] ·Tout    (7)
 
The following is obtained by substitution of the expression (6) and the expression (7) into the expression (3).
 
 T   C1   =Tout +(λ1/λ2)· Tout −[(λ1·λ5+λ2·λ5+λ1·λ4)/{λ2·(λ4+λ5−λ3)}]· Tout T   C1 =[1+(λ1/λ2)−(λ1·λ5+λ2·λ5+λ1·λ4)/{λ2·(λ4+λ5−3)}]· Tout    (8)
 
     For example, assuming that the engine torque is at the maximum value in the fourth forward speed (assuming that the input torque is at the maximum value), and that T C1  is the value obtained by multiplying the torque capacity of the clutch C- 1  in the normal state by the safety ratio, T c2  is the value obtained by multiplying the torque capacity of the clutch C- 2  in the normal state by the safety ratio, and T C3  is the value of the torque capacity in the case where the line pressure P L  is supplied to the hydraulic servo  43  of the clutch C- 3 , the values in which the torque capacities of the respective clutches C- 1 , C- 2 , and C- 3  when a failure has occurred in the worst condition are converted to the output shaft torques can be calculated by substituting the gear ratios λ1, λ2, λ3, λ4, and λ5 into the expressions (6), (7), and (8). 
     The calculation result is the fifth failure case Fa 5  shown in  FIG. 7 . That is, even in the worst condition where the linear solenoid valve SLC 3  has failed to engage the clutch C- 3  at the line pressure P L  and the accelerator is released to bring the engine torque (input torque) to zero while driving in the fourth forward speed with the throttle of the engine  2  fully opened, the converted value of the clutch C- 1  with respect to the output shaft torque and the converted value of the clutch C- 2  with respect to the output shaft torque become less than a limit torque Ttire at which the drive wheel slips. 
     Therefore, the clutch C- 1  of which the converted value with respect to the output shaft torque becomes minimum (in other words, the clutch C- 1  in which the torque received from the output shaft (drive wheel) becomes maximum) is caused to slip based on the inertia force of the vehicle received from the drive wheel without the drive wheel slipping. Thus, the clutch C- 2  and the clutch C- 3  are in the engaged states without the three clutches C- 1 , C- 2 , and C- 3  engaging simultaneously, i.e., in the state of the fifth forward speed, to ensure the driving state without being brought to a stalled state. 
     That is, in the fourth forward speed in the normal state, the clutches C- 1  and C- 2  are not engaged at the line pressure P L , but are engaged by the normal state hydraulic pressure setting unit  72  at the engagement pressures P C1  and P C2 , which are hydraulic pressures as low as possible, in consideration of the safety ratio so as not to slip due to the input torque. Thus, based on the change of the torque distribution by the clutch C- 3  engaging at the time of failure, the clutch C- 1  is set to slip even at the time of the failure. Accordingly, the stalled state can be prevented even if the failure occurs. 
     An example of the fifth failure case Fa 5  as a case where the clutch C- 3  is engaged due to a failure in the fourth forward speed has been described above.  FIG. 7  shows states of a first failure case Fa 1  to a fourth failure case Fa 4  and a sixth failure case Fa 6  to a tenth failure case Fa 10  covering all possible failure cases when the torque distribution is calculated in a similar manner. 
     That is, in the first failure case Fa 1  as a case where the clutch C- 2  is engaged due to a failure of the linear solenoid valve SLC 2  in the worst condition in the second forward speed, the converted value of the brake B- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 2  are in the engaged states, i.e., in the state of the fourth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the second failure case Fa 2  as a case where the clutch C- 3  is engaged due to a failure of the linear solenoid valve SLC 3  in the worst condition in the second forward speed, the converted value of the brake B- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 3  are in the engaged states, i.e., in the state of the third forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the third failure case Fa 3  as a case where the clutch C- 2  is engaged due to a failure of the linear solenoid valve SLC 2  in the worst condition in the third forward speed, the converted value of the clutch C- 3  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 2  are in the engaged states, i.e., in the state of the fourth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the fourth failure case Fa 4  as a case where the brake B- 1  is engaged due to a failure of the linear solenoid valve SLB 1  in the worst condition in the third forward speed, the converted value of the brake B- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 3  are in the engaged states, i.e., in the state of the third forward speed, and the driving state is ensured without being brought to the stalled state. Note that, in the fourth failure case Fa 4 , the line pressure P L  is supplied to the hydraulic servo  44  of the brake B- 1 . Since the winding direction of the brake band of the brake B- 1  is opposite to the rotational direction of the drum-shaped member  18  as described above, i.e., the brake B- 1  is rotated in a releasing direction by the rotation of the drive wheel, the converted value of the brake B- 1  with respect to the output shaft torque becomes particularly small. 
