JP4931217B2 - Automatic transmission input joint - Google Patents

Description

The present invention relates to an automatic transmission for a vehicle, and more particularly to an input joint of a multi-stage automatic transmission for driving front wheels controlled using a hydraulic clutch and a brake.

Generally, as an input joint of an automatic transmission for a vehicle, there is a joint having a lock-up clutch that mechanically conducts direct coupling inside a torque converter having a fluid conduction function including a pump impeller, a turbine runner, and a wheel stator. Used. Many of these lock-up clutches use a hydraulic pressure inside the torus of the torque converter to press the piston integrated with the friction member against the side wall of the front cover and transmit power to the turbine runner via the torsion damper. The oil chamber and the oil chamber on the front cover side are separated and formed by the friction surface of the piston, and the oil supply to both chambers is switched from the outside of the input joint so that the lockup clutch can be engaged and disengaged. Yes.

However, in terms of strength of the torque converter, the hydraulic pressure inside the torus cannot be increased as much as the hydraulic pressure used for the clutch of the normal transmission, and the friction surface is also one, limiting the transmission capacity, and the torus side oil chamber and the front cover side oil chamber. Are separated from each other by the friction surface of the piston, the oil flows from the torus side to the oil chamber on the front cover side when engaged, and the time for moving the piston having a large pressure receiving area becomes longer. In addition to such poor responsiveness, the hydraulic pressure inside the torus varies depending on the speed ratio between the pump impeller and the turbine runner, which makes it difficult to control hydraulic pressure from the outside. The pressure acting on the piston here is the pressure between the turbine runner and the front cover, which is different from the oil pressure inside the torus where a large amount of oil circulates, but is influenced by the oil pressure inside the torus. I have a problem. As a method for increasing the transmission capacity of the lock-up clutch, a number of methods have been devised in which the friction surface is doubled and the piston is pressed by the hydraulic pressure inside the torus, but the problem of engagement / disengagement response cannot be solved.

Many of the conventional automatic transmissions for vehicles have four forward speeds, one reverse speed, and the torque converter is locked up only at a certain speed of the maximum speed or more, and the engine brake is in reverse drive. However, the frequency of use of the lock-up is low, such as when the lock-up is released, and the lock-up clutch has a simple structure as described above and has no problem. However, since the frequency of lock-up use is low, the torque transmission capacity and efficiency at a high speed ratio must be increased, and it is necessary to increase the outer diameter or increase the torus area. However, if the torque transmission capacity of the high speed ratio is increased, the transmission capacity of the low speed ratio is also increased more than necessary.Therefore, the baffle plate effect that disturbs the flow inside the torus, which is often used in fluid couplings without a wheel stator, is provided. Some have a stator. These measures cause an increase in weight, which not only leads to an increase in cost but also against the improvement in fuel consumption.

As is well known, in recent years, there has been a strong demand for fuel saving in automobiles due to global environmental problems, and automatic transmissions for vehicles have been promoted to multistage forward gears of 6 or more that can reduce the difference in gear ratio between adjacent gears. However, it is no longer necessary to conduct the fluid to reduce fuel consumption except at low speed stages. In addition, it is a matter of course that the fuel of the prime mover is cut by lock-up in the vehicle coasting and reverse drive state. The torque converter lock-up clutch required for these multi-stage automatic transmissions can be used for lock-up engagement when the speed ratio between the pump impeller and the turbine runner is large, or for reducing shocks during gear shifting, including interlaced shifting. It is required that the release and slip can be performed with good response, and the conventional lock-up clutch as described above is insufficient.

Therefore, this multi-speed automatic transmission with 6 or more forward speeds has an independent hydraulic engagement chamber in the input joint that can be fastened regardless of the speed ratio between the pump impeller and the turbine runner, and is transmitted by having multiple friction surfaces. A lock-up clutch that can increase the capacity is suitable. This type of lock-up clutch is mainly used in torque converters for automatic transmissions for trucks and buses, and has a long history. For example, in a hydraulic circuit diagram of a transmission using this type of lockup clutch described in Patent Document 1 in 1955, oil is supplied into the torus from between the pump impeller and the wheel stator, and the turbine runner A circuit that controls the three-way circuit to the torque converter that supplies oil to the hydraulic engagement chamber of an independent lock-up clutch provided on the front cover side and the circulation circuit that is discharged from between the motor and the wheel stator is shown. However, even if the control valves are different, the system is controlled by these three passages.

In order to simplify the three passages to the outside, in Patent Document 2, the torus chamber of the torque converter and the two independent hydraulic engagement chambers provided on the front cover side in this type of lockup clutch are communicated with each other through a through hole. In addition, the flow of oil from the torus chamber to the hydraulic engagement chamber of the lock-up clutch can be controlled, and control from the outside can be performed by using a check valve that prevents the reverse flow and an orifice with a small diameter in the through hole. Proposals have been made in which the circuit has two paths. However, it has factors that make the operating pressure of the piston unstable, such as variations in pressure difference due to changes in the amount of oil passing through the orifices of the two chambers and the check valve through holes, and the reliability of the check valve.

Moreover, as an embodiment using a torque converter for mass production in which a front cover and a pump impeller are welded integrally on the outer peripheral portion and a hydraulic engagement chamber of an independent lock-up clutch is provided on the front cover side, as an example using a mass production torque converter in recent years, There are 6-speed automatic transmissions of Non-Patent Documents 3 and 4 developed and described in SAE PAPER. As shown in the structural diagram of this torque converter, in this type of lockup clutch, the piston that engages the lockup slides with the front cover or the boss integrated with the front cover sealed on the inner and outer peripheries. The front cover and the input shaft portion of the transmission have a double seal structure that provides a unique rotational seal. In addition, in FIGS. 11 and 12 of Patent Document 5 in which many lock-up clutch configurations are described, there is also a method in which the outer periphery of the front cover, which is a connecting portion of the friction member, is plastically processed in a spline shape and the retaining ring groove is processed. As described in Non-Patent Documents 3 and 4, however, the retaining ring is plastically processed into a spline shape because the outer peripheral groove processing part of the press material for mass production is insufficient in strength against the centrifugal hydraulic pressure inside. A method is adopted in which another member that has been grooved is welded to the front cover. Each of these is a complicated and expensive structure with an increased number of parts, and is not suitable for an input joint for driving front wheels that is compact and requires low cost.

By the way, in the multi-speed automatic transmission with 6 or more forward speeds, the number of parts increases and the axial direction has to be longer as compared with the conventional forward 4-speed automatic transmission. Therefore, further weight reduction and axial compactness are required. Most of the conventional manual transmissions for vehicles are 5 forward speeds and 1 reverse speed, and the traction capability is the forward 5 speed stage where the forward speed stage is the lowest speed stage and the forward speed stage is the fastest speed stage. It is determined by the gear range divided by the gear ratio, and its value is generally 5-6. In a 6-speed automatic transmission developed in recent years, this value is about 6 which is almost the same as that of a 5-speed manual transmission. Since the performance of the prime mover cannot be fully used as in a manual transmission, there is torque amplification. A torque converter is used. However, it is not necessary to increase the torque ratio as much as the torque converter used in the forward four-speed automatic transmission which has a gear range of only 4 to 5 in the forward six-speed automatic transmission, and also has a high speed ratio for locking up. There is no need to increase the torque transmission capacity, and the torque converter can be compact with a reduced torus area, such as by reducing the torque ratio and reducing the outer diameter or increasing the inner diameter. In addition, a fluid coupling that does not require a baffle plate that reduces the capacity at a low speed ratio and that is more compact and low in cost can be applied as an input coupling. The present invention is a proposal relating to a torque converter and a fluid coupling for such a multi-stage automatic transmission.

US2726557 JP 52-11364 A SAE PAPER 2004-01-0652 SAE PAPER 2004-01-0653 Special table 2001-503126

A first object of the present invention is to provide a control oil passage 2 from the outside of an input joint of a lockup clutch that presses and engages a piston in an independent hydraulic engagement chamber provided on the front cover side in the input joint. A simple hydraulic engagement chamber structure that can be used as a passage is proposed, and an input joint that can be applied to a fluid joint that makes it difficult to control three passages is proposed.

The second problem of the present invention is to reduce the number of internal components of the input joint equipped with a lock-up clutch that presses and engages the piston in an independent hydraulic engagement chamber provided on the front cover side in the input joint. A simple and compact structure.

The third problem of the present invention is that the connecting portion of the friction member of the lock-up clutch that presses and engages the piston in an independent hydraulic engagement chamber provided on the front cover side in the input joint is simple and compact. It is a structure.

