Control system for torque converter of automatic transmission

Abstract: A control system for a torque converter (20) including lock-­ up clutch (26) provided in a torque converter (20) for connecting an input (61) and output (23) members of the torque converter (20) directly, engaging and disengaging hydraulic device (64,65) for controlling an engaging and disengaging action of the lock-up clutch (26), shift valve (80) provided with a first and second spools (83,84) in series for controlling an introduction of a driving hydraulic pressure introduced into the hydraulic device (64,65). The first spool (83) is subjected to a first hydraulic pressure at one end. The second spool (84) is subjected to a second hydraulic pressure at the other end. The first and second spool (83,84) are subjected to a third hydraulic pressure therebetween. The control system further includes first control device (101) for controlling the first hydraulic pressure, second control device (102) for controlling the second hydraulic pressure, adjusting valve (90) for adjusting the driving hydraulic pressure introduced into the engaging or disengaging hydraulic device, and third control device for controlling a hydraulic pressure for controlling the adjusting valve. A converter, lock-up and slip conditions of the torque converter (20) can be selectively established by switching shift positions of the single shift valve (80).

Description: The present invention relates to a torque converter for an automatic transmission having a lock-up clutch, more specifically to a slip control valve system for the lock-up clutch. Reference is also made to EP-A-0310117. In the field of a torque converter incorporated in an automatic transmission of an automotive vehicle, there has been proposed a lock-up clutch mechanism for providing a direct connection between engine output member and input member of the transmission under a predetermined vehicle operating condition wherein there occurs no amplification of a torque, or no change in torque from an engine, so that a fuel consumption performance of the vehicle is improved. When the lock-up clutch mechanism is engaged, the torque from the engine is transmitted bypassing the torque converter to the transmission through the lock-up clutch mechanism. Under this operation, it is advantageous in that the fuel consumption performance can be improved because there is no energy loss caused by a fluid friction in the torque converter. On the other hand, it is disadvantageous in that the torque amplification and a shock absorbing function of the torque converter cannot be obtained to thereby deteriorate a riding comfort. In view of the above characteristics of lock-up mechanism of the torque converter, JP-A-62-297567, laid open to the public on December 24, 1987, discloses a slip control system for a transmission of motor vehicle having a torque converter and lock-up clutch mechanism wherein the lock-up clutch mechanism is constituted by a converter cover connected with an engine output shaft and a damper piston connected with a turbine shaft with the damper piston being brought into frictional engagement with the converter cover in accordance with a hydraulic pressure introduced into hydraulic chambers formed at front and back side of the damper piston. The transmission is adapted to carry out a slip control in which the lock-up clutch allows a certain slippage so as to take selectively a lock-up condition wherein an engine torque is transmitted through only the lock-up clutch mechanism, a converter condition wherein the engine torque is transmitted through only the torque converter and a slip condition wherein the engine torque is transmitted through both the lock-up clutch and the torque converter. JP-A-57-33253 (corresponding to US-A-4468988 issued on September 4, 1984) laid open to the public in 1982, discloses a slip control system for establishing a proper slip condition wherein respective rotation speeds of an input and output members of the torque converter are detected and a hydraulic pressure for controlling the lock-up clutch is controlled for converging a rotation speed difference between the input and output members to a predetermined value by means of a feedback control. US-A-4580671, US-A-4618041, US-A-4669441, US-A-4687083 and US-A-4468988 also disclose conventional transmission control systems respectively. It should however be noted that in the conventional transmissions, a complicated hydraulic control mechanism including pressure regulator, control valve, gate valves, air bleed valves, check valves and the like had been needed for carrying out the slip control, thus, providing with the above three different conditions of the torque converter thereof. It is therefore an object of the present invention to provide a simple slip control system for an automatic transmission in structure. It is another object of the present invention to provide a hydraulic control system for an automatic transmission which can selectively take a lock-up condition, a converter condition and a slip condition in accordance with an vehicle operating