JP6689504B2 - Multi-speed transmission - Google Patents

Description

  The present invention relates to an automatic transmission (AT) that is a vehicular automatic transmission that controls a planetary gear using a hydraulic clutch and a brake, and particularly relates to a ninth forward speed with good gear meshing efficiency and a multi-speed transmission that exceeds that speed. Regarding the structure.

  The most important factor for a multi-speed transmission using a plurality of planetary gear trains and clutch and brake engagement elements is the gear ratio, and in a typical vehicle, the step value of the gear ratio from the lowest speed to the next is 1. It is desirable that the step value of the gear ratio from the next stage to the highest stage is around 60, which is around 1.15, that is, the step value gradually decreases as it shifts to the higher stage. As a result, the gear ratio range (Gear Range) obtained by dividing the gear ratio of the lowest speed by the gear ratio of the highest speed is 8 for the seventh forward speed, 9 for the eighth forward speed, and 9 for the forward speed. If so, 10 is appropriate, and if it is the 10th forward speed, about 11 is appropriate.

  The existing MT (Manual Transmission) has a structure in which a dedicated counter gear meshes with each gear, so that a free gear ratio can be obtained. Therefore, the degree of freedom of the gear ratio is reduced. Therefore, a certain number of planetary gear trains and fastening elements are required to set an appropriate gear ratio.

  Looking back on history, the 3AT, which was popular for passenger cars in the 1960s, had three planetary gear trains and four fastening elements (two clutches and two brakes) in the 3rd forward 3rd reverse 1AT. In the 1980s, 4AT, which was increased in number to 1 by increasing the number of fastening elements (clutch), became popular in order to improve the characteristics. However, there was a problem that the step value from the third forward speed to the fourth speed was too large. On the other hand, in trucks and buses where traction is the most important factor, the number of planetary gear trains is increased by 1 to 3 and four ATs are used, with 5 fastening elements (2 clutches and 3 brakes) to obtain an ideal gear ratio. In the 1970's, the 6AT, which is equivalent to the 4AT, can achieve an appropriate gear ratio with three planetary gear trains and five fastening elements (two clutches and three brakes). It is a GM (Allison) 6AT (A type) devised in 52-149562. Further, it is connected to Lepeltierer's 6AT (B type) according to JP-A-4-219553 and Mercedes-Benz's 5AT (C type) according to US Pat. No. 5,435,791. In other words, three planetary gear trains and five fastening elements are required in order to obtain 5, 6 AT that can obtain an appropriate gear ratio.

  In the forward deceleration stages of any of the A, B, and C types, the input shaft is connected to the constituent element on the main driving side (driving side) of the four constituent elements consisting of the two planetary gear trains of the main transmission mechanism (MAIN GEAR). Rotation, or the same method of selectively inputting the decelerated rotation of the input shaft obtained from one planetary gear train and one engagement element of the front transmission mechanism (FRONT GEAR). There is a difference in the speed change method, which causes a difference in characteristics. In the forward deceleration stage, the rotation input to the main component of the main transmission mechanism is the rotation of the input shaft for type A, the deceleration rotation of the input shaft for type B, the rotation of the input shaft for type C, and The input shaft is decelerated rotation. The A and B types have the same system as the conventional 3 and 4AT in which the driving side is fixed and the passive side (braking side) is changed to change gears, while the C type is This is a system in which the passive side is fixed and the main drive side is changed to change gears.

  In recent years, 7 and 8 ATs have been put into practical use in order to further improve fuel efficiency and drivability. Daimler's C type 7AT described in JP 2000-266138 based on C type 5AT, and Toyota B type 8AT described in JP 2001-182785 (Aisin Seiki) based on B type 6AT, It is D type 8AT of ZF described in Japanese Patent Publication No. 2008-527267. The D type is a system in which two or more constituent elements of three or more planetary gear trains are connected to each other and a clutch is arranged in the connecting portion to change the connection and change the speed. C type 7AT is a C type 5AT with one planetary gear train and one fastening element (brake) added, four planetary gear trains and six fastening elements (3 clutches, 3 brakes), and B type 8AT is a B type 6AT with one additional fastening element (clutch) and consists of three planetary gear trains and six fastening elements (four clutches and two brakes). D type 8AT has four planetary gears. It consists of rows and 5 fastening elements (3 clutches, 2 brakes). Regarding the most important gear ratio, the B and D type 8AT have a poor step value of the gear ratio, and the Gear Range is also poor, and the B type 8AT is 6.7 and the D type 8AT is 7.3. , Which is considerably smaller than 9 which is suitable for 8 AT. As expected, the C type 7AT with four planetary gear trains and six fastening elements using one more planetary gear train or fastening elements than the B and D type 8ATs has an ideal gear ratio step value, but Gear Range is as small as about 6 and applicable vehicles are limited. The B and D type 8AT, which has a clutch between the components of the planetary gear train, has a more complicated structure than the A and C types, which has a clutch between the input shaft and the components of the planetary gear train. It should be noted that the supply passages are also complicated, and simply reducing the number of planetary gear trains and fastening elements does not lead to simpleness and compactness. By the way, in the skeleton diagram (schematic diagram), the clutch looks smaller than the planetary gear train, but in the actual structure it requires a space as large as or more than the planetary gear train, and the layout of this clutch influences the simplicity and compactness. . In addition to these B and D type 8ATs, a large number of patents have been proposed for 8AT consisting of three planetary gear trains and six fastening elements, and 8AT consisting of four planetary gear trains and five fastening elements. However, a gear ratio suitable for all 8AT is not obtained, and the arrangement of the clutch often complicates the entire structure. In other words, for a passenger car with a small weight, a Gear Range of about 9 is sufficient, and if the step value of the gear ratio, the weight of the transmission and the transmission efficiency are appropriate, the number of gear stages is 7 or 8 (7,8AT). However, with the current 7 and 8 ATs, it is not possible to satisfy it.

  When the number of gear stages and the number of gears are increased by decreasing the number of planetary gear trains and fastening elements, the clutch must change the connection of the components of the planetary gear train at multiple places, which increases the number of connecting parts and complicates the clutch arrangement structure. In addition, since it is necessary to change the power passages of the same planetary gear train in a complicated manner, the step value is inevitably disturbed, and there is a high possibility that the meshing efficiency will deteriorate and the gear shifting performance will deteriorate. On the contrary, even if the number of planetary gear trains and the number of fastening elements are increased, the cost will not be so high if the clutch arrangement and the planetary gear trains are simply connected, and the running cost will be better due to better running performance and fuel efficiency. .

  Therefore, one planetary gear train or one fastening element is added to 8AT having a poor gear ratio, and a forward gear 8 or higher multistage that can obtain a gear ratio suitable for a multistage transmission using four planetary gear trains and six fastening elements. It is necessary to consider a fast AT. Naturally, in addition to the most important gear ratio, the meshing efficiency of the planetary gears and the accompanying rotation loss of the friction member are also important, and weight reduction and compactness that affect mountability are inevitable. Although a torque converter having a torque amplification function is used as a starting device in the AT, a transmission having a Gear Range of about 9 does not require torque amplification, and other means including a motor generator (MG) can be used. It is also necessary to consider the launch device. Naturally, since it is necessary to improve the efficiency by directly connecting to the prime mover, the meshing efficiency of the planetary gears becomes even more important, and a multi-stage transmission including these must be considered. It should be noted that a commercial vehicle whose prime mover has a smaller power than GVW (gross vehicle weight) requires more gears than a passenger car having a large power, and this commercial vehicle must be taken into consideration.

  By the way, it seems that the meshing efficiency of planetary gears is not correctly understood, and in many patent cases, the rotation speed and torque of meshing gears, which are indispensable for efficiency, are not taken into consideration, and the ring gear (internal teeth) and pinion The meshing loss of the gear (external teeth) is less than half the meshing loss of the pinion gear (external teeth) and sun gear (external teeth), and it is ignored, and the meshing efficiency of the planetary gears is insufficiently evaluated. Not only that, but it is wrong. Incidentally, as described in Japanese Patent Laid-Open No. 2013-145016 (the applicant of the present application), the practically used Toyota 8AT has the worst planetary gear meshing efficiency in this class, and many patents are currently applied to the later-mentioned Toyota. The meshing efficiency of the planetary gears of the "5-TYPE" 10AT according to the series is also poor. In addition, Toyota's 2005 HEV, which is a hybrid vehicle that requires the highest priority on efficiency, has a planetary speed reduction structure of the motor generator (MG) on the output side, which is coaxial with the engine, and the meshing efficiency of the planetary gears. A reverse speed reduction mechanism is used that deteriorates. Even Toyota, who claims to "emphasis on fuel efficiency", is in this state. The applicant of the present application obtains a transmission torque and a relative rotation speed between gears that transmit power and meshes with each other, calculates and sums the meshing losses, and calculates and evaluates the meshing efficiency of each gear. In meshing the planetary gears, the power is dispersed and the meshing loss between the ring gear (internal teeth) and pinion gear (external teeth) is half the meshing loss between the pinion gear (external teeth) and sun gear (external teeth). There is a gear stage in which the meshing efficiency exceeds 99% because it does not satisfy, and there are many gear stages in which the gear meshing efficiency is better than MT.