A fourth problem of the present invention is that an internal component member of a torque converter equipped with a lock-up clutch that presses and engages a piston in an independent hydraulic engagement chamber provided on the front cover side in the input joint is automatically It arrange | positions so that the axial direction of a transmission may become compact.

The present invention according to claim 1 relates to the structure and control of a hydraulic engagement chamber in a lockup clutch, and is a means for solving the first and second problems, and is integrated by a front cover and a pump impeller. An input member having an enclosed oil chamber, and a turbine runner that is arranged between the front cover and facing the pump impeller and that is connected to the input shaft of the transmission, or in addition, a pump impeller and a turbine runner Adjacent to the front cover, a fluid conducting portion comprising a wheel stator disposed between the friction member, a friction member for connecting the input member and the turbine runner, a torsion damper mechanism provided between the friction member and the turbine runner, The oil chamber is held in an axially movable but non-rotatable manner and is divided into an oil chamber B on the front cover side and an oil chamber A on the pump impeller side. The oil chambers A and B of the piston of the input joint of the automatic transmission having a mechanical transmission portion by a lock-up clutch comprising a piston that presses and engages the friction member with the hydraulic pressure supplied to the chamber B are formed separately. The oil chamber B is hermetically sealed at the inner and outer sliding portions in a state where the pressure of the oil chamber B is higher than the pressure of the oil chamber A, and the oil chambers A and B are sealed in a state where the pressure of the oil chamber A is higher than the pressure of the oil chamber B Using an elastic seal member that communicates, the oil chambers A and B are communicated to the outside of the input joint, and hydraulic pressure is supplied to the oil chamber B in a conductive state where the lock-up clutch is engaged and the oil chamber A is supplied. The oil in the oil chamber A is supplied to the oil chamber A in a fluid conduction state where the lockup clutch is released, and the oil in the oil chamber A is discharged to the outside of the input joint through the oil chamber B. .

The present invention according to claim 2 relates to the detailed seal structure of the hydraulic engagement chamber of the lock-up clutch according to claim 1, and is a means for solving the first and second problems. The seal member of the outer peripheral sliding portion that separates and forms the chambers A and B has a shape in which the hydraulic pressure of the oil chamber A acts on the outer peripheral side and the hydraulic pressure of the oil chamber B acts on the inner peripheral side. The seal member has a shape in which the oil pressure in the oil chamber A acts on the inner peripheral side and the oil pressure in the oil chamber B acts on the outer peripheral side, and is elastically deformable in the radial direction by the differential pressure between the oil chambers A and B. It was welded to or attached independently.

The present invention according to claim 3 relates to the arrangement structure of the connecting portion of the friction member of the lockup clutch and the piston and the operational stability of the seal member attached to the piston according to claims 1 and 2, A means for solving the second and third problems, wherein the input member is provided with a radial step so that the inner diameter in the distal direction increases, and a plurality of protrusions extending in the axial direction on the distal side from the radial step are provided . The outer peripheral part of the pump impeller that has it is arranged on the inner side in the radial direction of the outer peripheral part of the front cover and welded . The pump impeller is fitted with the protrusion and is prevented from rotating, and is in contact with the radial step and moved axially toward the pump impeller. and end plates but are not limited to, a friction member is connected via the torsion damper mechanism to the turbine runner, or in addition to them, axially moved while being prevent rotation fitted with projections It arranged sequentially and freely driven plate from the pump impeller side, piston or to distribution freely axially move along the inner diameter of the protruding portion or the front cover the outer periphery as well as prevent rotation mates with the protrusion, the piston The driven plate or the friction member is pressed by making it thin and fixing the inner diameter side to the front cover and preventing it from rotating and elastically deforming the outer diameter side so as to be axially movable.

The present invention according to claim 4 is an embodiment relating to a fluid coupling comprising a pump impeller and a turbine runner according to claims 1 and 3, and is means for solving the first and second problems. A friction member integrally formed on the outer periphery of a clutch hub connected to the turbine runner via a torsion damper mechanism so as to be swingable in a rotational direction through a torsion damper mechanism. It was arranged between the end plate and the piston.

The present invention according to claim 5 is an embodiment relating to a torque converter comprising a pump impeller, a turbine runner and a wheel stator according to claims 1 and 3, and is means for solving the first and second problems. The outer periphery of the clutch hub, which is connected to the turbine runner through a torsion damper mechanism and is pivotable in the rotational direction through a torsion damper mechanism , is bent into an L shape toward the pump impeller side. The torque converter is composed of a pump impeller, a turbine runner, and a wheel stator. The drive plate sandwiched between the end plate and the piston is engaged with the cylindrical portion by a spline, and a friction member is provided on the drive plate, and the pump impeller side end portion of the spline that engages the drive plate of the clutch hub a projection provided on, regulations movement to the front cover direction of the clutch hub drive plate It was no way to.

The present invention according to claim 6 is another embodiment relating to the fluid coupling and torque converter according to claims 1 and 3, and is a means for solving the first problem. fluid coupling comprising a turbine runner or a torque converter comprising a pump impeller and the turbine runner and the wheel stator, the outer periphery of the piston and the front cover and the sliding, the inner circumference of the integral to the turbine runner and or the front cover boss sliding dynamic and, the inner periphery sealing the input shaft integral boss and the automatic transmission to the turbine runner or the front cover to slide the seal member, a clutch hub and the friction member for connecting via a torsion damper mechanism to the turbine runner and relatively movable in the axial direction, the member or between the front of the front cover and the turbine runner and the pump impeller It arranged a thrust bearing between the members of Tokaba and the turbine runner and the wheel stator and the pump impeller.

The present invention according to claim 7 relates to the arrangement of the internal components of the torque converter according to claims 1 and 6 , and is a means for solving the fourth problem. A pump impeller, a turbine runner, and a wheel stator. The one-way clutch is arranged on the radially outer side of the housing boss portion so that the inner diameter r2 of the one-way clutch of the wheel stator of the torque converter is larger than the outer diameter r1 of the housing boss portion serving as the transmission-side shaft support portion of the torque converter, The thrust bearing between the turbine runner and the wheel stator is arranged radially inside so as to be smaller than the inner diameter r2 of the one-way clutch, and the thrust bearing between the wheel stator and pump impeller is arranged radially outside so as to be larger than the inner diameter r2 of the one-way clutch. did.

According to the first aspect of the present invention, the turbine is disposed between the input member having the oil chamber integrally surrounded by the front cover and the pump impeller and the front cover facing the pump impeller and connected to the input shaft of the transmission. In addition to the runner, or in addition to them, a fluid conducting portion comprising a wheel stator disposed between the pump impeller and the turbine runner, a friction member for connecting the input member and the turbine runner, a friction member and the turbine runner A torsion damper mechanism provided between the front cover and the oil chamber B which is movable in the axial direction and is held in a relatively non-rotatable manner is separated into an oil chamber B on the front cover side and an oil chamber A on the pump impeller side. An automatic changer provided with a mechanical transmission part by a lock-up clutch comprising a piston that presses and engages the friction member with the hydraulic pressure supplied to the chamber B The oil chamber B is hermetically sealed in the state where the pressure of the oil chamber B is higher than the pressure of the oil chamber A, on the inner and outer peripheral sliding portions of the input joint of the apparatus that separate and form the oil chambers A and B of the piston. Using an elastic seal member that connects oil chambers A and B when the pressure in oil chamber A is higher than the pressure, and connecting oil chambers A and B to the outside of the input joint and mechanically connected directly to the lock-up clutch. The hydraulic pressure is supplied to the oil chamber B in the conductive state and the oil in the oil chamber A is discharged to the outside of the input joint, and the hydraulic pressure is supplied to the oil chamber A in the fluid conductive state in which the lockup clutch is released and the oil chamber A Oil is discharged to the outside of the input joint through the oil chamber B, so that a large amount of oil can be checked with a simple structure without using a special check valve. Stable lock-up control with two oil passages can be performed.

In the configuration according to claim 2, the oil pressure of the oil chamber A acts on the outer peripheral side and the oil pressure of the oil chamber B acts on the inner peripheral side of the seal member of the outer peripheral sliding portion that separates and forms the oil chambers A and B of the piston. The seal member of the inner peripheral sliding portion has a shape in which the oil pressure of the oil chamber A acts on the inner peripheral side and the oil pressure of the oil chamber B acts on the outer peripheral side, and the differential pressure between the oil chambers A and B Since it can be elastically deformed in the radial direction and is welded to the piston or mounted independently, the sliding that separates and forms the oil chambers A and B of the piston with a slight elastic margin of the seal member A large opening is made in the part, and a large amount of oil can be checked. Further, the responsiveness of the lock-up clutch can be ensured by the relief effect of the oil flow from the oil chamber A to the oil chamber B by the elastic force of the seal member and the quick check action.