condition. It is still another object of the present invention to provide a control system for an automatic transmission which can ease the torque shock. It is a further object of the invention to provide a control valve structure which can accomplish a simplified hydraulic control mechanism for an automatic transmission adapted to carry out a slip control of a lock-up clutch mechanism. US-A-3985046 discloses a control valve according to the preamble of claim 1 for a torque converter lock-up clutch in which the transition between fluid-drive at low vehicle speeds and non-fluid lock-up drive at higher vehicle speeds is relatively smooth. According to the present invention, there is provided a control system for a torque converter of an automatic transmission, the control system comprising: lock-up clutch means provided in a torque converter for connecting an input member and an output member of the torque converter directly; engaging and disengaging hydraulic means for controlling an engaging and disengaging action of the lock-up clutch means respectively; shift valve means provided with a first spool and a second spool in series for controlling an introduction of a driving hydraulic pressure for said engaging and disengaging hydraulic means, said first spool being subjected to a first hydraulic pressure at a first end thereof, said second spool being subjected to a second hydraulic pressure at a second end thereof opposite to said first end of the first spool, and said first and second spools being subjected to a third hydraulic pressure therebetween; first control means for controlling said first hydraulic pressure; adjusting valve means for controlling said driving hydraulic pressure for said engaging or disengaging hydraulic means; and, third control means for adjusting a control pressure for controlling said adjusting valve means; characterised by: second control means for controlling said second hydraulic pressure. According to the above control system, the first and second spools are moved in the same direction when the first control device changes the first hydraulic pressure from high to low or low to high. Where the second hydraulic pressure is changed by the second control device and the first hydraulic pressure is maintained at a low value, the first spool stays at its position and only the second spool is moved by virtue of the third hydraulic pressure. As a result, the shift valve can take selectively a first position wherein both the first and second spool are positioned at one side in a bore of the shift valve, a second position wherein the first and second spool are positioned at the other side in the bore, and a third position wherein the first spool is positioned at the one side and the second spool is positioned at the other side. The shift valve selectively takes the three different positions of the spools in accordance with vehicle operating condition to switch a hydraulic pressure introducing path for the lock-up clutch mechanism so that the lock-up mechanism establishes a lock-up condition when a hydraulic pressure is introduced only into the engaging hydraulic device, a converter condition when a hydraulic pressure is introduced only to the disengaging hydraulic device and a slip condition when a hydraulic pressure is introduced into both the engaging and disengaging hydraulic devices. The third control device controls a controlling hydraulic pressure for the adjusting valve so that the adjusting valve adjusts the hydraulic pressure for the disengaging device based on the controlling hydraulic pressure. This control of the hydraulic pressure for the disengaging device enables a pressure difference between the engaging and disengaging device to be changed during the slip condition so that a slippage in the lock-up clutch mechanism during the slip condition can be controlled with or without step. In addition, the third control device may be controlled by means of a feedback control in which a speed difference between engine and turbine speed is detected for the better slip control. The above and other features of the present invention will be apparent from the following description taking reference to the accompanying drawings. In the drawings:- Figure 1 is a schematic view of an automatic transmission to which a control system in accordance with the present invention can be applied; Figure 2 is a hydraulic control circuit incorporated into the automatic transmission of Figure 1 for controlling a lock-up mechanism; Figure 3A is a sectional view showing a shift valve and control valve employed in the hydraulic control circuit of Figure 2; Figure 3B is a sectional view showing the shift valve shown in Figure 3A but showing a different position; Figure 4 shows a relationship between the throttle opening and the throttle modulator pressure; Figure 5 shows a relationship between a duty ratio of a solenoid valve and a duty pressure; Figure 6 shows a relationship of a throttle opening, a duty ratio, a pressure difference between line pressures to respective chambers of a torque converter. Referring to the drawings, preferred