  In recent years, two types of 9AT have been put into practical use as a multi-stage transmission using four planetary gear trains and six fastening elements. All are C type which inputs the speed change rotation to the constituent elements on the main driving side (driving side) of the four constituent elements consisting of the two planetary gear trains of the main transmission mechanism. It consists of a combination of a front speed change mechanism consisting of engagement elements, two planetary gear trains and a main speed change mechanism consisting of two engagement elements. In the C1 type 9AT that was put into practical use in advance, the planetary gear train of the main transmission mechanism is the main transmission mechanism of the Daimler C type 7AT described in Japanese Patent Laid-Open No. 2000-266138. Five kinds of rotations of input shaft deceleration rotation 2 kinds, 0 rotation, and input shaft reverse rotation are selectively inputted, and ZF is 9 AT for FF based on the special table 2011-513662 (ZF). It has been put to practical use. The next commercialized C3 type 9AT is two planetary gear trains (Simpson planetary gear trains) arranged in the order of numbers 1 to 4 on the common speed diagram of the main transmission mechanism of the Daimler C type 7AT. ) Using the planetary gear train of the above-mentioned GM ALLISON 6AT main transmission mechanism in which the input / output components of (1) are reversed, the input shaft rotation, the input shaft deceleration rotation, the input shaft rotation Five types of rotations, namely, speed-up rotation, zero rotation, and reverse rotation of the input shaft, are selectively input. This 9AT was put into practical use by Daimler based on Patent Document 1. That is, the main transmission mechanism can obtain one kind of decelerated rotation and one kind of accelerated rotation by decelerated rotation of one kind of input shaft from the front speed change mechanism, and can be obtained by one kind of accelerated rotation of the input shaft. Since two kinds of decelerated rotations can be obtained, the C1 type 9AT inputs two kinds of decelerated rotations from the front transmission mechanism to the main transmission mechanism to obtain two kinds of decelerated rotations and two kinds of accelerated rotations. On the other hand, the C3 type 9AT is one in which one type of decelerated rotation and one type of increased rotation are input to the main transmission mechanism from the front transmission mechanism to obtain three types of reduced rotation and one type of increased rotation. is there.

  These two types of C1 and C3 type 9AT are a multi-stage transmission that can have a wide Gear Range of about 9 to 11, have an appropriate gear ratio and planetary gear meshing efficiency, and can be completed. Comparing the C1 type 9AT and the C3 type 9AT, the C1 type 9AT has four input shaft rotation speed increasing stages and four deceleration stages, while the C3 type 9AT has three input shaft rotating speed increasing stages. There are 5 speed reduction stages, and the C3 type 9AT is easier to use. However, the accompanying rotation loss caused by the friction members of the clutch and the brake is larger in the C3 type 9AT having a larger reduction ratio. On the other hand, the C1 type 9AT front transmission mechanism can be made simpler than the C3 type 9AT front transmission mechanism because the conventional 3AT can be used. In addition, the C1 type 9AT can employ various planetary gear trains for the front transmission mechanism and the main transmission mechanism depending on the application. 9 AT is proposed. Although the planetary gear train of the main transmission mechanism in the C3 type 9AT is limited, the input / output connection structure is simpler than that of the C1 type 9AT and the meshing efficiency of the planetary gears is improved. There is proposed a multi-stage transmission in which the front transmission mechanism is variously changed by using a main transmission mechanism including two C3 type 9AT planetary gear trains and two fastening elements according to Patent Document 1.

  Many proposals are for skeleton diagrams only, but there are also proposals for some structures. In the C3 type 9AT by Daimler, which is a proposal related to this structure and has been put into practical use, at least one clutch of the front speed change mechanism is arranged around the input shaft (upper part), and hydraulic oil is supplied from the dualized input shaft. It has a complicated structure. The C3 type 9AT is a multi-stage transmission that uses four planetary gear trains and six fastening elements (three clutches and three brakes), and if simply arranged, a planetary gear that requires a space in the axial direction. Four rows and three clutches are required for a total of seven, which lengthens the axial direction. Furthermore, in this speed change mode, since the input shaft is connected to both the front speed change mechanism and the main speed change mechanism, it is necessary to dispose the input shaft at the center of rotation over the entire width of the speed change device. It is unavoidable that the distance between the shaft supports becomes long, and it is necessary to take measures such as dualizing the input shaft. In addition, the main transmission mechanism commonly used in these proposals has a simple connection structure, but the suspended ring gear of the planetary gear train on which one of the brakes is applied requires thrust bearings, and thrust bearings are required. If it is taken between the other planetary gear trains, the axial direction becomes longer, so it is better to take it on the opposite side. This application is for FR for solving various problems of a multi-stage transmission having a shift stage of more than 9AT in which the front shift mechanism is variously changed by using the main shift mechanism of C3 type 9AT and C3 type 9AT. And the structure of a multi-stage transmission for FF. Many of the multi-stage transmissions to which the present application is applied have already applied for patents and can be classified into the following six types (1 to 6-TYPE). It should be noted that the structure of patents that have been put into practical use and structures are clear, but the structure of patents with only a skeleton diagram is unclear, so it is presumed by the arrangement of planetary gear trains and clutches and brakes. The structure will be described.

<1-TYPE>"C3-1,9-15AT"
In the patent proposed by the applicant of the present application in Patent Document 2, the front speed change mechanism includes a main front speed change mechanism including at least one planetary gear train, and one speed change rotational speed of an input shaft of the main front speed change mechanism. It is composed of a sub-positional front speed change mechanism including one planetary gear train capable of selectively inputting only one and a plurality of engagement elements including at least first and second clutches (C1, C2). In Patent Document 2, two types, 15AT using 9AT and A type 6AT for the front transmission mechanism, are listed as examples for FR, but C type 5AT and B type equipped with the auxiliary front transmission mechanism of the same form. Various types of multi-stage transmissions such as 13AT and 14AT using the type 6AT as the front transmission mechanism, and 10AT and 11AT are possible. However, although Patent Document 2 describes an embodiment for FR, which partially overlaps the claimed content of the present application, it is not proposed for FF. It should be noted that the A type 6AT, which is the front transmission mechanism used for the 15AT of Patent Document 2, has a more compact axial direction than the currently commercialized 4AT, and has a power train and structure that can be commercialized as 6AT. With this type of 9AT, the gear ratio and the meshing efficiency of the planetary gears are also appropriate, but if the number of gears is increased beyond that range, the gear ratios do not improve very well.

<2-TYPE> “C3-2, 9AT, 11AT”
Starting with Patent Document 1, DE102010063634 (Japanese Patent Laid-Open No. 2014-500463) (ZF), JP2013-199957 (JATCO, Nissan), JP2013-199958 (JATCO, Nissan), JP2013-199959 (JATCO, Nissan). ), The front transmission is composed of two planetary gear trains and four engagement elements including at least first and second clutches (C1, C2). One planetary gear is changed by changing the connection state of two planetary gear trains with the first and second clutches (C1, C2) and using two fastening elements other than the first and second clutches (C1, C2). One of the four components is connected to the other planetary gear train having output components by transmitting three types of rotations and outputting four types of rotations, and connecting two components each. It has become possible with the gear-shift mode to output. Among these proposals, the proposal of Patent Document 1 is appropriate in terms of gear ratio, planetary gear meshing efficiency, etc., and Daimler put C9 type original 9AT for FR into practical use. The first and second clutches (C1, C2) of the practically used front speed change mechanism are arranged at different axial positions, and the second clutch (C2) has an oil hole provided inside the input shaft. Since the hydraulic oil is supplied from the structure, the structure is complicated and the front transmission mechanism is elongated in the axial direction. It is not proposed for FF. Further, although not described in these patents, if one of the four connected components can be braked by an additional fourth brake, another kind of decelerated rotation of the input shaft can be obtained. This is 11AT (C3-2-2) proposed by the present application. However, 11AT is not established unless the capacity of the second clutch (C2) is further increased, and the step value of the gear ratio is slightly worse than 9AT. The present application is a structural proposal regarding the 9AT proposed by Daimler and the 11AT proposed by the present application.

<3-TYPE> “C3-3, C3-4, 11AT”
The front speed change mechanism has two planetary gear trains consisting of four constituent elements and four clutch elements including the first and second clutches (C1, C2), and further includes one clutch or brake, and C3. It is possible to output by adding one decelerated rotation of the input shaft to the type 9AT. As a patent, Fig. 1 of JP-A-2015-161311 (Honda) with a structure with one additional brake (C3-3-2) and JP-A-2014-2014 with structure with one additional clutch (C3-4-1). 77535 (HUNDAI) and FIG. 1 of JP-A-2015-78763 (HUNDAI) are the subject of the present application. The speed change mode of the main speed change mechanism is the same as 11 AT of (C3-2-2), and one type of deceleration rotation output of the input shaft is added to the front speed change mechanism. However, this front shifting mechanism is the same as the conventional 4AT, and there are power trains other than these patent examples, but JP2015-161311 (Honda) conflicts with JP2012-247057 (HUNDAI). Probability is high. FIG. 1 of Japanese Unexamined Patent Application Publication No. 2015-161311 (Honda) is for FF, and since the output gear is arranged between the front speed change mechanism and the main speed change mechanism, the clutch is arranged between the ring gear R3 of the main speed change mechanism and the planet carrier C4. Of the ring gear R4 becomes complicated and the holding mechanism that must be held by the thrust needle bearing in the small diameter portion of the ring gear R4 becomes complicated, and the axial direction is lengthened. Fig. 1 of 78763 (HUNDAI) is for FR, and the two clutches of the front speed change mechanism are arranged on the side opposite to the main speed change mechanism side of the front speed change mechanism, so that the input shaft has four speed change gears including the front speed change mechanism and the main speed change mechanism. Since the planetary gear train and other parts are axially supported, the diameter of the input shaft must be increased to prevent vibration, resulting in an increase in weight and cost. In these patents, the clutch of the main transmission mechanism is provided between the input shaft and the planet carrier, but the clutch arranged between the ring gear R3 used in the C3 type 9AT according to Patent Document 1 and the planet carrier C4. Has a smaller capacity, and this structure should be established. The sequence of step values of the gear ratio is the same as that of the above-mentioned “C3-2-2, 11AT”, which is worse than 9AT, but the Gear Range is properly taken, and the meshing efficiency of the planetary gears is not bad.