In the configuration of claim 3, wherein the input member is provided with a radial step to the tip end direction of the inner diameter is increased, the front of the pump impeller outer peripheral portion having a plurality of projections extending in the axial direction on the distal end side than the radial step An end plate which is arranged and welded radially inside the outer periphery of the cover , is fitted with the protrusion and is prevented from rotating, and is in contact with the radial step and is restricted from moving in the axial direction toward the pump impeller. Friction members connected to the turbine runner via a torsion damper mechanism, or in addition to them, a driven plate that fits with the protrusion and is prevented from rotating and is axially movable is arranged in order from the pump impeller side. , piston or to distribution freely axially move along the inner diameter of the protruding portion or the front cover the outer periphery as well as prevent rotation mates with the protrusion, the thin piston The inner diameter side is fixed to the front cover and prevented from rotating, and the outer diameter side is elastically deformed so as to be axially movable so as to press the driven plate or the friction member. The connection part which does not have the problem of the strength by centrifugal oil pressure can be arranged simply and compactly without using a special member in the outer peripheral space of the front cover. In addition, the piston that separates and forms the oil chambers A and B can be centered, and a constant elastic margin can be secured in the radial direction for the seal member of the outer and inner peripheral sliding portion, and the reverse of the oil from the oil chambers A to B Stabilization is stable.

According to a fourth aspect of the present invention, friction is integrally provided on the outer periphery of a clutch hub that is coupled to the turbine runner through a torsion damper mechanism so as to be able to swing in the rotational direction of a fluid coupling comprising a pump impeller and a turbine runner. Since the member is disposed between the end plate whose axial direction is regulated and the piston, the turbine runner is floated in the axial direction by the amount of piston breakage, and the thrust force generated in the turbine runner is the friction member of the lockup clutch. Since it is transmitted, it can be used together with fluid conduction, and there is no need to provide a thrust bearing, and the inner periphery sliding part seal of the piston and the rotation seal of the input shaft of the transmission can be shared, and the number of components can be reduced.

According to the fifth aspect of the present invention, the outer periphery of the clutch hub, which is connected to the turbine runner through the torsion damper mechanism so as to be able to swing in the rotational direction, of the torque converter including the pump impeller, the turbine runner, and the wheel stator is directed to the pump impeller side. A cylindrical portion is bent into an L-shape, and a drive plate sandwiched between an end plate and a piston is engaged with the cylindrical portion by a spline, and a friction member is provided on the drive plate and engages with a drive plate of the clutch hub. a projection on the pump impeller side end portion of the spline is provided, since no movement of the front cover direction of the clutch hub so as to regulate the drive plate, usually, the thrust force of the turbine runner in fluid conducting state only the pump impeller side Acts and does not act on the front cover side. Movement of Nran'na limits friction member without disposing a thrust bearing, the rotary seal of the input shaft of the inner peripheral sliding seal with the transmission of the piston also can reduce the components can be shared.

In the configuration of claim 6, the fluid coupling comprising a pump impeller and a turbine runner serving as a fluid conducting part or a torque converter comprising a pump impeller and the turbine runner and the wheel stator, the outer periphery of the piston and the front cover and the sliding , the inner circumferential slides integral with the boss to the turbine runner and or the front cover, the input shaft of the inner circumference of the integral to the turbine runner or the front cover slides boss and the automatic transmission is sealed by a sealing member, a turbine runner The clutch hub and the friction member, which are connected to each other via a torsion damper mechanism, can be moved relative to each other in the axial direction, and between the front cover, the turbine runner, and the pump impeller , or between the front cover, the turbine runner, the wheel stator, and the pump impeller. Thrust bearings are arranged between members, so the components The thrust force generated in the turbine runner does not act on the friction member of the lock-up clutch in any state, but the control oil passage from the outside of the input joint has more stable lock-up control. Can do.

According to the seventh aspect of the present invention, the inner diameter r2 of the one-way clutch of the wheel stator of the torque converter including the pump impeller, the turbine runner, and the wheel stator is larger than the outer diameter r1 of the housing boss that serves as the transmission side shaft support portion of the torque converter. The one-way clutch is arranged on the outer side in the radial direction of the housing boss, and the thrust bearing between the turbine runner and the wheel stator is arranged on the inner side in the radial direction so as to be smaller than the inner diameter r2 of the one-way clutch, and the thrust between the wheel stator and the pump impeller Since the bearing is arranged radially outward so as to be larger than the inner diameter r2 of the one-way clutch, the lock-up clutch of this system that is compact by arranging the torsion damper mechanism on the inner diameter side of the torus is a transmission device that supports the torque converter. Boss Can be avoided as much as possible the interference of the inner component may be distribution such that the axial direction of the automatic transmission torque converter becomes compact.

1 to 6 show the fluid coupling of the present invention, and FIGS. 7 to 12 show the torque converter of the present invention. FIG. 13 shows a connecting portion of a friction member of a lock-up clutch commonly used for a fluid coupling and a torque converter as an input joint of the present invention, and FIG. 14 shows details of a seal member of a piston sliding portion. 2 and 3 and FIGS. 8 and 9 are obtained by mounting the input joint of the present invention to the multi-stage automatic transmission proposed by the inventor in Japanese Patent Application No. 2006-306162. The structural diagram of the automatic transmission including this input joint is based on the concept that the torque of the prime mover is 300 Nm, and the total length is 350 mm or less.

The separation structure of the oil chamber B engaging the piston on the front cover side and the oil chamber A on the pump impeller side and the control from the outside of the input joint according to claim 1 of the present invention are shown in FIGS. It is shown in all the drawings of FIGS. 1 to 14 except for FIG. The non-return action of the oil from the oil chamber A to the oil chamber B of the seal member of the piston sliding portion according to claim 2 is illustrated in FIG. 14 and the seal structure welded to the piston is shown in FIGS. FIG. 5, FIG. 11 and FIG. 12 show a seal structure that is mounted independently. FIG. 13 shows a detailed structure in which the pump impeller member according to claim 3 is used for the connecting portion of the friction member of the lock-up clutch, and FIG. 4 shows a structure in which the piston is arranged so that the core in the radial direction comes out. 5, 10, 11, and 12 show a structure in which the piston is thinned and the inner diameter side is fixed to the front cover to prevent rotation, and the core comes out. The axial floating structure of the turbine runner of the fluid coupling in claim 4 is shown in FIGS. 4, 5 and 6, and the front cover side axial floating structure of the turbine runner of the torque converter in claim 5 is shown in FIG. 10. Shown in FIG. 11 shows a structure in which the inner peripheral sliding portion of the piston according to claim 6 is a turbine runner, and FIG. 12 shows a structure in which the boss is integrated with the front cover. FIG. 8 shows the overall positional relationship of the torque converter according to claim 7 in the automatic transmission, and FIGS. 7 and 10 show details of the relative positional relationship with the automatic transmission housing.

FIG. 1 shows a fluid coupling 1a using a hydraulic impulsion chamber of an independent lock-up clutch as an input joint of a multi-stage automatic transmission, using a pump impeller and a turbine runner. Therefore, it is a joint in which the area of the torus portion is reduced without the need to increase the torque transmission capacity and the need for a baffle plate for lowering the capacity at the low speed ratio accompanying the capacity increase in the high speed ratio region. If the fluid coupling 1a is used in a simple forward 6-speed automatic transmission having a gear range of 7 or more as shown in FIG. 2, it becomes more compact than the conventional forward 4-speed automatic transmission, and the increase in cost is reduced.