embodiments of the present invention will be described hereinafter. Referring specifically to Figure 1, there is shown a schematic view of an automatic transmission 10. The transmission 10 is provided with a torque converter 20 and a multiple stage transmission gear mechanism 30. The gear mechanism 30 is provided with a plurality of frictional elements such as clutch, brake and the like for switching a power transmitting path to selectively establish a desirable gear stage among a first through fourth stage in a range D, the first through third stage in a range 2, the first and the second stage in a range 1, and a range R. The torque converter 20 is provided with a pump 22 mounted on one side of a case 21 fixed to an engine output shaft 1 for rotating together with the shaft 1, a turbine 23 facing the pump 22 and rotatably mounted on the other side of the case 21 for being driven by the pump 22 through a hydraulic oil, and a stator 25 disposed between the pump 22 and the turbine 23 for amplifying the torque when a speed ratio of the turbine 23 to the pump 22 is smaller than a predetermined value. The torque converter 20 is also provided with a lock-up clutch 26 for directly connecting the engine output shaft 1 or an input member of the torque converter 20 and the turbine 23 or an output member of the torque converter 20. The gear mechanism 30 is constituted by Ravigneaux-type planetary gear mechanism. The gear mechanism 30 is provided with a small sun gear 31 slidably mounted on the turbine shaft 27 and a large sun gear 32 slidably mounted on the turbine shaft 27 at a rear position of the small sun gear 31, a plurality of short pinion gears 33 meshed with the small sun gear 31, long pinion gear 34 meshed with the short pinion gears 33 at a front portion and with the large sun gear 32 at a rear portion thereof, a carrier 35 rotatably carrying the long pinion gear 34 and the short pinion gears 33, and a ring gear 36 meshed with a front portion of the long pinion gear 34. There are provided a forward clutch 41 and a first one-way clutch 51 in series between the turbine shaft 27 and the small sun gear 31. There is provided a coast clutch 42 in parallel with the clutches 41 and 51. A reverse clutch 44 is disposed between the turbine shaft 27 and the large sun gear 32. A 2-4 brake 45 constituted by a band brake is provided between the large sun gear 32 and the reverse clutch 44 for fixing the large sun gear 32. A 3-4 clutch 43 is disposed between the turbine shaft 27 and the carrier 35. Between the carrier 35 and a transmission case 11 are provided a second one-way clutch 52 and a low & reverse brake 46 for fixing the carrier 35 in parallel with each other. The ring gear 36 is connected with an output gear 14 so that a rotative power is transmitted from the output gear 14 to right and left wheels (not shown) through a differential mechanism. Hereinafter, there is described operations of frictional elements or brakes and clutches 41-46 and one-way clutches 51 and 52 in relation to the gear stages. In the first stage, the forward clutch 41 is engaged and the first and second one-way clutches 51 and 52 are locked. As a result, the output torque of the torque converter 20 is introduced from the turbine shaft 27 to the small sun gear 31 of the planetary gear mechanism through the forward clutch 41 and the one-way clutch 51. In this case, the carrier 35 is locked by virtue of the second one-way clutch 52 so that no differential operation occurs in the planetary gear mechanism when the power is transmitted from the small sun gear 31 to the ring gear 36 through the short pinion gear 33 and the long pinion gear 34. Thus, the first gear stage having a large reduction speed ratio corresponding to a gear ratio of the small sun gear 31 to the ring gear 36 can be obtained. In the second stage, the forward clutch 41 is engaged to lock the one-way clutch 51 and the 2-4 brake 45 is engaged to fix the large sun gear 32 and to race the second one-way clutch 52. As a result, the power transmitted from the turbine shaft 27 to the small sun gear 31 is transmitted to the long pinion gear 34 through the short sun gear 33 to rotate the long pinion gear 34 around the large sun gear 32 which is fixed in this situation. Thus, the carrier 35 is rotated. As a result, the rotation speed of the ring gear 36 is increased compared with the first stage due to the rotation of the carrier 35 (rotation of the long pinion gear around the large sun gear 32) so that the second stage having a smaller reduction speed ratio than the first stage can be established. In the third stage, the 2-4 brake 45 is disengaged and the 3-4 clutch 43 is engaged in the above mentioned condition of the second stage. The rotative power of the turbine shaft 27 is transmitted to the small sun gear 31 through the forward clutch 41 and the first one-way clutch 51 and concurrently to the carrier 35 through the 3-4 clutch 43. As a result, the planetary gear mechanism is integrally rotated as if one piece so that the third stage in which the ring gear 36 is rotated with the same speed as the turbine