<4-TYPE> “C3-5, 14AT”
The front speed change mechanism allows a main front speed change mechanism consisting of one planetary gear train, and two brakes to the main front speed change mechanism to selectively input two types of decelerated rotation speeds of the input shaft. It is composed of an auxiliary front speed change mechanism consisting of a planetary gear train, five first and second clutches (C1, C2) for controlling the main front speed change mechanism, and five fastening elements including one brake. . The form of this front shift mechanism is the same as "C3-1, 9AT", and one main brake and planetary gear train are provided for the same main front shift mechanism and sub front shift mechanism as in "C3-1, 9AT". By adding two types of deceleration rotation speeds to the main front transmission mechanism in addition to the auxiliary front transmission mechanism, speed increase rotation, deceleration rotation, reverse rotation and rotation of the input shaft of each two types of input shafts can be achieved. Eight kinds of rotations of 0 rotation are input to the main transmission mechanism to be 14AT. The meshing efficiency of the "C3-1, 9AT" planetary gear is the highest level, and 14AT is also the highest level of the meshing efficiency of the planetary gear. However, the capacity of the clutch must be increased. This method has not been published yet, and is filed as a patent application. The front transmission mechanism used for this 14AT is a 5AT that is more compact in the axial direction than the currently commercialized 4AT, which is also a form that can be commercialized as 5AT.

<5-TYPE> “C4-1, 10AT”
Although the same speed change mode as the C3 type is used in the deceleration stage of the transmission, the "C4 type" is used because the speed change type is different in the speed increasing stage. As patents, starting from US20090054196 (GM), JP-A-2015-52377 (AW), JP-A-2015-64042 (Aisin Seiki), JP-A-2015-64099 (AW, Toyota), JP-A-2015-72061 (AW), Many proposals have been made regarding JP-A-2015-83853 (Aisin Seiki), JP-A-2015-105721 to 105726 (AW), JP-A-2015-183855 (GM), power train, and structure. The front transmission mechanism includes two planetary gear trains composed of four constituent elements, first and second clutches (C1 and C2) that can be connected to the first constituent element of the main transmission mechanism, and the first constituent element of the main transmission mechanism. A fourth clutch (C4) that can be connected to the four constituent elements, and four fastening elements that are brakes that enable braking of a part of the planetary gear train, and the same four planetary gear trains as C3 type 9AT. With 6 fastening elements, a gear position of 10 AT is added. However, the Gear Range is as small as about 8 and the step value of the gear ratio becomes a geometric series between the third forward speed and the seventh forward speed, and the fourth component of the main transmission mechanism and the front transmission mechanism can be connected. The connection of the fourth clutch (C4) that makes the structure complicated. Further, in the power train of Japanese Patent Laid-Open No. 2005-105721 to 105726 (AW) for which many structural patents have been applied, the meshing efficiency of the planetary gears in the third forward speed and the fifth forward speed becomes extremely poor. Further, in all of the many proposed structural drawings, the hydraulic oil to the first and second clutches (C1, C2) passes through the input shaft from the transmission case and is supplied to the clutches, so that the conduit resistance increases. In addition to that, the structure of the clutch is complicated. Naturally, the input shaft additionally requires oil holes for the oil supplied to the torque converter and the lubricating oil, and the input shaft must be made into a two-piece complex in order to enable hole processing. In addition, in this structure where the shaft support of the input shaft becomes long, the diameter of the input shaft must be increased to prevent vibration between the shaft supports, and the speed-increasing rotation part that is arranged around the input shaft more than in a normal transmission The diameter must be increased, which is a disadvantage. There are skeleton diagrams of FIGS. 6, 7, and 10 of JP-A-2015-64099 (AW, Toyota) for FF, and an output gear is arranged on the outer peripheral portion of the main transmission mechanism. The ring gear and the sun gear of this planetary gear train are provided. The ratio of the number of teeth is 3.6, which is larger than usual, and the output gear is larger than that, which is not realistic. Since the front transmission mechanism and the main transmission mechanism are not arranged in order in the axial direction, they are not the subject of the present application. However, in JP-A-2005-183793 (AW, Toyota), an output gear is provided between two planetary gear trains of the main transmission mechanism. Although the arrangement for the FF in which the four gears are arranged is described, a power train in which as many as four clutches must be arranged and an output gear is pivotally supported between two planetary gear trains of the main transmission mechanism, for example, a transmission. If a partition that is integrated with is provided, the axial direction becomes long and it is difficult to establish.

<6-TYPE> “C5-1, 10AT”
Since only the reverse speed stage of the transmission uses a shift mode different from the C3 type, the shift type is set to "C5 type". The reverse gear is the same as the C type 5AT according to Daimler US Pat. No. 5,435,791 and the C type 7AT according to Japanese Patent Laid-Open No. 2000-266138, and the second component of the main transmission mechanism can be braked by the fourth brake (B4). It is provided. The front transmission mechanism is composed of two planetary gear trains composed of four constituent elements and four engagement elements including first and second clutches (C1, C2). As patents, FIG. 1 of JP-A-2015-161313 (Honda) and FIG. 1 of JP-A-2015-161313 (Honda) are the subject of the present application. Although there are power trains other than these patent examples, JP-A-2012-225506 (HUNDAI) has the same speed change mode and some of them have the same planetary gear train as JP-A-2015-161313 (Honda). The main transmission mechanism to which the present invention is applied has the fourth brake (B4) arranged in the second component, and therefore can be applied only to the FF. 1 of Japanese Unexamined Patent Publication No. 2015-161313 (Honda) and FIG. 1 of Japanese Unexamined Patent Publication No. 2015-161313 (Honda), the output gear is arranged between the front transmission mechanism and the main transmission mechanism as in “C3-3, 11AT”. Therefore, the holding mechanism of the ring gear R4, which must be held by the thrust needle bearing, becomes complicated, and the axial direction is lengthened. In addition, in this skeleton diagram in which four planetary gear trains having no compact element are arranged in the axial direction, the axial direction becomes too long and it becomes difficult to use for FF, and therefore, in FIG. 4 of JP-A-2015-161313 (Honda). As described above, the output gear must be arranged on the outer peripheral part of the planetary gear train of the main transmission mechanism, but the diameter of the output gear and the meshing counter gear must be increased, resulting in an increase in weight and volume, and a narrow engine. It adversely affects the placement in the room. Only, the two planetary gear trains of the front transmission mechanism of JP-A-2015-161313 (Honda) can be constructed in two stories and can be used for FF, but with this power train, four forward speeds are possible. The meshing efficiency of the planetary gears in the gear becomes extremely poor. Incidentally, the Gear Range can be properly taken, and the series of step values of the gear ratio is a better one within 10 AT.

From the above, <1-TYPE> “C3-1, 9AT” and <2-TYPE> “C3-2 using four planetary gear trains and six fastening elements (three clutches, three brakes) are used. , 9AT ”is the most important step sequence of gear ratios, Gear Range, and the meshing efficiency of planetary gears are good, and the same number of four planetary gear trains and six (4 clutches, 2 brakes) It can be seen that <5-TYPE> “C4-1, 10AT” using the individual fastening elements deteriorates the chain of step values of the gear ratio, the Gear Range, and the meshing efficiency of the planetary gears. As described in the paragraph “0006”, “For a passenger car, it is enough that the Gear Range is about 9, and if the step value of the gear ratio, the weight of the transmission and the transmission efficiency are appropriate, the number of gears is 7th or 8th speed (7,8AT). ) Is sufficient. ”At this stage, it is clear that“ C3-1,9AT ”and“ C3-2,9AT ”are superior to“ C4-1,10AT ”in driving performance and fuel efficiency. . The "11AT" with one additional fastening element is doubtful as to the need for a passenger vehicle, but if it can be compactly arranged, it can be an appropriate transmission for a commercial vehicle. "11AT" is a speed-increasing stage of <2-TYPE>"C3-2,11AT" or <3-TYPE>"C3-3, C3-4, 11AT", which is one more deceleration rotation of the front transmission mechanism. "11AT", which has 3 speed stages, is the "1-TYPE""C3-1,11AT", which has one speed increase rotation of the front speed change mechanism, compared to "11AT" which has 4 speed stages. Is good. Regarding "10AT" of <6-TYPE>"C5-1", which is an exclusive FF with one additional fastening element, the meshing efficiency of the planetary gears in the fourth forward speed is extremely poor, but has a simple structure. Value comes out if you can. One of the purposes of the present application is to identify the limit of multi-stage of efficient and compact AT, and further multi-stage <4-TYPE> “C3-5, 14AT” and <1-TYPE> “C3- 1, 15 AT ”will also be examined. He said, "For a passenger car, the number of gears is sufficient for 7th and 8th speed (7,8AT)." However, if the Gear Range is 3 times wider than 6AT, it is recommended to use 14AT or 15AT. If possible, an engine speed of 1000 to 2000 RPM will suffice, and if an engine with good fuel consumption can be developed with this specification, it will bring a great change. Naturally, commercial vehicles such as trucks and buses will bring even more changes.
Although not the main subject of the present application, regarding the above six types of multi-stage transmissions, in addition to the already known succession of step values of gear ratios and Gear Range, in the present situation where the meshing efficiency of the planetary gears cannot be calculated, Calculates the meshing efficiency of the planetary gears, examines the structure, and makes a comprehensive evaluation.