     In the sixth failure case Fa 6  as a case where the brake B- 1  is engaged due to a failure of the linear solenoid valve SLB 1  in the worst condition in the fourth forward speed, the converted value of the clutch C- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutch C- 2  and the brake B- 1  are in the engaged states, i.e., in the state of the sixth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the seventh failure case Fa 7  as a case where the clutch C- 1  is engaged due to a failure of the linear solenoid valve SLC 1  in the worst condition in the fifth forward speed, the converted value of the clutch C- 3  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 2  are in the engaged states, i.e., in the state of the fourth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the eighth failure case Fa 8  as a case where the brake B- 1  is engaged due to a failure of the linear solenoid valve SLB 1  in the worst condition in the fifth forward speed, the converted value of the clutch C- 3  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutch C- 2  and the brake B- 1  are in the engaged states, i.e., in the state of the sixth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the ninth failure case Fa 9  as a case where the clutch C- 1  is engaged due to a failure of the linear solenoid valve SLC 1  in the worst condition in the sixth forward speed, the converted value of the brake B- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 1  and C- 2  are in the engaged states, i.e., in the state of the fourth forward speed, and the driving state is ensured without being brought to the stalled state. 
     In the tenth failure case Fa 10  as a case where the clutch C- 3  is engaged due to a failure of the linear solenoid valve SLC 3  in the worst condition in the sixth forward speed, the converted value of the brake B- 1  with respect to the output shaft torque becomes less than the limit torque Ttire at which the drive wheel slips based on the change of the torque distribution due to the failure. Thus, the clutches C- 2  and C- 3  are in the engaged states, i.e., in the state of the fifth forward speed, and the driving state is ensured without being brought to the stalled state. 
     Note that, as described above, since the engagement pressure P B2  supplied to the brake B- 2  is the control pressure P SLC2  of the linear solenoid valve SLC 2 , and is switched by the C- 2  relay valve  23  to be supplied to the hydraulic servo  45 , the control pressure P SLC2  is supplied to the hydraulic servo  42  in the case where the linear solenoid valve SLC 2  has failed. That is, the brake B- 2  cannot be caused to engage by a failure. Even if another clutch C- 2  or C- 3  or the brake B- 1  is engaged in the first forward speed, the shift speed merely changes without the simultaneous engagement of the three friction engagement elements. Therefore, the first to tenth failure cases Fa 1  to Fa 10  cover all possible single failures in which one linear solenoid valve fails. 
     Even if the first to tenth failure cases Fa 1  to Fa 10  occur as described above, merely a downshift of two speeds from the sixth forward speed to the fourth forward speed occurs in the ninth failure case Fa 9 , and merely an upshift or a downshift of one speed occurs in other failure cases. Therefore, the driving stability of the vehicle is not greatly influenced by the failure in particular. 
     As described above, according to the present invention, the normal state hydraulic pressure setting unit  72  sets the engagement pressure of two friction engagement elements such that slippage does not occur in the two friction engagement elements in a state where a shift speed is formed by the engagement of the two friction engagement elements (clutch or brake) and such that at least one of three friction engagement elements is caused to slip even if another friction engagement element engages based on the line pressure P L  while the two friction engagement elements are engaged. Therefore, the torque transmission can be performed between the driving source and the drive wheel without causing slippage of the friction engagement element while driving in the shift speed formed by the engagement of the two friction engagement elements, and the driving state can be ensured by one of the three friction engagement elements being caused to slip when another friction engagement element is engaged. Accordingly, provision of a cut-off valve becomes unnecessary to achieve a reduction in size, weight, and cost of the hydraulic control device. 
     The normal state hydraulic pressure setting unit  72  sets each engagement pressure such that the torque capacities of the two friction engagement elements forming the shift speed become torque capacities in which the safety ratios are taken into consideration in addition to the calculated transmission torques of the two friction engagement elements. Therefore, even if the torque fluctuates while driving in the shift speed formed by the engagement of the two friction engagement elements, the two friction engagement elements can reliably be prevented from slipping. 
     Further, since one of the three friction engagement elements is caused to slip by the inertia force of the vehicle, one of the three friction engagement elements can be reliably caused to slip without controlling the engine or the like in particular. 
     Specifically, when another friction engagement element engages while the two friction engagement elements are engaged, the torque distribution of the three friction engagement elements changes and one of the three friction engagement elements is applied with less than the limit torque Ttire at which the drive wheel slips. As a result, the one friction engagement element is caused to slip. Therefore, one of the three friction engagement elements can reliably be caused to slip. 