In FIG. 2 showing the structure of this forward six-speed automatic transmission, a fluid coupling 1a is disposed surrounded by a partition wall 41 between the housing 40 and the transmission on the left side of the drawing. The power of the prime mover is transmitted from the crankshaft 50 to the fluid coupling 1a via the flexible plate 51 and output to the input shaft 60 of the transmission. The main transmission gear of the transmission is a simple planetary gear train composed of a sun gear S1, a planet carrier P1, and a ring gear R1, and a simple planetary gear train composed of a sun gear S2, a planet carrier P2, and a ring gear R2, a sun gear S1 and a ring gear R2. Are connected as the first component, and the planet carrier P1 and the planet carrier P2 are connected as the second component, the ring gear R1 is the third component, and the sun gear S2 is the fourth component. The power from the input shaft 60 is directly transmitted to the planetary carriers P1 and P2 of the second component via the clutch C2, or the ring gear R0 of the simple planetary gear train for reduction to which the sun gear S0 is fixed. Is decelerated through the planet carrier P0 from the fourth component sun gear S2 and the first component sun gear S via the clutch C1 or C3. And is selectively transmitted to the ring gear R2, is output from the ring gear R1 of the third component to the counter gear 61. Here, the planetary carriers P1 and P2 as the second component are braked by the one-way clutch OWC and the brake B1, and the sun gear S1 and the ring gear R2 as the first component are selectively braked by the brake B2. The counter gear 61 meshes with the counter gear 62 disposed on the relay shaft, and power is transmitted from the counter gear 63 integrated with the relay shaft to the counter gear 64 disposed on the output shaft meshing with the counter gear 63 to transmit the differential 70. Is output via.

In FIG. 3, which schematically illustrates the structure of the forward 6-speed reverse 1-speed automatic transmission shown in FIG. 2 and describes the speed diagram and gear ratio, the gear ratio obtained by dividing the ring gear of each planetary gear train by the sun gear is ZR0 / If ZS0 = 1.667, ZR1 / ZS1 = 1.847, ZR2 / ZS2 = 1.609, and selectively engage any one of clutches C1, C2, C3 and brakes B1, B2, forward 6 speed reverse A first speed shift is possible. The characteristic of this gear ratio is that the gear ratio of one speed stage is as large as 4.754. As the variable speed stage shifts to the high speed side, the step ratio between adjacent gear ratios gradually decreases and the gear range becomes 7 .34 can be taken greatly. This is equivalent to 1.2 to 1.5 times the forward five-speed manual transmission that is normally used in similar vehicles. The input joint of the manual transmission uses a dry clutch that does not amplify the torque. If the gear range can be increased by 1.2 to 1.5 times even with an automatic transmission that cannot use the full performance of the prime mover as much as the manual transmission, the torque converter A fluid coupling with no torque amplification can be applied without using the.

Unlike the torque converter, since the fluid coupling 1a does not have a wheel stator, at a low speed ratio, the oil that has passed through the turbine runner is circulated in a direction that prevents the rotation of the pump impeller, and the driving force of the pump impeller is larger than that of the torque converter. In such a fluid conduction device, oil circulates due to the difference in centrifugal force between the pump impeller and the turbine runner, and the difference in centrifugal force increases as the speed ratio decreases, so that a large amount of oil circulates and torque transmission capacity increases. . Accordingly, the capacity at the low speed ratio should be matched to the characteristics of the prime mover, and there is no torque amplification action like a torque converter, but the area of the torus is small, and the front wheel drive forward six-speed automatic transmission It is a low cost, light weight, compact input fitting suitable for the If the area of the torus portion is reduced, the capacity at the high speed ratio is reduced, but this is not a problem because the lockup occurs.

In FIG. 1, which is an enlarged view of the fluid coupling 1a shown in the forward 6-speed backward 1-speed automatic transmission shown in FIG. 2, the crankshaft 50 of the prime mover is sandwiched between washers 53 and 54 and the starting ring gear of the prime mover. A flexible plate 51 provided with 52 is connected by a bolt 55. The flexible plate 51 is connected to the boss 17 welded to the outer periphery of the front cover 3a of the fluid coupling 1a by a bolt 56, and the prime mover and the fluid coupling 1a are connected with a certain degree of freedom in the axial direction.

The front cover 3a and the pump impeller 2a are welded at the outer peripheral portion so that the pump impeller 2a is inside, and an oil chamber is formed inside. The pump impeller 2a has a plurality of protruding portions (annular protrusions and Description) is formed and is extended in the axial direction of the front cover 3a, and an impeller hub 18 is welded to the inner periphery of the pump impeller 2a and is disposed in a boss portion of a partition wall 41 connected to a housing 40 of the transmission. And the pump rotor 65 of the charging pump of the fluid coupling 1a and the speed change device arranged in the partition wall 41 are supported. A pilot boss 19 is welded to the center of the front cover 3a in the radial direction, and the crankshaft 50 forms the center of rotation. A turbine runner 4a is arranged in the oil chamber formed by the front cover 3a and the pump impeller 2a so as to face the pump impeller 2a. The torus formed by the blades of the pump impeller 2a and the turbine runner 4a depends on the kinetic energy of the oil. Fluid conduction takes place.

The turbine runner 4a includes a torsion damper mechanism, a runner portion having blades, and a turbine hub 9a. The torsion damper mechanism and the runner portion are integrally connected to the turbine hub 9a by a rivet 16, and the turbine hub 9a is fixed to the partition wall 41. 42 is splined to an input shaft 60 of a transmission supported by a bush 44 disposed on 42. The torsion damper mechanism is composed of a clutch hub 7a and clutch plates 11 and 12 that sandwich the torsion springs 12, 13 and 14 in the rotational direction. The clutch hub 7a has friction members attached to both sides of the outer peripheral portion. 12 is integrally connected by a stud pin 13 and sandwiches the clutch hub 7a between both sides to restrict axial movement and holds the torsion springs 12, 13 and 14 and is connected to the turbine hub 9a. This torsion damper mechanism has a multistage vibration absorbing function by the torsion springs 12, 13 and 14, and a vibration damping function by a frictional force at the contact surface between the clutch plates 11 and 12 and the clutch hub 7a.

The plurality of annular projections provided at the outer peripheral tip of the pump impeller 2a are processed to have a stepped shape so that the inner diameter of the tip portion is increased. An end plate 8 that is in contact with the pump impeller 2a and is restricted from moving to the pump impeller 2a side is disposed, and the end plate 8 and the clutch hub 7a adjacent to the end plate 8 are sandwiched to fit with a plurality of annular protrusions of the pump impeller 2a to prevent rotation. And a piston 6a that is movable in the axial direction. The outer periphery of the piston 6a is centered along the inner diameter of the front cover 3a or the annular protrusion of the pump impeller 2a, and a seal 100 is welded to the outer periphery of the annular protrusion formed at the center in the radial direction. The boss 17 welded to the outer peripheral portion is slidable in the axial direction while being non-rotatable at a constant interval in the radial direction between the inner periphery of the annular recess formed on the inner side in the radial direction, and on the radially inner side. A seal 101 is welded to the inner periphery of the hole, and can slide relative to the outer periphery of the distal end of the input shaft 60 of the transmission with a constant spacing in the radial direction and slide with a movement allowance of about 1 mm in the axial direction. Is done. The turbine runner 4a is interposed between the end plate 8 and the piston 6a so as to float in the movement margin of about 1 mm in the axial direction of the piston 6a, and is not provided with a thrust bearing.

1 and FIG. 4 showing the control and the upper part of the rotation center, the oil chamber surrounded by the front cover 3a and the pump impeller 2a is located on the pump impeller 2a side by the outer peripheral seal 100 and the inner peripheral seal 101 of the piston 6a. The oil chamber A and the oil chamber B on the front cover 3a side are separately formed. The oil chamber A is composed of an oil chamber A1 on the pump impeller 2a side and an oil chamber A2 on the piston 6a side with the turbine runner 4a interposed therebetween. The oil chamber A1 is in fluid communication between the impeller hub 18 and the support 42. The oil chamber B communicates with the outside of the fluid coupling 1a through a hole opened at the center of the input shaft 60 at the tip end B1 of the input shaft 60.

The lock-up control valve 80a disposed outside the fluid coupling 1a is provided with a spool 81 and a return spring 82. In the fluid conduction state of the lock-up off, the spool 81 returns because the lock-up regulator pressure for controlling the lock-up is zero. The spring 82 is pushed to the left end of the figure. In this fluid conduction state, the internal oil circulates in the torus by both impellers due to the difference in centrifugal force between the pump impeller 2a and the turbine runner 4a, and the heat generated by the loss is cooled, so usually 5 to 10 liters per minute. The oil is poured into the torus from outside and the oil inside is discharged to the outside. The inside of the torus is kept at a low pressure to avoid cavitation, and the oil from the coupling regulator with a constant pressure passes through the lock-up control valve 80a and flows into the oil chamber A1 of the fluid coupling 1a, and the end plate 8 and the piston 6a from the outer periphery of the torus. It flows between the friction surfaces of the clutch hub 7a sandwiched between them and flows to the outer peripheral seal 100 of the piston 6a or flows to the inner peripheral seal 101. In this state, since the pressure in the oil chamber A is higher than that in the oil chamber B, the piston 6a is pushed toward the front cover 3a and has a movement margin of about 1 mm in the axial direction. The seals 100 and 101 welded to the piston 6a are elastically deformed in the axial direction due to this pressure difference to form a gap in the radial direction, and the oil flows from the oil chamber A to the oil chamber B and from the B1 portion to the center of the input shaft 60. It flows through the opened hole and flows from the lock-up control valve 80a toward the cooler. Details of the elastic deformation of the seals 100 and 101 will be described later with reference to FIG.