  The six types of power trains of the present application using the same main transmission mechanism as the C3 type 9AT that Daimler put into practical use for FR based on the patent document 1 are the patent documents 2 proposed by the present applicant as shown in paragraph “0013”. All of the patent applications have been filed for passenger cars, except for "C3-1, 9-15AT" according to US Pat. A vehicle transmission using a torque converter (Wandler) was originally put to practical use for a shared bus (city bus) in Germany, which developed Wandler, and then used by GM in the United States for passenger cars. Therefore, 100% of city buses are converted to ATs in Europe and America, and ATs with retarders are also popular. Drivers of trucks, buses, etc. are so-called professionals and require a certain level of skill, but there are many serious accidents due to driving errors, including lack of manpower and lack of skill. In recent years, the self-driving technology that has been researched has become necessary for the safety of trucks and buses with larger weight, and it is expected that the needs for automatic transmission will spread to trucks and buses in the future. As an AT for commercial vehicles such as trucks and buses, even if the number of planetary gear trains and the number of fastening elements increase, the gear ratio range (Gear Range) is wide enough to match the number of gear stages, and it is required that the axial direction be compact. A multi-stage transmission that requires 10 times the load frequency and 10 times the durability life of a passenger car is required. Due to its rigor, there are only a few manufacturers in the world, and the supply models are inadequate. The applicant of the present application has proposed C1 and C2 types 9 to 12 AT in Japanese Patent Laid-Open No. 2014-224547 and C3 types 9 to 15 AT in Patent Document 2 as ATs for commercial vehicles. In the present application, everything is considered for commercial vehicles, but as a concrete structural diagram, “C3-5, 14AT” shown in the paragraph “0016” in which the meshing efficiency of the planetary gear is particularly improved is targeted.

DE102008055526 (Daimler) Japanese Patent Application No. 2014-195316 (Applicant of the present application)

  The first problem of the present invention is that the Daimler described in paragraphs “0010” to “0012” has been put into practical use for FR based on Patent Document 1 and has a wide Gear Range of about 9 to 11, a gear ratio and a planet. <2-TYPE according to Patent Document 2 filed by the applicant of the present application, which uses a C2-type 9AT of <2-TYPE> having an appropriate gear meshing efficiency and the same main transmission mechanism as that of the C3 type 9AT, which is simpler and has the same or better performance. Regarding the C3 type 9AT of <>, although it is partially shown in Patent Document 2 for FR, it proposes a simpler and more compact structure as FR and FF specifications.

  A second object of the present invention is to improve the meshing efficiency of a planetary gear further multistaged based on C3 type 9AT of <1,2-TYPE> according to Daimler and Patent Documents 1 and 2 proposed by the present applicant. "11AT, 15AT" of <1-TYPE> shown in 0013 "," 11AT "of <2-TYPE> shown in paragraph" 0014 ", and" 14AT of <4-TYPE> shown in paragraph "0016". Is to be realized compactly as not only FR but also FF specifications.

  A third object of the present invention is to use the same main transmission mechanism as the C2-type 9AT of <2-TYPE> by Daimler in Japanese Patent Laid-Open Publication No. 2001-331, which is referred to in paragraph <0015>, which is already filed. “C3-3, C3-4, 11AT” of TYPE>, “C4-1, 10AT” of <5-TYPE> shown in paragraph “0017”, and of <6-TYPE> shown in paragraph “0018”. We propose a simple and compact structure for the three types of "C5-1, 10AT" that have some drawbacks.

  A fourth object of the present invention is to describe "C3-5, 14AT" of <4-TYPE> shown in paragraph "0016" similar to Patent Document 2 proposed by the applicant of the present application, in which a patent application for power train has not yet been made. Is applied as a structural patent.

  The present invention according to claim 1 relates to a common structure of a multi-stage transmission that uses a main transmission mechanism used in a C3 type 9AT and is combined with various front transmission mechanisms. A first planetary gear train (10), which is a means for solving the problems, and which has components of a first sun gear (S1), a first planetary carrier (P1), and a first ring gear (R1), which are simple planetary gears, , The second sun gear (S2) consisting of a simple planetary gear, the second planetary carrier (P2), the second planetary gear train (20) having the components of the second ring gear (R2), the connected first sun gear ( S1) and the second sun gear (S2) as the first constituent element, and the first planetary carrier (P1) whose input shaft is connectable with the third clutch (C3) as the second constituent element and also the first ring Gear (R1) and No. The planetary carrier (P2) is the third component, or the first planetary carrier (P1) connected to the input shaft is the second component and the first ring gear (3) is connectable with the third clutch (C3). R1) and the second planet carrier (P2) are the third constituent elements, the second ring gear (R2) that can be braked by the third brake (B3) is the fourth constituent element, and the second planet carrier (P2) is Rotation of the input shaft and at least input to the first component of the main transmission mechanism, which is connected to the output shaft and has a common velocity diagram in which the first, second, third, and fourth components are arranged side by side in the axial direction It has at least two planetary gear trains and at least four engagement elements including first and second clutches (C1 and C2), which can selectively input a plurality of variable speed rotations including an accelerated rotation of the shaft. It is equipped with a front speed change mechanism and is the first of the main speed change mechanisms. A multi-stage transmission in which at least the ninth forward speed and the first reverse speed are obtained by the third constituent element by selectively restricting the rotation of any two constituent elements of the second and fourth constituent elements. In addition, the front transmission mechanism and the main transmission mechanism are arranged in the axial direction in the transmission case of the multi-stage transmission, and the input shaft is integrally arranged at the center of rotation of the front transmission mechanism and the main transmission mechanism. The input shaft is pivotally supported by the inner peripheral end of the cylindrical member provided on one end of the transmission case on the front transmission mechanism side on the main transmission mechanism side, and the input shaft is pivotally supported on the other end of the transmission case. Alternatively, the input shaft is rotatably supported by the output shaft rotatably supported at the other end of the transmission case, and the friction members are connected to the first and second clutches (C1, C2) of the pre-transmission mechanism. It is rotatably arranged on the outer side in the outer peripheral direction of the cylindrical member so as to overlap in two steps in the direction, The hydraulic oil is supplied from the outer circumference of the cylindrical member to the first and second clutches (C1, C2), and at least 2 of the front speed change mechanism is axially arranged from the inner peripheral end of the cylindrical member toward the main speed change mechanism. Among the planetary gear trains, at most one planetary gear train or a pair of planetary gear trains in which two planetary gear trains are radially stacked in two stories is arranged.

  The present invention according to claim 2 relates to the structure and arrangement of the first and second clutches (C1, C2) commonly used in various front transmissions, and a hydraulic oil supply circuit. 3 is a means for solving the problem of No. 3, in which a cylindrical input coupling member (Y) that is coupled to the input shaft is rotatably arranged on the outer peripheral direction outer side of the cylindrical member, and the first and second front transmission mechanisms are provided. The clutches (C1, C2) are provided with a dual clutch connecting member (X) in which each friction member is coaxially overlapped in two radial steps, and the input connecting member (Y) and the dual clutch connecting member (X) are integrated. The two clutch coupling members (X) are connected to each other or rotatably arranged outside the input coupling member (Y) in the outer circumferential direction. A hydraulic chamber that serves as a servo mechanism for the first and second clutches (C1, C2) is provided Wherein the outer peripheral first member, was no to supply hydraulic fluid through the input connecting member (Y) to the second hydraulic chamber of the clutch (C1, C2).

  The present invention according to claim 3 relates to an arrangement structure of the FF specification of a multi-stage transmission that uses a main transmission mechanism used for a C3 type 9AT and is combined with various front transmission mechanisms. A means for solving the problem is to provide a partition wall that is integral with the transmission case at the central portion of the transmission case in the axial direction, and arrange an output counter gear that is axially supported by the partition wall. The first planetary gear train (10) of the main transmission mechanism is arranged on one side in the direction, and the second planetary gear train (20) of the main transmission mechanism and the front transmission mechanism are arranged on the other side.

According to a fourth aspect of the present invention, a hydraulic oil supply circuit for the third clutch (C3) of the main speed change mechanism according to the third aspect of the present invention, which has an FF specification, and a brake arrangement in a configuration in which the multi-stage transmission has four brakes. It relates to a structure and is a means for solving the first to third problems, in which the first planetary carrier (P1) connected to the input shaft of the main transmission mechanism is used as the second component and the third clutch ( In the configuration in which the first ring gear (R1) and the second planet carrier (P2) that can be connected by C3) are used as the third component, the first ring gear (R1) and the first ring gear (R1) are connected to the partition wall integrated with the transmission case. A passage for supplying hydraulic oil to the third clutch (C3) that enables connection of the two-planetary carrier (P2) is not provided, or in a configuration in which the multi-stage transmission has four brakes, one is provided in the partition wall. Blur None of to provide a hydraulic chamber which is a key servomechanism.

The present invention according to claim 5 provides paragraphs [0013] and [11AT] and paragraphs [0016] in paragraphs [0013] and [0014] with which the meshing efficiency of the C3 type 9AT and the planetary gears having multiple stages based on the C3 type 9AT are improved. The structure of the first and second clutches (C1, C2) forming the front transmission mechanism of "14AT" shown in FIG. 2 is a means for solving the first and second problems, and at least 2 In a pre-transmission mechanism having a plurality of planetary gear trains, a planetary carrier of either one of the planetary gear trains can be fastened to the input shaft and other components by the first and second clutches (C1, C2). , A side member that pivotally supports the planet gears of the planet carrier is connected to a double clutch connecting member (X) of the first and second clutches (C1 and C2), and the side member has first and second clutches (C1 and C1). C ) Of no to provide a hydraulic chamber comprising at least one of the servo mechanism.