shaft 27 is established. In the fourth stage, the 2-4 brake 45 which is disengaged in the third stage is engaged again. As a result, the rotation of the turbine shaft 27 is transmitted to the carrier 35 from the 3-4 clutch 43 to rotate the long pinion gear 34 around the large sun gear 32. However, the large sun gear 32 is fixed by the 2-4 brake 45 so that the long pinion gear 34 is rotated around the large sun gear 32 and concurrently rotated by itself. This causes the rotation of the ring gear 36 meshed with the long pinion gear 34 to be increased by the self-rotation of the long pinion gear 34 in addition to the rotation of the carrier 35 ( rotation of the turbine shaft) so that the fourth stage of an over drive condition can be established. In this case, although the forward clutch 41 is maintained at the engaged condition, the one-way clutch 51 connected with the forward clutch 41 in series is racing. In addition, the coast clutch 42 is not engaged. Therefore, the rotation of the turbine shaft 27 is never transmitted to the small sun gear 31 in this stage. In the reverse stage, the reverse clutch 44 and the low & reverse brake 46 are engaged so that the rotation of the turbine shaft 27 is introduced into the large sun gear 32 and the carrier 35 is fixed. Consequently, the rotation is transmitted through an integrated gear train or through the large sun gear 32 and the long pinion gear 34 to the ring gear 36 so that a speed reduction ratio corresponding to a gear ratio of the ring gear 36 to the large sun gear 32 can be obtained. In this case, the ring gear 36 is rotated in a direction opposite to those of the turbine shaft 27 and the large sun gear 32. Thus, the reverse condition can be established. The one-way clutch 51 which transmits the rotation in the first through the third stage and the second one-way clutch 52 which bears a counter force in the first stage race under a coasting condition of the vehicle. However, the coast clutch 42 in parallel with the one-way clutch 51 is engaged in the third stage of the range D, the second and third stages of the range 2, the first and second stages of the range 1 and the low & reverse brake 46 in parallel with the second one-way clutch 52 is engaged in the first stage of the range 1 so that an engine brake effect can be obtained in those conditions. The table 1 shows a relationship between the respective gear stages and operations of the frictional elements 41-46 and the one-way clutches 51 and 52 of the transmission. In the table 1, reference * means that the corresponding element is raced under a coasting condition. Figure 2 shows a hydraulic control circuit 70 for controlling the lock-up clutch 26. The hydraulic circuit 70 is incorporated into an entire hydraulic control unit for controlling the whole automatic transmission as a part of it. Hydraulic control circuits for controlling a hydraulic pressure introduced into actuators of the respective frictional elements 41-46 are omitted from the illustration. The torque converter 20 is provided with the lock-up clutch 26 between the turbine 23 and the converter cover 61. The lock-up clutch 26 is constituted by a torsion damper 63 rotating integrally with the turbine shaft 27, a damper piston 62, and a friction plate (not shown) facing to a peripheral portion of the damper piston 62. The damper piston 62 divides an inner space of the converter cover 61 into a R-chamber 65 adjacent to the turbine 23 and a F-chamber 64 adjacent to the converter cover 61. The damper piston 62 is brought into a frictional engagement with the frictional plate by an engaging force in accordance with a pressure difference between the R-chamber 65 and the F-chamber 64. The lock-up clutch 26 selectively takes, according to the frictional engaging condition, a lock-up condition or engaged clutch condition wherein the rotation of the engine output shaft 1 is directly transmitted to the turbine shaft 27, a converter condition or disengaged clutch condition wherein the rotation of the engine output shaft 1 is transmitted to the turbine shaft 27 through the converter cover 61 and the turbine 23, and a slip condition wherein the converter cover 61 is brought into slidable engagement with the damper piston 62 so that the rotation of the engine output shaft 1 is transmitted through the pump 22 and the turbine 23 and partially through the lock-up clutch 26 to the turbine shaft 27. The hydraulic circuit 70 is provided with a lock-up shift valve 80 which selectively takes three positions corresponding to the above three conditions of the lock-up clutch 26, a lock-up control valve 90 for introducing a hydraulic pressure into the F-chamber 64 through the shift valve 80, a duty solenoid valve 101 for controlling a second pilot pressure of the shift valve 80 and the control valve 90 by means of a duty control, and a lock-up solenoid valve 102 for controlling a first pilot pressure of the shift valve 80 by means of an ON-OFF control. The three positions of the shift valve 80 mean a first position wherein the hydraulic pressure is introduced only into the R-chamber 65, a second