In this fluid conduction state, the turbine runner 4a receives a force toward the front cover 3a due to the circulation of oil inside the torus. However, the pump impeller 2a by centrifugal hydraulic pressure generated in the oil chamber A2 between the turbine runner 4a and the piston 6a. Since the force to the side is larger, it is pushed to the pump impeller 2a side. At this time, the friction surface of the clutch hub 7a of the turbine runner 4a pushes the end plate 8, where torque due to friction force is generated, but there is no influence on the performance of the fluid coupling without torque amplification. Normally, the turbine runner 4a is less likely to receive a force toward the front cover 3a. In this case, the turbine runner 4a is restricted from moving toward the front cover 3a with respect to the clutch hub 7a. No bearing is required.

Next, in the lockup-on lockup state shown in the lower part of FIG. 4, the lockup regulator pressure that controls the lockup pushes the spool 81 against the right end of the figure against the force of the return spring 82. In this state, the oil from the coupling regulator flows through the lockup control valve 80a directly to the cooler, and the oil inside the torus is drained from the portion A1 through the lockup control valve 80a. Further, the lockup regulator pressure flows into the oil chamber B from the B1 portion. In this state, since the pressure in the oil chamber B is higher than that in the oil chamber A, the seals 100 and 101 welded to the piston 6a are pushed in the radial direction so as to block the sliding portions, and the oil chamber B and the oil chamber A are separated. Is pressed to the pump impeller 2a side. The piston 6a presses the friction plate of the clutch hub 7a and the end plate 8 whose movement to the pump impeller 2a side is restricted. In this state, the power from the prime mover is transmitted from the end plate 8 that is non-rotatably connected to the front cover 3a and the piston 6a through the friction surface of the clutch hub 7a and directly from the turbine runner 4a to the input shaft 60 of the transmission. It becomes a state. Since this lock-up regulator pressure acts on the piston 6a as it is, accurate mechanical conduction is possible by controlling the lock-up regulator pressure. Details of the seals 100 and 101 in this state will be described later with reference to FIG.

In FIG. 5 showing an embodiment in which the seal of the piston 6a welded with the seals 100 and 101 in FIG. 4 is detachable, the driven plate 8b contacting the friction surface of the clutch hub 7a is the tip of the outer periphery of the pump impeller 2a. The piston 6b is fitted to the plurality of annular protrusions provided on the front cover 3a so as to be prevented from rotating while being axially movable and presses the driven plate 8b. The piston 6b engages with the plurality of annular protrusions of the pump impeller 2a. In addition, they are prevented from rotating and are arranged so as to be movable in the axial direction. The piston 6b is centered along the inner diameter of the annular protrusion of the front cover 3a or the pump impeller 2a, and a ring 27 that detachably holds the seal 102 is formed on the annular protrusion formed at the center in the radial direction. It was welded and slid in the axial direction while maintaining a certain distance in the radial direction between the inner periphery of the annular recess of the front cover 3a, and it was slid in the axial direction, and was welded to a thin metal material on the inner periphery of the hole in the radial direction. The seal 103 is press-fitted and is slidable relative to the distal end outer peripheral portion of the input shaft 60 of the transmission with a constant clearance in the radial direction and having a movement allowance of about 1 mm in the axial direction. The shapes of the seals 102 and 103 on the sliding surface are the same as the seals 100 and 101 of FIG. 4, and FIG. 5 shows an embodiment of the fluid coupling 1a. Is also applicable.

In FIG. 6, which shows another embodiment of the piston 6a in FIG. 4, a driven plate 8b that contacts the friction surface of the clutch hub 7a is fitted with a plurality of annular protrusions provided at the outer peripheral tip of the pump impeller 2a. Thus, it is prevented from rotating and is arranged on the front cover 3a side so as to be movable in the axial direction. The piston 6c is thinned and fixed by being squeezed by a plurality of protrusions protruding from the inner peripheral side surface of the front cover 3b in a state where the outer peripheral portion can be flexed flexibly in the axial direction. The seal 104 welded to the outermost peripheral portion of the piston 6c in the radial direction slides in the axial direction at a constant radial distance between the pump impeller 2a and the inner peripheral portion of the front cover 3a to be welded, and is a driven plate. The seal 105, which presses 8b and is welded to the inner peripheral portion, is fixed in the axial direction between the distal end outer peripheral portion of the input shaft 60 of the transmission and performs relative rotation with a constant spacing in the radial direction. The shapes of the seals 104 and 105 are the same as those of the seals 100 and 101 of FIG. 4, and FIG. 5 shows an embodiment in the fluid coupling 1a, but it can also be applied to a torque converter 1b described later.

FIG. 7 shows a torque converter 1b using a simple and compact pump impeller, a turbine runner, and a wheel stator provided with an independent lock-up clutch hydraulic engagement chamber as an input joint of a multi-stage automatic transmission, as shown in FIG. When used in a simple forward 6-speed automatic transmission with such a cross gear ratio, the compactness is equal to or less than that of a conventional forward 4-speed automatic transmission, and the cost increase is reduced.

In FIG. 8 showing the structure of this forward six-speed automatic transmission, a torque converter 1b is disposed surrounded by a partition wall 41 between the housing 40 and the transmission on the left side of the drawing. The power of the prime mover is transmitted from the crankshaft 50 to the torque converter 1b via the flexible plate 51 and output to the input shaft 60 of the transmission. The main transmission gear of the transmission is a simple planetary gear train composed of a sun gear S1, a planet carrier P1, and a ring gear R1, and a simple planetary gear train composed of a sun gear S2, a planet carrier P2, and a ring gear R2, a sun gear S1 and a ring gear R2. Are connected as the first component, and the planet carrier P1 and the planet carrier P2 are connected as the second component, the ring gear R1 is the third component, and the sun gear S2 is the fourth component. The power from the input shaft 60 is transmitted directly to the fourth component sun gear S2 and the second component planetary carriers P1, P2 via the clutch C1 or C2, or the sun gear S0 is fixed. The sun gear S1 of the first component is decelerated from the planet carrier P0 of the double planetary gear train for reduction through the ring gear R0 through the clutch C3. Selectively transmitted to the micro-ring gear R2, is output from the ring gear R1 of the third component to the counter gear 61. Here, the planetary carriers P1 and P2 as the second component are braked by the one-way clutch OWC and the brake B1, and the sun gear S1 and the ring gear R2 as the first component are selectively braked by the brake B2. The counter gear 61 meshes with the counter gear 62 disposed on the relay shaft, and power is transmitted from the counter gear 63 integrated with the relay shaft to the counter gear 64 disposed on the output shaft meshing with the counter gear 63 to transmit the differential 70. Is output via.

In FIG. 9, which schematically illustrates the structure of the forward 6-speed reverse 1-speed automatic transmission shown in FIG. 8 and describes the speed diagram and gear ratio, the gear ratio obtained by dividing the ring gear of each planetary gear train by the sun gear is ZR0 / If ZS0 = 1.938, ZR1 / ZS1 = 1.8, ZR2 / ZS2 = 2.07, and selectively engage any one of clutches C1, C2, C3 and brakes B1 and B2, 6 forward speeds A first speed shift is possible. The characteristic of this gear ratio is that the step ratio between adjacent gear ratios gradually decreases as the variable speed step shifts to the high speed side, and the gear range is 5.8, so that the step ratio can be reduced. 2 and FIG. 3, the speed is reduced to the planetary gear train when there is a gear ratio 1 in which the power does not pass through the shifting planetary gear at all at the forward 4 speed stage as compared with the forward 6-speed reverse 1-speed automatic transmission described above. In the forward 3 and 5 speed stages that pass through, only a part of them passes, which is advantageous in terms of gear transmission efficiency and strength. However, since the gear range is not so different from that of a manual transmission device with a gear range of 5.8 and five forward speeds, a torque converter is used to make the traction characteristics advantageous.