  The present invention according to claim 6 relates to the front transmission mechanism of "9AT" according to Patent Document 2 filed by the applicant of the present application, "11AT" shown in paragraph "0013", and "14AT" shown in paragraph "0016". And a means for solving the first, second, and fourth problems, in which the front speed change mechanism includes a main front speed change mechanism and a sub front speed change mechanism. The mechanism makes it possible to connect the input shaft with the first clutch (C1) to the constituent element E, which is a planet carrier of one simple planetary gear in which three constituent elements D, E, and F are arranged side by side in the axial direction. At the same time, the component E and the component D or F can be connected by the second clutch (C2), the component F and the first component of the main speed change mechanism are connected, and the component E is connected by the second brake (B2). It is the first main front shifting mechanism that is capable of braking. In other words, it is a second main front transmission mechanism in which the constituent element D of the first main front transmission mechanism can be braked by the fourth brake (B4), and the sub front transmission mechanism is A, B, or C. A first sub-positional transmission mechanism in which the input shaft is connected to the constituent element C of one planetary gear train in which the constituent elements are arranged in the axial direction and the constituent element A can be braked by the first brake (B1). Alternatively, the input shaft is connected to the component A that is composed of two planetary gear trains in which four components A, B, C, and G are arranged in the axial direction, and the component G is connected to the first brake. (B1) is a second auxiliary front speed change mechanism that can be braked and component C can be braked by a fourth brake (B4). The component B of the speed change mechanism is connected to the first and second clutches (C1, C2), and the first and second brakes (B1, B). ) By engaging any two of them, the component F of the first main front speed change mechanism is configured to rotate the input shaft, reduce the speed of the input shaft, increase the speed of the input shaft, rotate 0, Reverse rotation of the shaft and five types of rotation are selectively obtained, and rotation of any two of the first, second and fourth constituent elements of the main transmission mechanism is selectively restricted. As a result, the third component is configured to obtain a shift position of 9 forward gears and 1 reverse gear, or the component D of the second main front transmission mechanism and the component B of the first sub front transmission mechanism are connected. Of the first and second clutches (C1, C2) and the first, second, and fourth brakes (B1, B2, B4) are engaged to configure the second main front transmission mechanism. Element F has six types: input shaft rotation, input shaft deceleration rotation, input shaft speed-up rotation, two rotations, zero rotation, and input shaft reverse rotation. Of the main transmission mechanism by selectively restricting the rotation of any two of the first, second, and fourth constituent elements of the main transmission mechanism to advance the third constituent element. The 11th speed reverse speed 1st speed is obtained, or the component D of the first main front transmission mechanism and the component B of the second auxiliary front transmission mechanism are connected to each other, and the first and second clutches ( C1, C2) and any one of the first, second, and fourth brakes (B1, B2, B4) are engaged so that the component F of the first main front transmission mechanism can rotate the input shaft. , Two types of decelerating rotation of the input shaft, two types of increasing rotation of the input shaft, zero revolution, and two types of reverse rotation of the input shaft are selectively obtained. By selectively restricting the rotation of any two of the first, second, and fourth components, the third component advances 1. Was none to obtain the gear position of the speed one reverse speed.

  The present invention according to claim 7 relates to the power train of the front transmission mechanism of "9AT" according to Patent Document 1 shown in paragraph "0014" and "11AT" proposed in the present application, and has the first and second problems. This is a means for solving the problem, and the front speed change mechanism has a third planetary gear having a third sun gear (S3), a third planetary carrier (P3), and a third ring gear (R3). The third sun gear (S3) of the fourth planetary gear train (40) including the gear train (30), the fourth sun gear (S4), the fourth planet carrier (P4), and the fourth ring gear (R4). ) Is connected to the input shaft, the third planet carrier (P3) and the fourth ring gear (R4) can be connected by the second clutch (C2), and the third ring gear (R3) and the fourth planet carrier (P4) are connected. And enter as the third planet carrier (P3) The shaft is connectable with the first clutch (C1), the fourth ring gear (R4) is connected with the first component of the main transmission mechanism, and the third planetary gear is connected with the first and second clutches (C1, C2). A transmission mechanism for selectively changing the connection between the train (30) and the fourth planetary gear train (40), wherein the third planet carrier (P3) can be braked by the first brake (B1), and the fourth sun gear (P3) S4) can be braked by the second brake (B2), and by engaging any two of the first and second clutches (C1, C2) and the first and second brakes (B1, B2), the main shift A fourth ring gear (R4) connected to the first component of the mechanism rotates the input shaft, decelerates the input shaft, accelerates the input shaft, zero revolutions, reverse rotation of the input shaft, 5 types of rotation are selectively obtained, and the first, second, and fourth components of the main transmission mechanism are required. By selectively restricting the rotation of any two of the above components, the third component is made to obtain a shift speed of 9 forward gears and 1 reverse gear, or the third planetary carrier (P3) is The first brake (B1) can be braked, the fourth sun gear (S4) can be braked by the second brake (B2), and the connected third ring gear (R3) and fourth planet carrier (P4) can be braked by the fourth brake ( B4) enables braking, and by engaging any one of the first and second clutches (C1, C2) and the first, second, and fourth brakes (B1, B2, B4), the main transmission mechanism The fourth ring gear (R4) connected to the first component includes rotation of the input shaft, two types of decelerated rotation of the input shaft, increased rotation of the input shaft, zero rotation, and reverse rotation of the input shaft. 6 rotations of the main transmission mechanism are selectively obtained. 4 by selectively restricting rotation of any two elements of the component, the third component is no to obtain the gear position of the forward 11 speeds and one reverse speed.

  In the configuration according to claim 1, a first planetary gear train (10) having components of a first sun gear (S1), a first planetary carrier (P1) and a first ring gear (R1), which are simple planetary gears. , A second sun gear (S2) consisting of a simple planetary gear, a second planetary carrier (P2), a second planetary gear train (20) having the components of a second ring gear (R2), the connected first sun gear ( S1) and the second sun gear (S2) as the first constituent element, and the first planetary carrier (P1) whose input shaft is connectable with the third clutch (C3) as the second constituent element and also the first ring The gear (R1) and the second planet carrier (P2) can be used as the third component, or the first planet carrier (P1) connected to the input shaft can be used as the second component and can be connected by the third clutch (C3). No. 1 ring gear R1) and the second planet carrier (P2) are the third constituent elements, the second ring gear (R2) that can be braked by the third brake (B3) is the fourth constituent element, and the second planet carrier (P2) is Rotation of the input shaft and at least input to the first component of the main transmission mechanism, which is connected to the output shaft and has a common velocity diagram in which the first, second, third, and fourth components are arranged side by side in the axial direction It has at least two planetary gear trains and at least four engagement elements including first and second clutches (C1 and C2), which can selectively input a plurality of variable speed rotations including an accelerated rotation of the shaft. By providing a front transmission mechanism and selectively restricting rotation of any two of the first, second and fourth constituent elements of the main transmission mechanism, the third constituent element can move forward at least in the 9th speed and reverse speed. A multi-stage transmission that is adapted to obtain a first speed shift stage, The front speed change mechanism and the main speed change mechanism are axially arranged inside the transmission case of the stepped speed change device, and the input shaft is integrally arranged at the center of rotation of the front speed change mechanism and the main speed change mechanism. , The input shaft is rotatably supported by the inner peripheral end portion of the main transmission mechanism side of the cylindrical member provided at one end on the front transmission mechanism side, and the input shaft is rotatably supported by the other end portion of the transmission case, or The input shaft is rotatably supported by the output shaft rotatably supported at the other end of the transmission case, and the first and second clutches (C1, C2) of the front speed change mechanism are arranged such that each friction member has two stages in the radial direction. Is arranged rotatably outside in the outer circumferential direction of the cylindrical member so as to supply hydraulic oil from the outer circumference of the cylindrical member to the first and second clutches (C1, C2). In the axial direction from the main transmission mechanism to at least two planetary gear trains of the front transmission mechanism. Of these, at most one planetary gear train or a pair of planetary gear trains in which two planetary gear trains are radially stacked in two stories is arranged, so a simple and compact structure is provided. Becomes In particular, in addition to being able to integrate the input shaft, the distance between the shafts of the input shaft can be shortened, and it is not necessary to make the input shaft larger than the thickness corresponding to the torque capacity, and the diameter of the portion around the input shaft is not increased and the weight can be reduced.

  According to the configuration of claim 2, a cylindrical input connecting member (Y) that is connected to the input shaft is rotatably arranged on the outer peripheral direction outer side of the cylindrical member, and the first and second clutches (C1, C2) is provided with a dual clutch connecting member (X) in which each friction member is coaxially overlapped in two stages in the radial direction, and the input connecting member (Y) and the dual clutch connecting member (X) are integrally connected, or , The two-clutch connecting member (X) is rotatably arranged on the outer peripheral direction outer side of the input connecting member (Y), and the first and second connecting clutch members (Y) or the two-clutch connecting member (X) are provided. A hydraulic chamber serving as a servo mechanism for the clutches (C1, C2) is provided, and hydraulic oil is supplied from the outer circumference of the cylindrical member to the hydraulic chambers of the first and second clutches (C1, C2) through the input connecting member (Y). As a result, the first and second clutches (C1, C2) are axially engaged. Disposed extract, comprising first, second clutch (C1, C2) responsive also shorter pipeline resistance is reduced clutch engaging and disengaging the supply oil passage to the hydraulic chamber well of.