position wherein the hydraulic pressure is introduced only into the F-chamber 64, and a third position wherein the hydraulic pressure is introduced into both the R-chamber 65 and the F-chamber 64. The hydraulic circuit 70 is provided with various hydraulic passages such as, a torque converter line L1 in which a hydraulic pressure from the pump 22 is adjusted to a predetermined line pressure by virtue of a regulator valve (not shown), a first pilot line L2 for introducing the first pilot pressure into the shift valve 80, a throttle modulator line L3 into which a throttle modulator pressure adjusted to a value corresponding to a throttle opening of the engine through a throttle modulator valve and a duty solenoid valve (not shown) is introduced, and a second pilot line L4 into which a predetermined second pilot pressure is introduced after being reduced by virtue of a solenoid reducing valve (not shown). The torque converter line L1 is divided into a line L11 connected to a port 82R of the shift valve 80 and a line L12 connected to a port 92F of the control valve 90. According to the illustration of Figure 2, spools 83 and 84 of the shift valve 80 are positioned at a right position or the first position. In this position, a port 81R is connected to the port 82R so that the line L11 introduces a hydraulic pressure of the torque converter into a line LR through the ports 81R and 82R. The hydraulic pressure from the line 12 is slightly reduced by the control valve 90 and introduced into a port 82F of the shift valve 80 through a port 91F and line L13 and concurrently introduced into a port 97 of the control valve 90 through a line L15. In the first position, a port 82F of the shift valve 80 is closed so that the torque converter pressure is not introduced into a line LF from the line L13. The line LR is connected with the R-chamber 65 through the pump 22 of the torque converter 20. The line LF is connected with the F-chamber 64. The hydraulic pressures in the R-chamber 65 and the F-chamber 64 are determined by the hydraulic pressures from the lines LR and LF. Since the lock-up clutch 26 is operated by an engaging force corresponding to a pressure difference between the R-chamber 65 and the F-chamber 64, the lock-up engaging force is determined by a pressure difference between a line pressure Pr of the line LR and a line pressure Pf of the line LF. For instance, in the first position as illustrated in Figure 2, the shift valve 80 provides only the line LR with the torque converter pressure so that the pressure difference between the R-chamber 65 and the F-chamber 64 is maximized. As a result, the lock-up condition or the engaged condition of the lock-up clutch 26 can be established. The first pilot line L2 is connected with a port 86 of the shift valve 80 through an orifice and with the solenoid valve 102 through a line L21. The solenoid valve 102 allows the first pilot line L2 to be drained in the ON-condition thereof so that the pilot pressure acting on a right end surface of the spool 84 as illustrated is reduced. The throttle modulator line L3 is connected with a port 93 of the control valve 90 to introduce the throttle modulator pressure into the port 93 as a pilot pressure for the control valve 90. The throttle modulator pressure is adjusted by virtue of the throttle modulator valve and the duty solenoid valve (not shown) in a manner that the value is increased as the throttle opening is increased. The second pilot line L4 is divided into a line 41 and a line 42. The line 41 is connected with a port 88 formed on a central portion of the shift valve 80 to introduce a predetermined hydraulic pressure such as 4 kg/cm2 as a pilot pressure between the spools 83 and 84. The line 42 is connected with a port 87 of the shift valve 80 through an orifice, with a port 96 of the control valve 90 through a line 44 and with the duty solenoid valve 101 through a line 43. The duty solenoid valve 101 allows the second pilot line L4 to be drained in an ON-condition thereof to thereby reduce or leak a pilot pressure on a left end surface of the spool 83 of the shift valve 80 and a pilot pressure on a right end surface of a spool 95 of the control valve 90 as illustrated. The duty solenoid valve 101 is subjected to an ON-OFF control in a certain cycle (duty ratio control). In other words, the pilot pressure of the line L42 is controlled by a duty ratio(a ratio of a valve open time period to a total period of a cycle) in the duty control applied for the duty solenoid valve 101. A line L6 is connected with a port 89 of the shift valve 80 to connect the port 89 to an oil cooler 75. A line LC in which a check valve is intervened connects the F-chamber 64 of the torque converter 20 with the oil cooler 75. Hereinafter, there will be described structures and operations of the shift valve 80 and the control valve 90 taking reference with Figures 3A and 3B. As illustrated, the shift valve 80 is provided with the spool 83 arranged within a sleeve thereof and urged by a spring 85 toward right, the spool 84 arranged rightward of the spool 83 in the sleeve and movable