     Since the automatic speed change mechanism is brought to one of the shift speeds when another friction engagement element engages while the two friction engagement elements are engaged and one of the three friction engagement elements is caused to slip, the driving state can be ensured. 
     Particularly in the configuration of the automatic speed change mechanism  5 , even if the line pressure P L  is supplied to the hydraulic servo of another friction engagement element in the state where two friction engagement elements (clutch or brake) are engaged in the second forward speed to the sixth forward speed, the change of the torque distribution of the three friction engagement elements due to the simultaneous engagement of the three friction engagement elements causes the friction engagement element of which the torque distribution converted to the output shaft torque becomes minimum (the torque received from the drive wheel becomes maximum) to have a torque capacity, with respect to the drive wheel, of less than the inertia force of the vehicle (to receive a torque greater than the torque capacity of the hydraulic pressure setting in the normal state from the drive wheel) so as to slip. Accordingly, the driving state can be ensured. Since the engagement of the clutch C- 1  and the locking of the one-way clutch F- 1  achieve the first forward speed, the engagement of another friction engagement element merely causes a shift to one of the shift speeds, and the driving state can be ensured. 
     The brake B- 1  is formed of the band brake, and the band brake is arranged such that the rotational direction of the drum-shaped member  18  in the second forward speed to the sixth forward speed is the opposite direction of the winding direction of the brake band  19 . Therefore, the brake B- 1  can be made to easily slip by the inertia force of the vehicle even if the line pressure P L  is supplied to the hydraulic servo  44  of the brake B- 1 . Accordingly, particularly in the configuration of the automatic speed change mechanism  5 , one friction engagement element can be reliably caused to slip even if another friction engagement element is engaged in the state of any of the shift speeds. 
     Further, by a configuration in which the linear solenoid valves SLC 1 , SLC 2 , SLC 3 , and SLB 1  supply the control pressures P SLC1 , P SLC2 , P SLC3 , and P SLB1  as the engagement pressures P C1 , P C2 , P C3 , and P B1  in correspondence with the respective hydraulic servos  41 ,  42 ,  43 , and  44  as in the hydraulic control device  6 , the torque capacity (hydraulic pressure setting) of each friction engagement element can be set individually, and the torque capacity of each friction engagement element can be set such that two friction engagement elements do not slip while forming a shift speed and one friction engagement element slips when another friction engagement element is engaged. 
     Further, the brake B- 2  is arranged in parallel with the one-way clutch F- 1  and is formed of a brake which stops the rotation of the carrier CR 2  during coasting in the first forward speed, and the C- 2  relay valve  23  switches the hydraulic servo  42  of the clutch C- 2  and the hydraulic servo  45  of the brake B- 2  to supply the control pressure P SLC2  regulated by the linear solenoid valve SLC 2 . Therefore, a particularly small torque capacity suffices for the brake B- 2  during coasting in the first forward speed. Furthermore, the brake B- 2  is not engaged even if the linear solenoid valve SLC 2  outputs the control pressure P SLC2  in the shift speed other than the coasting in the first forward speed. Therefore, one friction engagement element can be reliably caused to slip even if another friction engagement element is engaged in the state of any of the shift speeds. 
     Note that, in the embodiment described above, when setting the hydraulic pressure of the friction engagement element to be engaged in the normal state, the torque capacity is set to a value obtained by multiplying the torque capacity based on the torque distribution and the input torque by the safety ratio. However, the safety ratio is to be set to an appropriate value in consideration of the output performance of the driving source (engine), and the grip performance of the drive wheel, for example. That is, as long as the hydraulic pressure setting in the normal state is performed such that slippage does not occur in two friction engagement elements forming a shift speed and such that one of the friction engagement elements slips when another friction engagement element is engaged, the safety ratio may take any value. Further, any method may be used as a calculation method of the hydraulic pressure setting in the normal state. 
     The automatic transmission  3  of the embodiment described above has been described as one example that can achieve the sixth forward speed. However, the present invention is obviously not limited thereto, and may be applied to any automatic transmission, as long as the hydraulic pressure setting of the automatic transmission in the normal state as described above can cause two friction engagement elements to not slip when the two friction engagement elements form a shift speed and cause one of three friction engagement elements to slip when another friction engagement element engages to prevent a simultaneous engagement. 
     The control device for an automatic transmission according to the present invention can be used for an automatic transmission mounted on a passenger car, a truck, a bus, or an agricultural machine, for example, and is particularly suitable for use in an automatic transmission which can ensure the driving state even if three friction engagement elements are simultaneously engaged and which is required to be reduced in size, weight and cost. 