  According to the configuration of claim 3, a partition wall that is integral with the transmission case is provided at the central portion of the transmission case in the axial direction, an output counter gear that is axially supported by the partition wall is arranged, and the output counter gear is sandwiched in the axial direction. The first planetary gear train (10) of the main transmission mechanism is arranged on one side, and the second planetary gear train (20) of the main transmission mechanism and the front transmission mechanism are arranged on the other side. A thrust bearing for holding the axial direction of the second ring gear (R2) floating in the axial direction of the two planetary gear train (20) is compactly arranged between the second planetary gear train (20) and the front transmission mechanism. In addition, the first ring gear (R1), which is an output component of the first planetary gear train (10), has an FF specification arrangement structure in which the output counter gear is arranged in the axial center of the transmission case. And the output configuration of the second planetary gear train (20) Connecting the second planet carrier (P2) and an output counter gear to be also good balance compact.

  According to the configuration of claim 4, the first planetary carrier (P1) connected to the input shaft of the main transmission mechanism is used as the second component, and the first ring gear (R1) is connectable by the third clutch (C3). ) And the second planet carrier (P2) as the third constituent element, a third ring gear (R1) and a second planet carrier (P2) can be connected to a partition integrated with the transmission case. A passage for supplying hydraulic oil to the clutch (C3) is not provided, or in the multi-stage transmission having four brakes, a partition is provided with a hydraulic chamber that serves as a servo mechanism for one brake. Therefore, the third clutch (C3), to which hydraulic oil is supplied from the partition wall, can be compactly arranged between the partition wall and the shaft of the first planetary gear train (10). Further, when the hydraulic chamber of the brake is provided in the partition wall, the axial direction becomes compact, and the effect becomes remarkable especially in the configuration having four brakes.

In the structure according to claim 5, the planetary carrier of either one of the planetary gear trains of the front speed change mechanism having at least two planetary gear trains is the first and second clutches (C1, C2) and serves as the input shaft and the other. In the configuration that can be fastened to the component of (1), the side member that pivotally supports the planetary gear of the planet carrier is connected to the double clutch connecting member (X) of the first and second clutches (C1, C2), and the side member is also connected. Since the hydraulic chamber that serves as the servo mechanism for at least one of the first and second clutches (C1, C2) is provided in the above, the clutch cover member is shared with the side member that axially supports the planet gears of the planet carrier. It can be simple and compact.

  In the structure according to claim 6, the front transmission mechanism comprises a main front transmission mechanism and a sub front transmission mechanism, and the main front transmission mechanism has three constituent elements D, E, and F arranged in axial order. The input shaft can be connected with the first clutch (C1) to the component E, which is the planet carrier of the arranged simple planetary gear, and the component E and the component D or F can be connected with the second clutch (C2). A first main front speed change mechanism that is connectable and that connects the constituent element F and the first constituent element of the main transmission mechanism so that the constituent element E can be braked by the second brake (B2). It is a second main front transmission mechanism in which a component D of the front transmission mechanism can be braked by a fourth brake (B4), and the sub front transmission mechanism has three constituent elements A, B, and C as axes. The input shaft is connected to the constituent element C of one planetary gear train arranged side by side in the direction of the constituent element A. It is a first auxiliary front speed change mechanism that can be braked by the first brake (B1), or from two planetary gear trains in which four constituent elements A, B, C, and G are arranged side by side in the axial direction. A second sub-positional transmission mechanism in which an input shaft is connected to the component A, the component G can be braked by the first brake (B1), and the component C can be braked by the fourth brake (B4). , The component D of the first main front transmission mechanism and the component B of the first auxiliary front transmission mechanism are connected to each other, and the first and second clutches (C1, C2) and the first and second brakes (B1, By connecting any two of B2), the component F of the first main front transmission mechanism causes the rotation of the input shaft, the decelerated rotation of the input shaft, the increased rotation of the input shaft, and the zero rotation. Reverse rotation of the input shaft and five types of rotation, which are reverse rotation, are selectively obtained, and the first, second and fourth configurations of the main transmission mechanism are provided. By selectively restricting the rotation of any two of the constituent elements, the third constituent element is made to obtain a shift speed of 9 forward speeds and 1 reverse speed, or the second main front speed change mechanism is provided. Any of the first and second clutches (C1, C2) and the first, second, and fourth brakes (B1, B2, B4) that connects the component D and the component B of the first auxiliary front transmission mechanism. By connecting two or more of them, the constituent element F of the second main front speed change mechanism is configured such that the input shaft rotates, the input shaft decelerates and rotates, the input shaft accelerates two types, and the input shaft rotates zero. Reverse rotation and six kinds of rotations are selectively obtained, and rotations of any two of the first, second and fourth constituent elements of the main transmission mechanism are selectively restricted. As a result, the third component is configured to obtain the 11th forward gear and the 1st reverse gear, or the component D of the first main front transmission mechanism. And the component B of the second sub pre-position transmission mechanism are connected to each other, and any one of the first and second clutches (C1 and C2) and the first, second and fourth brakes (B1, B2 and B4) By engaging the first main front speed change mechanism, the constituent element F causes the rotation of the input shaft, two types of decelerated rotation of the input shaft, two types of increased rotation of the input shaft, zero revolution, and Eight kinds of rotations, that is, two kinds of reverse rotations are selectively obtained, and rotations of any two of the first, second and fourth constituent elements of the main transmission mechanism are selectively restricted. As a result, the third component obtains the forward 14th speed and reverse 1st speed, so that the “9AT” described in Patent Document 2 is simplified, and the fourth brake (B4) is used as the main front speed change mechanism. By adding to the 11th forward speed and 1st reverse speed of the same speed change mode, one planetary gear train and the fourth brake ( 4) can forward 14 speeds and one reverse speed multistage of the same gear-shift mode by adding the sub before 置変 speed mechanism, both a simple structure.

  In the configuration according to claim 7, the front transmission mechanism has a third planetary gear (S3), a third planetary carrier (P3), and a third ring gear (R3), which are simple planetary gears. The third sun gear (S3) of the fourth planetary gear train (40) including the gear train (30), the fourth sun gear (S4), the fourth planet carrier (P4), and the fourth ring gear (R4). ) Is connected to the input shaft, the third planet carrier (P3) and the fourth ring gear (R4) can be connected by the second clutch (C2), and the third ring gear (R3) and the fourth planet carrier (P4) are connected. To connect the third planet carrier (P3) and the input shaft with the first clutch (C1), and also connect the fourth ring gear (R4) to the first component of the main transmission mechanism. With the second clutch (C1, C2) and the third planetary gear train (30) A transmission mechanism for selectively changing the connection of the fourth planetary gear train (40), wherein the third planet carrier (P3) can be braked by the first brake (B1), and the fourth sun gear (S4) is driven by the second brake. (B2) enables braking, and by engaging any two of the first and second clutches (C1, C2) and the first and second brakes (B1, B2), the first constituent element of the main transmission mechanism The fourth ring gear (R4) connected to the input shaft rotates five kinds of rotations of the input shaft, the input shaft's decelerated rotation, the input shaft's accelerated rotation, 0 rotation, and the input shaft's reverse rotation. By selectively restricting the rotation of any two of the first, second, and fourth constituent elements of the main transmission mechanism, the third constituent element can move forward in the ninth forward speed and reverse speed. Not to obtain the 1st speed, or the 3rd planetary carrier (P3) to the 1st break. (B1) can be braked, the fourth sun gear (S4) can be braked by the second brake (B2), and the connected third ring gear (R3) and fourth planet carrier (P4) can be braked by the fourth brake (B4). The first and second clutches (C1, C2), and any one of the first, second, and fourth brakes (B1, B2, B4) are engaged to make the first transmission of the main transmission mechanism. The fourth ring gear (R4) connected to the constituent elements includes the rotation of the input shaft, two types of decelerated rotation of the input shaft, the increased rotation of the input shaft, zero rotation, and the reverse rotation of the input shaft. By selectively restricting the rotation of any two of the first, second and fourth constituent elements of the main transmission mechanism, the third constituent element is configured so that the third constituent element Since the shift speed of the 11th forward speed and the 1st reverse speed is obtained, the "9AT" according to Patent Document 1 is Be pulled further, the fourth brake the (B4) can forward 11 speeds and one reverse speed in the same gear-shift mode by adding before 置変 speed mechanism, the proposed "11AT" is the simple structure.