relative to the spool 83. There is disposed a spring 85a between the spools 83 and 84 as a retainer for providing smooth movements of the spools 83 and 84. As aforementioned, the left end surface of the spool 83 is subjected to a pilot pressure Pd from the line L42 and the right end surface 84a of the spool 84 is subjected to a pilot pressure Ps from the line L2. The spool 84 is provided with an enlarged portion at the right end thereof so that the right end surface 84a is subjected a larger pressure than the left end surface of the spool 83 because of the larger area than that of the left end surface of the spool 83. The control valve 90 is provided with the spool 95 arranged within a sleeve thereof and urged by a spring 99 rightward as illustrated. The spool 95 is subjected to the throttle modulator pressure (hereinafter referred to as Pm) from the line L3 through a port 93 at a left end surface thereof and subjected to the line pressure of the line L13 at a stepped portion 95c formed on a left land portion thereof through the line L14 in which an orifice is intervened and a port 94. The spool 95 is subjected to a duty pressure Pd of a line 44 at a right end surface 95a thereof and also subjected to a line pressure of the line L12 or the torque converter pressure at a stepped portion 95b formed on a right land portion thereof through the line 15 and the port 97. In operation, in the converter condition of the lock-up clutch 26, the duty ratio of the solenoid valve 101 is controlled to 0% so that the duty pressure Pd0 of the second pilot line L4 after being slightly reduced by the orifice is introduced into the left end surface of the spool 83 through the port 97. In this case, the solenoid valve 102 is in an OFF condition so that the pilot pressure Ps of the first pilot line L2 is introduced into the end surface 84a of the spool 84 through the port 86. As aforementioned, the pilot pressure Pc of the line L41 is normally introduced between the spools 83 and 84 through the port 88. Since the end surface 84a of the spool 84 is relatively large in the pressure area so that the leftward force on the spool 84 is greater than the rightward force on the spool 83 due to the resilient force of the spring 85 and the duty pressure Pd0, both the spools 83 and 84 are stayed at left positions by an upper half portion over a horizontal center line of the spools as illustrated in Figure 3A. In this position in which both the spools 83 and 84 are positioned at leftmost places, the port 82F and port 81R are connected with the port 81F and 89 respectively so that the torque converter pressure of the line L1 after being slightly reduced by the control valve 90 is introduced into the line LF and the line LF is brought into connection with the oil cooler 75 through the port 89. However, the shift valve 80 takes the second position. As a result, the torque converter pressure is introduced into the F-chamber 64 through the line LF, and drained through the R-chamber 65, the line LR and the line LC to the oil cooler 75. Thus, the lock-up clutch 26 is disengaged to establish the converter condition. On the other hand, in the lock-up condition of the lock-up clutch 26, the solenoid valve 102 is turned on as the duty ratio of the solenoid 101 is maintained at a value of 0 % so that the pilot pressure Ps of the line L2 is reduced. The resilient force of the spring 85 acting on the spool 83 overcomes the force acting on the end surface 84a of the spool 84 due to the pilot pressure so that the spool 83 is moved rightward together with the spool 84 in the illustration. As a result, the port 82R of shift valve 80 is brought into a connection with the port 81 R and the port 81F is brought into connection with a drain line D. In other words, the first position of the shift valve 80 is established, which is shown by a lower half portion divided by the horizontal center line of the spools 83 and 84 in Figure 2. In this condition, the torque converter pressure is introduced into the R-chamber 65 through the line LR and the pump 22, while the line LF is drained to release the hydraulic pressure in the F-chamber 64. The damper piston 62 is brought into frictional engagement with the frictional plate due to the pressure difference between the R-chamber 65 and the F-chamber 64 to establish the lock-up condition of the lock-up clutch 26. The duty ratio of the duty solenoid valve 101 is controlled beyond a predetermined value, for example, 20% as the solenoid valve 102 is maintained at the ON condition thereof so that the duty pressure Pd on the left end surface of the spool 83 is reduced. As a result, the spool 83 is moved in a left direction due to the pilot pressure Pc normally acting between the spools 83 and 84 while the spool 84 is stayed at the right position. In this case, the spool 83 is retained at a place where a rightward force due to the resilient force of the spring 85 and the duty pressure Pd is balanced with the leftward force due to the pilot pressure Pc. In other words, the shift valve 80 takes the third position as shown in Figure 3B. The spool 83 opens the