Unlike the torque converter of the conventional forward 4-speed reverse 1-speed automatic transmission, the torque converter 1b can reduce the torque ratio, so that the input torque capacity at the low speed ratio increases. In such a fluid conduction device, oil circulates due to the difference in centrifugal force between the pump impeller and the turbine runner, and the difference in centrifugal force increases as the speed ratio decreases, so that a large amount of oil circulates and torque transmission capacity increases. . Therefore, it is only necessary to match the capacity at the low speed ratio with the characteristics of the prime mover, and the capacity at the high speed ratio is low, but this is not a problem because it locks up, and the area of the torus part can be reduced. it can. In the torque converter 1b, the arrangement of the one-way clutch 5 is arranged on the radially outer side of the boss portion in order to avoid interference with the boss portion of the partition wall 41, the torus inner diameter is increased correspondingly, and the area of the torus portion is reduced. It is a light weight and compact input joint suitable for a 6-speed forward drive automatic transmission.

In FIG. 7, which is an enlarged view of the torque converter 1b shown in the forward 6-speed backward 1-speed automatic transmission shown in FIG. 8, the crankshaft 50 of the prime mover is sandwiched between washers 53 and 54 and the starting ring gear of the prime mover. A flexible plate 51 provided with 52 is connected by a bolt 55. The flexible plate 51 is connected to the boss 17 welded to the outer periphery of the front cover 3a of the torque converter 1b by a bolt 56, and the prime mover and the torque converter 1b are connected with a certain degree of freedom in the axial direction.

The front cover 3a and the pump impeller 2b are welded at the outer peripheral portion so that the pump impeller 2b is inside, and an oil chamber is formed inside, and a plurality of annular protrusions are formed at the outer peripheral tip of the pump impeller 2b, and the axial direction of the front cover 3a In addition, the impeller hub 18 is welded to the inner periphery of the pump impeller 2b and is pivotally supported by a bush 43 disposed on a boss portion of a partition wall 41 that is connected to the housing 40 of the transmission, and is also disposed on the partition wall 41. And the pump rotor 65 of the charging pump of the torque converter 1b. A pilot boss 19 is welded to the center of the front cover 3a in the radial direction, and the crankshaft 50 forms the center of rotation. A turbine runner 4b is disposed in the oil chamber formed by the front cover 3a and the pump impeller 2b so as to face the pump impeller 2b, and a wheel stator 5 is disposed between the pump impeller 2b and the turbine runner 4b. Fluid conduction is performed by oil kinetic energy in a torus portion formed by the impeller 2b, the turbine runner 4b, and the blades of the wheel stator 5.

A one-way clutch 21 that allows the wheel stator 5 to freely rotate in the rotational direction of the pump impeller 2b and not to rotate in the reverse direction is provided on the radially inner side of the wheel stator 5 and an outer race 22 on the outer circumferential side and an inner race 20 on the inner circumferential side. It is placed between. The outer race 22 and the inner race 20 are concentric at the side portion of the side washer 23 disposed on the turbine runner 4b side of the one-way clutch 21 and the wheel stator 5 disposed on the pump impeller 2b side. The case 42 is splined to a support 42 integral with the case 40. Here, the one-way clutch 21 is disposed at a position where the inner diameter r2 is larger than the outer diameter r1 of the housing boss portion of the partition wall 41 that supports the impeller hub 18, and a thrust needle roller serving as a thrust bearing between the side washer 23 and the turbine runner 4b. The bearing 24 is disposed at a position smaller than r2, and a thrust needle roller bearing 25 serving as a thrust bearing between the side portion of the wheel stator 5 and the pump impeller 2b is disposed at a position larger than r2.

What is required of the input joint to make the automatic transmission compact is not only to flatten the torus of the torque converter, but also to reduce the interference width between the mounting bolt 55 to the motor on the inner peripheral side and the housing boss portion of the partition wall 41. Must be small. In particular, in this structure in which a hydraulic engagement chamber of an independent lock-up clutch is provided, the friction member is arranged in the outer peripheral space of the turbine runner 4b and the front cover 3a, and the torsion damper mechanism is arranged on the inner diameter side of the friction member, The connecting portion of the torsion damper mechanism, the one-way clutch 21, and the thrust needle roller bearings 24 and 25 serving as thrust bearings are arranged so as to overlap the radial positions of the bolt 55 and the housing boss portion of the partition wall 41. Therefore, the inner diameter of the torus is increased, the one-way clutch 21 and the thrust needle roller bearing 25 are arranged at a position larger than the outer diameter r1 of the boss part of the partition wall 41, and only the thrust needle roller bearing 24 is arranged so as to overlap the boss part of the partition wall 41. By doing so, the automatic transmission can be made compact. In addition, since the area of the torus portion can be reduced, it is not necessary to increase the outer diameter of the torus, and the weight of the automatic transmission can be prevented from being increased.

The turbine runner 4b includes a torsion damper mechanism, a runner portion having blades, and a turbine hub 9b. The torsion damper mechanism and the runner portion are integrally connected to the turbine hub 9b by a rivet 16, and the turbine hub 9b is fixed to the partition wall 41. 42 is splined to an input shaft 60 of a transmission supported by a bush 44 disposed on 42. The torsion damper mechanism is composed of a clutch hub 7c that sandwiches torsion springs 12, 13 and 14 in the rotational direction and clutch plates 11 and 12, and the clutch hub 7c is bent in an inverted L shape at the outer peripheral portion and splined. The plates 11 and 12 are integrally connected by a stud pin 13 so as to restrict the movement in the axial direction with the clutch hub 7c sandwiched between both sides, and hold the torsion springs 12, 13 and 14 and are connected to the turbine hub 9b. This torsion damper mechanism has a multistage vibration absorbing function by the torsion springs 12, 13 and 14, and a vibration damping function by a frictional force at the contact surface between the clutch plates 11 and 12 and the clutch hub 7c. A drive plate 7b having friction members attached to both sides is splined to the outer peripheral spline of the clutch hub 7c, and the clutch hub 7c is connected to the drive plate 7b by a protrusion 28 provided on the pump impeller 2b side of the outer peripheral spline of the clutch hub 7c. On the other hand, movement to the front cover 3a side is restricted and movement to the pump impeller 2b side is possible.

The plurality of annular protrusions provided at the outer peripheral tip of the pump impeller 2b are processed to be stepped so that the inner diameter of the tip is increased. An end plate 8 which is in contact with the pump impeller 2b and is restricted from moving to the pump impeller 2b side is arranged, and the end plate 8 and a drive plate 7b adjacent to the end plate 8 are sandwiched to fit with a plurality of annular protrusions of the pump impeller 2b to prevent rotation. And a piston 6a that is movable in the axial direction. The outer periphery of the piston 6a is centered along the inner diameter of the annular protrusion of the front cover 3a or the pump impeller 2b, and a seal 100 is welded to the outer periphery of the annular protrusion formed at the radial center. The boss 17 welded to the outer peripheral portion is slidable in the axial direction while being non-rotatable at a constant interval in the radial direction between the inner periphery of the annular recess formed on the inner side in the radial direction, and on the radially inner side. A seal 101 is welded to the inner periphery of the hole, and can slide relative to the outer periphery of the distal end of the input shaft 60 of the transmission with a constant spacing in the radial direction and slide with a movement allowance of about 1 mm in the axial direction. Is done. The plurality of annular protrusions of the pump impeller 2b, the front cover 3a, and the piston 6a are the same as the fluid coupling 1a shown in FIG. The turbine runner 4b is sandwiched between the end plate 8 and the piston 6a, and the front cover 3a is floated on the movement of the piston 6a about 1 mm in the axial direction, and a thrust bearing on the front cover 3a side is not provided.

7 and FIG. 10 showing the control center upper part and control, the oil chamber surrounded by the front cover 3a and the pump impeller 2b is on the pump impeller 2b side by the outer peripheral seal 100 and the inner peripheral seal 101 of the piston 6a. The oil chamber A and the oil chamber B on the front cover 3a side are separately formed. The oil chamber A includes an oil chamber A1 between the wheel stator 5 and the pump impeller 2b, an oil chamber A2 between the turbine runner 4b and the front cover 3a, and an oil chamber A3 between the turbine runner 4b and the wheel stator 5. The oil chamber A1 is communicated to the outside of the torque converter 1b between the impeller hub 18 and the support 42, and the oil chamber B is a hole formed in the center of the input shaft 60 at the tip B1 of the input shaft 60. Communicated with In FIG. 10, the oil chamber A3 is also communicated between the support 42 and the input shaft 60 to the outside of the torque converter 1b. This is a special case provided because it is necessary to cool the heated oil in the oil chamber A even during the up operation, and normally a communication circuit to the outside of the oil chamber A3 is not required.