<1-TYPE> “C3-1-1, 9AT” A schematic diagram and a velocity diagram showing an embodiment of a C3 type 9AT according to Patent Document 2 of the present invention, and a table showing a gear ratio and gear meshing efficiency. <1-TYPE> “C3-1-2, 11AT” A schematic diagram and a velocity diagram showing an embodiment of a C3 type 11AT according to Patent Document 2 of the present invention, and a table showing a gear ratio and gear meshing efficiency. <1-TYPE> “C3-1-3, 15AT” A schematic diagram and a velocity diagram showing an embodiment of a C3 type 15AT according to Patent Document 2 of the present invention, and a table showing a gear ratio and meshing efficiency of gears. <2-TYPE> “C3-2-1, 9AT” A schematic diagram and a velocity diagram showing an embodiment of a C3 type 9AT according to Patent Document 1 of the present invention, and a table showing a gear ratio and meshing efficiency of gears. <2-TYPE> "C3-2-2, 11AT" A schematic diagram and a velocity diagram showing an embodiment of a C3 type 11AT in which a brake (B4) is added to Patent Document 1 of the present invention, and a gear ratio and meshing of gears. Table showing efficiency. <3-TYPE> "C3-3-1, 11AT" A schematic diagram and a velocity diagram showing an embodiment of the C3 type 11AT of the present invention, and a table showing a gear ratio and gear meshing efficiency. <3-TYPE> “C3-3-2, 11AT” A schematic diagram and a velocity diagram showing an example of a C3 type 11AT according to JP-A-2015-161311 (Honda) and JP-A-2012-247057 (Hyundai) of the present invention, And a table showing gear ratio and gear meshing efficiency. <3-TYPE> "C3-4-1, 11AT" A schematic diagram and a speed diagram showing an embodiment of a C3 type 11AT according to JP-A-2014-77535 (Hyundai) of the present invention, and a table showing a gear ratio. <4-TYPE> “C3-5-1, 14AT” A schematic diagram and a velocity diagram showing an embodiment of a C3 type 14AT by a new power train similar to Patent Document 2 of the present invention, and a gear ratio and gear meshing efficiency. Table showing. <4-TYPE> “C3-5-1, 14AT” A combination of four types of planetary gears forming a sub pre-position transmission mechanism including four components A, B, C, and G in FIG. 9 is used. Show. <5-TYPE> “C4-1-1, 10AT” A schematic diagram showing a working example of a C4 type 10AT based on US2009-0054196A1 (GM) of the present invention, a velocity diagram, and a table showing a gear ratio. <5-TYPE> "C4-1-2, 10AT" A schematic diagram and a velocity diagram showing an embodiment of a C4 type 10AT according to JP-A-2015-105722 (AW) of the present invention, and a gear ratio and gear meshing efficiency. Table showing. <6-TYPE> "C5-1-1, 10AT" A schematic diagram and a velocity diagram showing an embodiment of a C5 type 10AT according to JP-A-2015-161313 (Honda) and JP-A-2012-2225506 (Hyundai) of the present invention, And a table showing gear ratio and gear meshing efficiency. <1-TYPE> “C3-1-1, 9AT, C3-1-2, 11AT” Structural diagram of FR specifications for passenger cars of “FIG. 1” and “FIG. 2”. <1-TYPE> “C3-1-3, 15AT” FIG. 3 is a structural diagram of FR specifications for passenger cars. <2-TYPE> “C3-2-1, 9AT” Structural diagram of FR specifications for passenger cars of “FIG. 4”. <3-TYPE> "C3-3-1, 11AT" Structural diagram of FR specifications for passenger cars of "Fig. 6". <4-TYPE> “C3-5-1, 14AT” Structural diagram of FR specifications for commercial vehicles of “FIG. 9”. <5-TYPE> "C4-1-2, 10AT" Fig. 12 is a structural diagram of FR specifications for passenger cars. <1-TYPE> “C3-1-1, 9AT” FIG. 1 is a structural diagram of FF specifications for the passenger car. <1-TYPE> “C3-1-2, 11AT” FIG. 2 is a structural diagram of FF specifications for passenger cars. <2-TYPE> “C3-2-1, 9AT” FIG. 4 is a structural diagram of FF specifications for passenger cars. <4-TYPE> "C3-5-1, 14AT" Fig. 9 is a structural diagram of FF specifications for the passenger car of Fig. 9. <6-TYPE> “C5-1-1, 10AT” FIG. 13 is a structural diagram of FF specifications for passenger cars.

  The present invention relates to the structure of six kinds of multi-stage transmissions of "1-6-TYPE", and Figs. 1 to 13 show typical power trains and performances thereof. All patent applications have been filed except for "2-TYPE" 11AT in Fig. 5 and "4-TYPE" 14AT in Fig. 9 and Fig. 10, and the gear ratios written in the tables of Figs. Although based on this, the applicant of the present application appropriately created a different embodiment or a new embodiment. Note that "GEAR EFF" in the table represents the meshing efficiency of the planetary gears, which has not been calculated so far, and is an important item for evaluating the performance of the multi-stage transmission. The meshing loss of the planetary gears is mainly rolling loss and sliding loss of the tooth surface and is proportional to the passing power, so the relative speed and transmission torque of each meshing point are calculated and the loss is calculated and summed as the efficiency. In addition, the meshing loss between the planetary pinion gear and the ring gear is such that the tooth surface of the involute curve part meshes in the same direction, so the surface pressure is low and slippage is reduced. And about 40% of Sun Gear. Therefore, the GEAR EFFs shown in these figures are calculated by the same method and are values that can be compared accurately. The schematic diagrams of FIGS. 1 to 13 show the features of the structure of the present application and the arrangement of the planetary gears and the fastening elements, and are different from the schematic diagrams described in the patents already filed. "Claims 1 and 3" of the present application are represented by this schematic diagram, but other claims are represented by Fig. 14 to Fig. 23 which is a schematic diagram in addition to the schematic diagrams of Fig. 1 to Fig. 13. It

  Claim 1 of the present invention is shown in schematic diagrams of FIGS. 1 to 13 excluding FIG. 10, a structural diagram of FR specifications of FIGS. 14 to 19 and a structural diagram of FF specifications of FIGS. 20 to 24, Claim 2 is shown in the structure diagram of the FR specifications of FIGS. 14 to 19 and the structure diagram of the FF specifications of FIGS. 20 to 24. Claims 3 and 4 are the FF specifications of FIGS. 20 to 24. 5 is shown in the structure diagram of the FR specifications of FIGS. 14, 16 and 18, and the structure diagram of the FF specifications of FIGS. 20 to 23, and claim 6 is shown in FIG. , FIG. 9, FIG. 9 and FIG. 10 schematic diagrams and velocity diagrams, FIG. 14 and FIG. 18 structural drawings of FR specifications, and FIG. 20, FIG. Item 7 is shown in the schematic diagrams and velocity diagrams of FIGS. 4 and 5, the structural diagram of the FR specification of FIG. 16, and the structural diagram of the FF specification of FIG. Note that the FR structural drawing for commercial vehicles in FIG. 18 is based on the concept that the torque of the prime mover is 1000 Nm, and the structural drawings for FR and FF specifications other than FIG. 18 are for passenger cars with the torque of the prime mover being 300 Nm. It is a concept. Therefore, structural drawings for passenger cars can be compared in size by TYPE. The present application is limited to "1-TYPE", "2-TYPE", and "4-TYPE" that are particularly valuable for practical use, and the power train and structure are defined in "Claims 5, 6, and 7". Is.

<1-TYPE>
In the multi-stage transmission described in paragraph “0013”, the transmission mode of the power train of Patent Document 2 proposed by the applicant of the present application has a structure that is the invention of the present application. As an example of the present application, "C3-1-1, 9AT" FIG. 1, FIG. 14, FIG. 20 and "C3-1-2, 11AT" FIG. 2, (FIG. 14), FIG. 21 and "C3-1-" 3, 15AT ”The three types shown in FIGS. 3 and 15 will be described. In Patent Document 2, the power train of 11AT and FF specifications is not described as an example. Further, although part of the embodiments of Patent Document 2 includes the structure of the claims of the present application, the claims of Patent Document 2 do not include the claims of the present application, and in particular, regarding the FF specifications, what is claimed is Since this has not been done, an example will be added and its details will be described in the present application.

<1-TYPE> “C3-1-1, 9AT” FIGS. 1, 14 and 20
FIG. 1 is a schematic diagram of two types of C3 type 9AT, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift stage, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 1, the velocity diagram is divided into a MAIN GEAR (main transmission mechanism) and a FRONT GEAR (front transmission mechanism), and the FRONT GEAR (front transmission mechanism) is a main front transmission mechanism and a sub front transmission mechanism. It is divided into a stationary transmission mechanism. The speed diagram of the MAIN GEAR (main transmission mechanism) is such that the first, second, third, and fourth constituent elements are arranged in order from the right side of the figure, and the speed diagram of the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). The fifth, sixth, and seventh constituent elements (A, B, C) are arranged in order from the right in the figure, and the speed diagram of the main front transmission mechanism is the eighth, ninth, and tenth constituent elements in order from the right in the figure. (D, E, F) are arranged, the sixth component (B) and the eighth component (D) are connected by the second connecting member (8), and the first component and the tenth component (F). Are connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  In the two types of schematic diagrams in Fig. 1, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. A prime mover is located on the front left side (not shown), and power is input to the input shaft of the transmission via the torque converter. In the left diagram of the FR specification, the third planetary gear train (30), the fourth planetary gear train (40), the second planetary gear train (20), the first planetary gear train (20), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40) In the right figure of the FF specification in which the planetary gear train (10) is arranged, the first planetary gear train (10), the second planetary gear train (20), the second planetary gear train (20), the second planetary gear train (20) A four planetary gear train (40) and a third planetary gear train (30) are arranged, and the first and second planetary gear trains (10, 20) constitute a MAIN GEAR (main transmission mechanism), and third and fourth The planetary gear train (30, 40) constitutes a FRONT GEAR (forward gear change mechanism). Further, the third planetary gear train (30) constitutes a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), and the fourth planetary gear train (40) constitutes a main front transmission mechanism. In the right diagram of the FF specification, an output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20) and is axially supported by a partition wall integrated with the transmission case. The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The input shaft is rotatably supported at the other end of the transmission case, and in the FR specification, the input shaft is rotatably supported at the output shaft rotatably supported at the other end of the transmission case. C1 and C2) are arranged between the third planetary gear train (30) and the fourth planetary gear train (40) on the outer peripheral side of the cylindrical member so that the friction members are superposed in two stages in the radial direction. A fourth planetary gear train (S4, P4, R4) of the front transmission mechanism (FRONT GEAR) is arranged in the axial direction from the inner peripheral end portion toward the main transmission mechanism (MAIN GEAR).