     According to an exemplary aspect of the invention, the controller sets the engagement pressure of the two friction engagement elements such that slippage does not occur in the two friction engagement elements in the state where the engagement of the two friction engagement elements forms the shift speeds and such that, even if an additional friction engagement element engages based on the line pressure while the two friction engagement elements are engaged, one of the three friction engagement elements is caused to slip. Therefore, the torque transmission between the driving source and the drive wheel can be performed without causing slippage in the friction engagement elements while driving in the shift speeds formed by the engagement of the two friction engagement elements, and the driving state can be ensured by causing one of the three friction engagement elements to slip when an additional friction engagement element engages. Accordingly, provision of a cut-off valve can be made unnecessary, and a hydraulic control device can be reduced in size, weight, and cost. 
     According to an exemplary aspect of the invention, the controller sets the engagement pressure such that the torque capacity of the two friction engagement elements becomes the torque capacity in which the safety ratio is taken into consideration in addition to the calculated transmission torque of the two friction engagement elements. Therefore, slippage in the two friction engagement elements can reliably be prevented even if the torque fluctuates while driving in the shift speeds formed by the engagement of the two friction engagement elements. 
     According to an exemplary aspect of the invention, since one of the three friction engagement elements is caused to slip by the inertia force of the vehicle, one of the three friction engagement elements can reliably be caused to slip without controlling the driving source or the like in particular. 
     According to an exemplary aspect of the invention, the torque distribution of the three friction engagement elements changes when an additional friction engagement element engages while the two friction engagement elements are engaged and one of the three friction engagement elements is applied with less than the limit torque at which the drive wheel slips to cause slippage in the one friction engagement element. Therefore, one of the three friction engagement elements can reliably be caused to slip. 
     According to an exemplary aspect of the invention, the automatic speed change mechanism is brought to one of the shift speeds when an additional friction engagement element engages while the two friction engagement elements are engaged and slippage is caused in one of the three friction engagement elements. Thus, the driving state can be ensured. 
     According to an exemplary aspect of the invention, even if the line pressure is supplied to the hydraulic servo of an additional friction engagement element in a state where two friction engagement elements (clutch or brake) are engaged in the second forward speed to the sixth forward speed in particular, the change of the torque distribution of the three friction engagement elements due to the simultaneous engagement of the three friction engagement elements causes the friction engagement element of which the torque distribution converted to the output shaft torque becomes minimum (the torque received from the drive wheel becomes maximum) to have a torque capacity with respect to the drive wheel of less than the inertia force of the vehicle (to receive a torque greater than the torque capacity of the hydraulic pressure setting in the normal state from the drive wheel) so as to slip. Accordingly, the driving state can be ensured. Since the engagement of the first clutch and the locking of the one-way clutch achieve the first forward speed, the engagement of an additional friction engagement element merely causes a shift to one of the shift speeds, and the driving state can be ensured. 
     According to an exemplary aspect of the invention, the first brake is formed of the band brake, and the band brake is arranged such that the rotational direction of the drum-shaped member in the second forward speed to the sixth forward speed is the opposite direction of the winding direction of the brake band. Therefore, the first brake can be made to easily slip by the inertia force of the vehicle even if the line pressure is supplied to the hydraulic servo of the first brake. Accordingly, particularly in the structure of the automatic speed change mechanism, one friction engagement element can reliably be caused to slip even if an additional friction engagement element engages in the state of any of the shift speeds. 
     According to an exemplary aspect of the invention, with a configuration in which the first, second, third, and fourth solenoid valves supply the engagement pressures in correspondence with the respective hydraulic servos, the torque capacity (hydraulic pressure setting) of each friction engagement element can be set individually. Thus, the torque capacity of each friction engagement element can be set such that the two friction engagement elements do not slip while forming a shift speed and one friction engagement element slips when an additional friction engagement element engages. 
     According to an exemplary aspect of the invention, the second brake is arranged in parallel with the one-way clutch and is formed of the brake which stops the rotation of the second rotational element during coasting in the first forward speed, and the switching portion switches the hydraulic servo of the second clutch and the hydraulic servo of the second brake to supply the engagement pressure regulated by the second solenoid valve. Therefore, a small torque capacity suffices in particular for the second brake during coasting in the first forward speed. Since the second brake is not engaged even if the second solenoid valve outputs the engagement pressure in a shift speed other than the coasting in the first forward speed, one friction engagement element can be reliably caused to slip even if an additional friction engagement element engages in the state of any of the shift speeds.