port 82F to connect the line 13 with the line LF so that the torque converter pressure is introduced into the R-chamber 65 through the line LR and the pump 22 and the line pressure of the line L13 is introduced into the F-chamber 64 through the line LF. Consequently, the damper piston 62 is brought into the slidably frictional engagement with the frictional plate adjacent to the converter cover 61 in accordance with the pressure difference between the R-chamber 65 and the F-chamber 64 to establish the slip condition. In the slip condition, the control valve 90 controls the torque converter pressure in accordance with the duty ratio of the duty solenoid valve 101 and the throttle modulator pressure Pm of the throttle modulator line L3. In detail, the spool 95 is subjected to the duty pressure Pd through the line L44 at the right end surface 95a of the right land portion and to the torque converter pressure in the line L15 at a stepped portion 95b formed on a middle portion of the land portion as a pilot pressure. Since the duty pressure Pd and the throttle modulator Pm are introduced to the spool 95 in opposite directions to each other, the spool 95 is moved rightward to reduce the torque converter pressure when the duty ratio is increased or the duty pressure is reduced, and when the throttle opening is increased or the throttle module pressure Pm is increased. On the other hand, the spool 95 is moved leftward in the case of an increase of the duty pressure Pd or a decrease of the throttle modulator pressure Pm. Thus, the line pressure of the line L13 is adjusted in accordance with a hydraulic pressure balance between the duty pressure Pd and the throttle modulator pressure Pm. The spool 95 is subjected to a rightward force due to the resilient force of the spring 99, throttle modulator pressure Pm at a left land portion and subjected to a line pressure of the line L14 separated from the line L13 through an orifice at a stepped portion 95c formed on a middle portion of the left land portion thereof. This is effected to suppress a fluctuation of the torque converter pressure Pr on the R-chamber 65 and the line pressure Pf on the F-chamber 64. Figure 4 shows a relationship between the throttle opening and the throttle modulator pressure Pm. Figure 5 shows a relationship between the duty ratio of the duty solenoid valve and the duty pressure Pd, and Figure 6 shows a relationship of the throttle opening, the duty ratio, the pressure difference between the line pressures Pr and Pf. As shown in Figure 4, the throttle modulator pressure Pm is increased as the throttle opening is increased. On the other hand, the duty pressure Pd is reduced due to the increase of the duty ratio. The control valve 90 is controlled based on the duty pressure Pd and the throttle modulator pressure Pm for an improved and responsive control of the pressure of R and F-chambers 65 and 64. The line pressure Pf which determines the F-chamber 64 is controlled based on the control valve 90 when the shift valve 80 is set at the third position for establishing the slip condition of the lock-up clutch 26. At this moment, the torque converter pressure in the line L11 is introduced to the line LR to constitute the line pressure Pr which determines the pressure in the R-chamber 65 so that the pressure difference dP between the line pressures Pr and Pf can be controlled by the control valve 90 based on the pressure difference between the duty pressure Pd and the throttle modulator pressure Pm. Meanwhile, the duty solenoid valve 101 which controls the duty pressure Pd is controlled by means of a feedback control so as to reduce a speed difference between the engine and the turbine 23 within a certain level. The throttle modulator pressure Pm is increased as the throttle opening is increased. Thus, the spool 95 is moved due to the increase of the throttle opening to reduce the line pressure Pf so that the pressure in the F-chamber 64 is reduced to displace the lock-up clutch 26 toward the disengaged condition. In this control, the duty pressure Pd and the throttle modulator pressure Pm are controlled independently from each other so that the pressure difference Pd can be widely controlled. The control valve 90 is controlled based on not only the speed difference for the feedback control but also the throttle opening. This makes the slip control responsive. The control valve 90 intervenes the line L15 and the line L14 so that a pressure fluctuation can be reduced during the slip control of the lock-up clutch 26. According to the present invention, the three different conditions of the lock-up clutch 26 can be readily established by switching the positions of the single shift valve 80. In the slip condition, the engaging force of the lock-up clutch can be smoothly controlled through a control for the control valve 90. In the above embodiment of the present invention, the duty solenoid valve 101 is commonly provided for controlling both the shift valve 80 and the control valve 90. It should however be noted that solenoid valves can be provided for the valves 80 and 90 independently.