A spool 81 and a return spring 82 are arranged on the lock-up control valve 80b arranged outside the torque converter 1b, and the spool 81 returns because the lock-up regulator pressure for controlling the lock-up is zero in the fluid conduction state of the lock-up off. The spring 82 is pushed to the left end of the figure. In this fluid conduction state, the internal oil circulates in the torus due to the difference in centrifugal force between the pump impeller 2b and the turbine runner 4b, and normally 5 to 10 liters of oil per minute is externally supplied to cool the heat generated by the loss. Then, it flows into the torus and drains the oil inside. The inside of the torus is kept at a low pressure to avoid cavitation, and oil from the coupling regulator with a constant pressure passes through the lock-up control valve 80b and flows into the oil chamber A1 portion of the torque converter 1b, and from the outer periphery of the torus, the end plate 8 and the piston 6a It flows between the friction surfaces of the drive plate 7b sandwiched between the two and flows to the outer peripheral seal 100 of the piston 6a or to the inner peripheral seal 101. In this state, since the pressure in the oil chamber A is higher than that in the oil chamber B, the piston 6a is pushed toward the front cover 3a and has a movement margin of about 1 mm in the axial direction. The seals 100 and 101 welded to the piston 6a are elastically deformed in the axial direction due to this pressure difference to form a gap in the radial direction, and the oil flows from the oil chamber A to the oil chamber B and from the B1 portion to the center of the input shaft 60. It flows through the opened hole and flows from the lock-up control valve 80b toward the cooler. Details of the elastic deformation of the seals 100 and 101 will be described later with reference to FIG.

In this fluid conduction state, the turbine runner 4b receives a force toward the front cover 3a due to the circulation of oil inside the torus. However, the pump impeller 2b by centrifugal hydraulic pressure generated in the oil chamber A2 between the turbine runner 4b and the piston 6a. Since the force to the side is larger, it is pushed to the wheel stator 5 side. Since the turbine runner 4b is not restricted by the movement of the pump impeller 2b with respect to the drive plate 7b, thrust needle roller bearings 24, 25 are provided between the turbine runner 4b and the wheel stator 5 and between the wheel stator 5 and the pump impeller 2b. It receives a thrust force toward the pump impeller 2b. Normally, the turbine runner 4b receives little force toward the front cover 3a. In this case, the turbine runner 4b is restricted from moving toward the front cover 3a with respect to the drive plate 7b. No bearing is required.

Next, in the lock-up state shown in the lower part of FIG. 10, the lock-up regulator pressure that controls the lock-up pushes the spool 81 against the right end of the figure against the force of the return spring 82. In this state, the oil from the converter regulator flows directly to the cooler through the lock-up control valve 80b, and the oil in the torus is drained from the A1 portion through the lock-up control valve 80b. Further, the lockup regulator pressure flows into the oil chamber B from the B1 portion. In FIG. 10, the oil from the converter regulator is poured from the oil chamber A3 through the orifice 83, and the oil in the oil chamber A is drained from the portion A1, but this circuit is used in a special way with large heat generation. It is not necessary to use this circuit. In this state, since the pressure in the oil chamber B is higher than that in the oil chamber A, the seals 100 and 101 welded to the piston 6a are pushed in the radial direction so as to block the sliding portions, and the oil chamber B and the oil chamber A are separated. Is pressed to the pump impeller 2b side. The piston 6a presses the friction surface of the drive plate 7b and the end plate 8 that is restricted from moving toward the pump impeller 2b. In this state, the power from the prime mover is transmitted from the turbine runner 4b to the input shaft 60 of the transmission directly from the end plate 8 and the piston 6a that are non-rotatably connected to the front cover 3a through the friction surface of the drive plate 7b. It becomes a state. Since this lock-up regulator pressure acts on the piston 6a as it is, accurate mechanical conduction is possible by controlling the lock-up regulator pressure. Details of the seals 100 and 101 in this state will be described later with reference to FIG.

The protrusions 28 on the outer peripheral spline of the clutch hub 7c in FIG. 10 are eliminated, and a thrust bearing 29 is provided between the turbine hub 9c and the front cover 3a, and a seal 106 is welded to the inner periphery of the piston 6d to contact the turbine hub 9c. FIG. 11 shows the result. In FIG. 11, the turbine runner 4b and the wheel stator 5 are stably arranged between the front cover 3a and the pump impeller 2b, with axial bearings being supported by thrust bearings 29 and thrust needle roller bearings 24 and 25. The turbine hub 9c and the input shaft 60 are sealed with an O-ring 108. A plurality of annular projections provided at the outer peripheral tip of the pump impeller 2b, the outer periphery of the end plate 8, the drive plate 7b, and the piston 6d to which the seal 100 is welded are the same as the torque converter 1b of FIG. Therefore, the seal 106 welded to the inner periphery of the piston 6d can be relatively rotated between the boss outer peripheral portion of the front cover 3a side of the turbine hub 9c with a certain distance in the radial direction, and can move about 1 mm in the axial direction. And is slid. Although FIG. 11 is an example applied to the torque converter 1b, it can also be applied to the fluid coupling 1a described above.

Further, the protrusion 28 of the outer peripheral spline of the clutch hub 7c in FIG. 10 is eliminated, the boss 30 is welded to the front cover 3a, a thrust bearing 31 is provided between the boss 30 and the turbine hub 9d, and a seal is provided on the inner periphery of the piston 6e. FIG. 12 shows 107 welded and brought into contact with the boss 30. In FIG. 12, the turbine runner 4b and the wheel stator 5 are supported by a thrust bearing 31 and thrust needle roller bearings 24 and 25 between a boss 30 integrated with the front cover 3a and a pump impeller 2b, and are stabilized. Arranged. A space between the boss 30 and the input shaft 60 is rotatably sealed by a seal ring 109. A plurality of annular projections provided at the outer peripheral tip of the pump impeller 2b, the outer periphery of the end plate 8, the drive plate 7b, and the piston 6e to which the seal 100 is welded are the same as the torque converter 1b of FIG. Accordingly, the seal 107 welded to the inner periphery of the piston 6e is slid with a movement allowance of about 1 mm in the axial direction while maintaining a constant radial distance from the outer periphery of the boss 30. Although FIG. 12 is an example applied to the torque converter 1b, it can also be applied to the fluid coupling 1a described above.

In FIG. 13 which shows the connection part of the friction member of the lockup clutch used for the torque converter 1b of FIG.7 and FIG.10, XX cross section sees through the friction member from the side surface side of the front cover 3a, Arrow view Y is a perspective view of the friction member from the top of the front cover 3a. A plurality of annular projections provided at the outer peripheral tip of the pump impeller 2b are equally divided in the circumferential direction as shown by the XX cross section and the arrow Y, and are projected toward the front cover 3a in parallel to the axial direction. The end plate 8 and the radial projection of the piston 6a are fitted into the circumferential opening to prevent rotation. The inner diameter of the annular protrusion is processed so as to be stepped on the side surface side of the front cover 3a at a position slightly closer to the side surface of the front cover 3a than the opening, and the end plate 8 has the stepped portion. The movement toward the pump impeller 2b is restricted. The outer diameter of the annular protrusion of the pump impeller 2b and the inner diameter of the front cover 3a are substantially the same diameter, and the front cover 3a prevents deformation of the annular protrusion due to centrifugal force. The piston 6a is processed so that the outer diameter of the protrusion fitted into the circumferential opening is along the inner diameter of the front cover 3a or the outer diameter of the recess is along the inner diameter of the annular protrusion of the front cover 3a. The core is aligned in the radial direction, is prevented from rotating by the protrusion of the pump impeller 2b, and is arranged so as to be movable in the axial direction. The centering of the piston 6a in the radial direction is to provide a certain radial elastic margin for the seals 100 and 101 on the sliding surface. The drive plate 7b is rotatably disposed between the piston 6a and the end plate 8, and a friction material having a lubricating groove formed on both sides is attached. The annular protrusion of the pump impeller 2b has a shape common to the fluid coupling 1a and the torque converter 1b serving as the input coupling of the present invention, and the friction member is connected in the same manner.

In FIG. 14 schematically showing the non-return action of oil from the oil chamber A to the oil chamber B in the sliding portion of the seal member welded to the piston 6a, the oil pressure in the oil chamber B is higher than that in the oil chamber A. In the increased lock-up on state, the outer peripheral side seal 100 slides because the force acting from the inner peripheral side hydraulic pressure of the oil chamber B to the outer peripheral side is larger than the force acting from the outer peripheral side hydraulic pressure of the oil chamber A to the inner peripheral side. The oil chamber B and the oil chamber A are separated from each other by being pressed against the front cover 3a, and the inner peripheral side seal 101 has an outer peripheral side hydraulic pressure of the oil chamber B by a force acting from the inner peripheral side hydraulic pressure of the oil chamber A to the outer peripheral side. Therefore, the oil chamber B and the oil chamber A are separated by being pressed against the input shaft 60 serving as a sliding surface. Further, in the lock-up-off state in which the oil pressure in the oil chamber A is higher than that in the oil chamber B, the outer peripheral side seal 100 is changed from the outer peripheral pressure in the oil chamber A to the inner periphery by the force acting from the inner peripheral pressure in the oil chamber B to the outer peripheral side. Since the force acting on the side is larger, a gap is formed between the front cover 3a which is elastically deformed on the inner peripheral side and becomes a sliding surface, and the oil flows from the oil chamber A to the oil chamber B. 101 is larger in force acting from the inner peripheral side hydraulic pressure of the oil chamber A to the outer peripheral side than the force acting from the outer peripheral side hydraulic pressure of the oil chamber B to the outer peripheral side. A gap is formed between the oil chamber A and the oil chamber B. FIG. 14 illustrates the seals 100 and 101, but the same operation is performed for the seals 102 and 103 where the seal member is not welded to the piston and the seals 104 and 105 where the seal member is welded to the thin piston.

Enlarged structural diagram of the mounting part of the automatic transmission equipped with the fluid coupling of the present invention Overall structure diagram of an automatic transmission equipped with the fluid coupling of the present invention Schematic diagram, speed diagram, and gear ratio of the gear train of the automatic transmission shown in FIG. Fig. 1 Upper radius structure and external control diagram Another embodiment in which the piston seal member of FIG. 1 is detachable Another embodiment in which the piston mounting method of FIG. 1 is changed FIG. 3 is an enlarged structural view of a mounting portion of an automatic transmission equipped with the torque converter of the present invention Overall structure of automatic transmission equipped with torque converter of the present invention Schematic diagram, speed diagram, and gear ratio of the gear train of the automatic transmission shown in FIG. Structure of the upper radius of FIG. Another embodiment in which the inner periphery sliding portion of the piston of FIG. 7 is a turbine hub Another embodiment in which the inner periphery sliding portion of the piston in FIG. 7 is a front cover boss Detailed view showing the connecting portion of the friction member of the present invention Detailed view showing the operation of the sealing member of the piston of the present invention

Explanation of symbols

1a Fluid coupling 1b Torque converter 2a, 2b Pump impeller 3a, 3b Front cover 4a, 4b Turbine runner 5 Wheel stators 6a, 6b, 6c, 6d, 6e Pistons 12, 13, 14 Torsion spring 40 Transmission case 50 Crankshaft 60 Input shafts 80a and 80b of transmission device Lock-up control valves A, A1, A2, A3, B, B1 Oil chamber

Claims (7)

Input member having an oil chamber surrounded integrally by the front cover and a pump impeller, a turbine runner for connecting the input shaft of the arranged is transmission between the front cover to face the pump impeller, or they In addition, a fluid conduction portion comprising a wheel stator disposed between the pump impeller and the turbine runner,
A friction member that connects the input member and the turbine runner, a torsion damper mechanism that is provided between the friction member and the turbine runner, and is axially movable adjacent to the front cover and is held in a relatively non-rotatable manner. A lockup clutch comprising a piston that separates and forms the oil chamber into an oil chamber B on the front cover side and an oil chamber A on the pump impeller side, and presses and engages the friction member with the hydraulic pressure supplied to the oil chamber B. An input joint of an automatic transmission equipped with a mechanical conduction part according to
The oil chamber B is hermetically sealed at the inner and outer periphery sliding portions that separate the oil chambers A and B of the piston in a state in which the pressure of the oil chamber B is higher than the pressure of the oil chamber A, and the oil chamber A from the pressure of the oil chamber B. Using an elastic seal member that communicates the oil chambers A and B with a high pressure of
Oil chambers A and B are connected to the outside of the input joint, and hydraulic pressure is supplied to the oil chamber B in a conductive state where the lock-up clutch is engaged and the oil in the oil chamber A is supplied to the outside of the input joint. An input joint of an automatic transmission that discharges and supplies hydraulic pressure to the oil chamber A in a fluid conduction state in which the lock-up clutch is released and discharges the oil in the oil chamber A through the oil chamber B to the outside of the input joint. . The seal member of the outer periphery sliding part that separates and forms the oil chambers A and B of the piston has a shape in which the oil pressure of the oil chamber A acts on the outer peripheral side and the oil pressure of the oil chamber B acts on the inner peripheral side. The seal member of the sliding portion has a shape in which the oil pressure in the oil chamber A acts on the inner peripheral side and the oil pressure in the oil chamber B acts on the outer peripheral side, and can be elastically deformed in the radial direction by the differential pressure between the oil chambers A and B. The input joint for an automatic transmission according to claim 1, wherein the input joint is welded to the piston or mounted independently. The input member is provided with a radial step to the tip end direction of the inner diameter is increased, the pump impeller outer peripheral portion of the front cover an outer peripheral portion having a plurality of projections extending in the axial direction on the tip side of the該径direction step An end plate which is arranged and welded radially inside, is fitted with the protrusion and is prevented from rotating, and is in contact with the radial step and is restricted in axial movement toward the pump impeller side; and Friction members connected to the turbine runner via a torsion damper mechanism, or in addition to them, a driven plate that is fitted with the protrusion and is prevented from rotating and is movable in the axial direction in order from the pump impeller side. arranged, axially movably along the inner diameter of the protrusion or the front cover the outer periphery as well as prevent rotation of the piston engaged with the protrusion The driven plate or the friction member is arranged by arranging the piston plate to be thin and fixing the inner diameter side to the front cover and preventing rotation and elastically deforming the outer diameter side to move axially. The input joint for an automatic transmission according to claim 1, wherein the input joint is pressed. The fluid conducting portion is a fluid coupling composed of a pump impeller and a turbine runner, and is a friction member integrally provided on the outer periphery of a clutch hub that is connected to the turbine runner via a torsion damper mechanism so as to be swingable in the rotational direction. The input joint of the automatic transmission according to claim 3, wherein: is arranged between the end plate whose axial direction is regulated and the piston. The fluid conducting portion is a torque converter including a pump impeller, a turbine runner, and a wheel stator, and an outer periphery of a clutch hub connected to the turbine runner via a torsion damper mechanism is swingable to the pump impeller side. A cylindrical portion is bent into an L shape, and the end plate and a drive plate sandwiched between the pistons are engaged with the cylindrical portion by a spline, and the friction member is provided on the drive plate . 4. The automatic transmission according to claim 3, wherein a protrusion is provided at a pump impeller side end portion of the spline engaged with the drive plate , and movement of the clutch hub in the front cover direction is restricted by the drive plate. Input fittings. Wherein the fluid-conducting part is a fluid coupling or torque converter comprising a pump impeller and the turbine runner and the wheel stator, and a pump impeller and the turbine runner, the outer periphery of the piston slides and the front cover, the inner periphery the turbine runner and or slides integral with the boss on the front cover, the inner circumference sealing the input shaft of the boss integral to the turbine runner or the front cover to slide the automatic transmission in the sealing member, A clutch hub coupled to the turbine runner via a torsion damper mechanism and the friction member can be moved relative to each other in the axial direction, and between the front cover, the turbine runner, and the pump impeller , or between the front cover and the turbine runner. the wheel stator and the pump impeller The input coupling of the automatic transmission according to claim 3 which arranged a thrust bearing between the members. The fluid conducting portion is a torque converter including a pump impeller, a turbine runner, and a wheel stator, and an inner diameter r2 of the one-way clutch of the wheel stator is determined from an outer diameter r1 of a housing boss portion serving as a transmission side shaft support portion of the torque converter. The one-way clutch is arranged on the radially outer side of the housing boss portion so as to be larger, and the thrust bearing between the turbine runner and the wheel stator is arranged on the radially inner side so as to be smaller than the inner diameter r2 of the one-way clutch, The input joint of the automatic transmission according to claim 6 , wherein a thrust bearing between the stator and the pump impeller is arranged radially outward so as to be larger than an inner diameter r2 of the one-way clutch.

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