Claims: A control system for a torque converter of an automatic transmission, the control system comprising: lock-up clutch means (26) provided in a torque converter (20) for connecting an input member (1) and an output member (27) of the torque converter (20) directly; engaging (65) and disengaging (64) hydraulic means for controlling an engaging and disengaging action of the lock-up clutch means respectively; shift valve means (80) provided with a first spool (84) and a second spool (83) in series for controlling an introduction of a driving hydraulic pressure for said engaging and disengaging hydraulic means (65,64), said first spool (84) being subjected to a first hydraulic pressure (Ps) at a first end thereof, said second spool (83) being subjected to a second hydraulic pressure (Pd) at a second end thereof opposite to said first end of the first spool, and said first and second spools (84,83) being subjected to a third hydraulic pressure (Pc) therebetween; first control means (102) for controlling said first hydraulic pressure (Ps); adjusting valve means (90) for controlling said driving hydraulic pressure for said engaging or disengaging hydraulic means; and, third control means for adjusting a control pressure for controlling said adjusting valve means; characterised by: second control means (101) for controlling said second hydraulic pressure (Pd); A control system according to claim 1, wherein said shift valve means (80) selectively takes a first position wherein both the first and second spools (84,83) are positioned at a first end portion of the shift valve means (80), a second position wherein the first and second spools (84,83) are positioned at a second end portion opposite to the first end portion, and a third position wherein the first spool (84) is positioned close to the first end portion and away from the second spool (83) and the second spool (83) is positioned at the second end portion. A control system according to claim 2, wherein said shift valve means (80) takes said first position when the first control means (102) provides the first hydraulic pressure (Ps) with a low value. A control system according to claim 2, wherein said driving hydraulic pressure is introduced into said engaging hydraulic means (65) to establish a lock-up condition of said torque converter when the shift valve means (80) takes said first position. A control system according to claim 2, wherein said shift valve means (80) takes said second position when the first control means (102) provides the first hydraulic pressure (Ps) with a high value. A control system according to claim 2, wherein said driving hydraulic pressure is introduced into said disengaging hydraulic means (64) to establish a converter condition of said torque converter when said shift valve means (80) takes said second position. A control system according to claim 2, wherein said shift valve means (80) takes the third position when the second control means (101) reduces said second hydraulic pressure (Pd) in the case when the first control means (102) provides the first hydraulic pressure (Ps) with a low value. A control system according to claim 2, where said driving hydraulic pressure is introduced into both said engaging and disengaging hydraulic means (65,64) to establish a slip condition of the torque converter when said shift valve means (80) takes said third position. A control system according to claim 8, wherein said second spool (83) is positioned in accordance with said second and third hydraulic pressures to define said driving hydraulic pressure for the disengaging hydraulic means. A control system according to any of claims 1 to 9, wherein said first control means (102) is constituted by a solenoid valve subjected to an ON-OFF control. A control system according to any of claims 1 to 10, wherein said second control means (101) is constituted by a duty solenoid valve. A control system according to any of claims 1 to 11, wherein said third control means controls said control pressure in accordance with throttle valve opening in a manner that the control pressure is increased as the throttle valve opening is increased. A control system according to any of claims 1 to 12, wherein said adjusting valve means (90) comprises a spool (95) which is subjected to said second hydraulic pressure at one end and to said control pressure at the other end. A control system according to claim 13, wherein said spool (95) is positioned in accordance with said second hydraulic pressure and said control pressure to change the driving hydraulic pressure for the disengaging hydraulic means. A control system according to any of claims 1 to 14, wherein said second hydraulic pressure is determined in accordance with a speed difference between an engine and a turbine of the torque converter.

Inventors: FUJIWARA TAKUJI - - MINAMITAIS, ISHII KOZO SHOFURYO - -, TAKEMOTO KAZUO, YOSHIMURA HIROSHI, FUJIWARA, TAKUJI 5-19-510, MINAMITAISHO-MACHI, ISHII, KOZO SHOFURYO 422, 5-11-5, TAKEMOTO, KAZUO, YOSHIMURA, HIROSHI
Applicants: MAZDA MOTOR, MAZDA MOTOR CORPORATION
Application year: 1989
Main classification: F16H61
IPC codes: F16H61