  In the two types of schematic diagrams and velocity diagrams of FIG. 1, the first and second sun gears (S1, S2) of the first and second planetary gear trains (10, 20) forming the MAIN GEAR (main transmission mechanism) are shown. The first component, the first planet carrier (P1) is the second component, the first ring gear (R1) and the second planet carrier (P2) are the third component, and the second ring gear (R2) is As a fourth component, the third sun gear (S3), the third planet carrier (P3), and the third ring of the third planetary gear train (30), which constitutes a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism). The gear (R3) is the fifth, sixth, and seventh constituent elements (A, B, C), and the fourth ring gear (R4) and the fourth ring gear (R4) of the fourth planetary gear train (40) constituting the main front transmission mechanism. 4th planet carrier (P4), 4th sun gear (S4) 8th, 9th and 10th structure Component (D, E, F).

  Here, the first and second sun gears (S1, S2) forming the first constituent element of the MAIN GEAR (main transmission mechanism) are connected and also connected to the first connecting member (7) to form the second constituent element. The first planet carrier (P1) that constitutes the first planetary carrier (P1) is connected to the input shaft, the second planetary carrier (P2) that constitutes the third component is connected to the output shaft, and the third clutch (C3) forms the first ring gear. The second ring gear (R2), which is connectable to (R1) and constitutes the fourth component, can be braked by the third brake (B3). The third sun gear (S3) that constitutes the fifth component (A) that is the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism) can be braked by the first brake (B1), and the sixth component ( The third planet carrier (P3) that constitutes B) is connected to the second connecting member (8), the seventh component (C) is connected to the input shaft, and the eighth component (main front transmission mechanism) is formed. The fourth ring gear (R) forming D) is connected to the second connecting member (8), and the fourth planet carrier (P4) forming the ninth component (E) is input by the first clutch (C1). The fourth sun gear (S4), which is connectable to the shaft and can be braked by the second brake (B2), and which constitutes the tenth component (F), is connected to the first connecting member (7) and the ninth The fourth planet carrier (P4) and the eighth component (D) constituting the component (E) A fourth ring gear which constitutes (R) are connectable by a second clutch (C2). In Patent Document 2, the second clutch (C2) that integrates the fourth planetary gear train (40) is provided with a fourth planet carrier (P4) and a tenth component (F) that form a ninth component (E). ) Is connected to the fourth sun gear (S4), but in order to have a simple structure in the present application, the second connecting member (8) and the fourth planetary carrier (8) connected to the fourth ring gear (R) are used. Used for ligation of P4).

  The table of FIG. 1 shows the engagement element (SHIFT) and the gear ratio (RATIO) at each gear, and the meshing efficiency (GEAR EFF) of the planetary gears. Regarding the gear ratio (RATIO) indicating the performance, the gear ratio step value (STEP), and the gear ratio width (RANGE), the gear ratio width (RANGE) becomes a value suitable for 9.86 and 9AT. The value (STEP) is also almost appropriate. The meshing efficiency (GEAR EFF) of the planetary gears is higher than 99% except for the first forward speed (1st) 98.4% and the third forward speed (3rd) 98.2%. It is considerably superior to the conventional 8AT. Since the shift mode such as the flow of power in each shift stage has already been disclosed in Patent Document 2, description thereof will be omitted.

  Fig. 14 "C3-1-1, 9AT, C3-1-2, 11AT (FR)" is a conceptual design of the FR specification shown on the left of Fig. 1 for passenger cars with the input power from the prime mover set to 300 Nm. It is a structural drawing. The arrangement structure of the fourth brake (B4) of FIG. 2 showing 11AT described later is shown in FIG. 14 for reference. In FIG. 14, a prime mover (not shown) is arranged on the left front side of the transmission, and power is input to the transmission via the torque converter (200a). The transmission case 1 is arranged as one body, and the third planetary gear train (30), the first and second clutches (C1, C2), and the fourth planetary gear train (40 ) And the first and second brakes (B1, B2) arranged on the outer periphery thereof, or the fourth brake (B4) used when 11AT is used, and the second planetary gear train ( 20), a first planetary gear train (10), a third clutch (C3), and a main transmission mechanism including a third brake (B3) in which a friction member is arranged around the outer periphery of the second planetary gear train (20). Arranged in order.

  A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened to the front portion of the transmission case 1 with bolts, and a gear shift for fixing the wheel stator of the torque converter (200a) to the holding member 2a. The cylindrical member 2b, which is cylindrically extended in the machine inward direction, is fastened with a bolt. Bushings 4a and needle roller roller bearings 4b are arranged on the inner circumferences of both ends of the cylindrical member 2b to support the input shaft 3a. An output shaft 3c rotatably supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e is arranged at the rear portion of the transmission case 1, and an input shaft 3a is arranged by a needle roller roller bearing 4d arranged on the inner circumference of the output shaft 3c. To support. Here, since the integrated input shaft 3a is pivotally supported at three points, the distance between the respective pivots is short, and the input shaft 3a is less susceptible to vibration due to unbalance in the circumferential direction, so that the torque transmission capacity is reduced. The diameter of the portion arranged around the input shaft 3a can be reduced, which leads to the weight reduction of the transmission.

  On the outer periphery of the cylindrical member 2b, a third planetary gear train (30) and a first and second clutches (C1, C2), which serve as a sub front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged from the left front. . The third planetary gear train (30) decelerates the rotation of the input shaft 3a and selectively transmits it to the fourth planetary gear train (40) which serves as a main front transmission mechanism. The third planetary carrier (P3) of the third planetary gear train (30) is rotatably held on the outer periphery of the cylindrical member 2b by a bush 4j press-fitted into the inner diameter of the right side member, which is cylindrically extended. Further, the third sun gear (S3) is rotatably held by the bush 4i press-fitted on the inner circumference of the third planet carrier (P3) on the outer circumference which is cylindrically extended on the inner circumference of the right side member of the third planet carrier (P3). A thin plate-shaped brake hub is welded to the front side of the third sun gear (S3), extended to the outer periphery of the third planetary gear train (30), and the friction member of the first brake (B1) is locked. The left side member of the third planet carrier (P3) extends rearward as the second connecting member (8) along the inner side of the transmission case 1 through the outer periphery of the third planetary gear train (30). The third ring gear (R3) is connected to the input hub welded to the input connecting member (Y). The input connecting member (Y) is spline-connected to the input shaft 3a behind the cylindrical member 2b, and extends forward in a cylindrical shape along the outer circumference of the cylindrical member 2b, and is rotatably held on the outer circumference of the cylindrical member 2b by a bush 4c. To be done. Further, a hub connected to the third ring gear (R3) is welded to the front of the input connecting member (Y), and a clutch hub that locks the friction member of the first clutch (C1) is also welded.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member 2b behind the third planetary gear train (30). The first clutch (C1) can connect the input shaft 3a and the fourth planet carrier (P4) of the fourth planetary gear train (40), and the second clutch (C2) connects the fourth planet carrier (P4) and the fourth planet carrier (P4). A second connecting member (8) that connects to the ring gear (R4) can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The left side member of the fourth planetary carrier (P4) constituting the fourth planetary gear train (40) is welded to the rear end of the fourth planetary gear train (P4). In front of the dual clutch connecting member (X), the friction of the first clutch (C1) arranged on the radial inner circumference of the dual clutch connecting member (X) and the clutch hub welded to the input connecting member (Y). A servo mechanism for pressing the member is disposed, and a second outer side of the fourth planetary carrier (P4) is disposed on the left side member and the outer periphery of the dual clutch coupling member (X) in the radial direction and the second coupling member (8). A servo mechanism for pressing the friction member of the clutch (C2) is provided. A partition plate is provided between the double clutch coupling member (X) and the left side member of the fourth planet carrier (P4) to supply the servomechanism arranged above the left side member of the fourth planet carrier (P4) in the radial direction. A passage for the working oil of the piston and the hydraulic cancel oil is formed. The servo mechanism of the second clutch (C2) arranged on the upper side in the radial direction of the dual clutch coupling member (X) and the servo mechanism of the first clutch (C1) arranged on the lower side in the radial direction include the outer circumference of the cylindrical member 2b. Is supplied with hydraulic oil and hydraulic cancel oil through an oil passage sealed by a piston ring provided in the input connecting member (Y). Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism for the second clutch (C2) of the left side member of the fourth planetary carrier (P4), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 6" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

A fourth planetary gear train (40) serving as a main front transmission mechanism of the FRONT GEAR (front transmission mechanism) is arranged behind the cylindrical member 2b and the first and second clutches (C1, C2). The fourth planetary gear train (40) selectively uses five types of rotations (up to one type for 11AT to increase six types) for the first and second planetary gear trains (10, 20) serving as the main transmission mechanism. Communicate to. The fourth ring gear (R4) of the fourth planetary gear train (40) is arranged along the inner side of the transmission case 1 on the left side member of the third planetary carrier (P3) of the third planetary gear train (30). A spline connection is made at the rear side of the extended material as the connection member (8). The left side member of the fourth planet carrier (P4) supporting the planet pinion gear that meshes with the fourth ring gear (R4) serves as a clutch cover for the second clutch (C2) and is welded to the dual clutch connecting member (X). The right side member is extended on the outer circumference of the fourth ring gear (R4). The output fourth sun gear (S4) is spline-connected to the first connecting member (7) at the inner circumference.

  The first brake (B1) enables braking of the third sun gear (S3) of the third planetary gear train (30). A thin plate brake hub welded in front of the third sun gear (S3) is extended to the outer periphery of the third planetary gear train (30), and one side of the friction of the first brake (B1) is applied to the spline formed on the outer periphery. The member is locked. The other friction member of the first brake (B1) is locked to a spline formed in front of the transmission case 1, and a piston is provided in a hydraulic chamber of a holding member 2a fixed to the front of the transmission case 1 by a bolt. The return spring is held and the hydraulic servo of the first brake (B1) is formed.

  The second brake (B2) enables braking of the fourth planet carrier (P4) of the fourth planetary gear train (40). The right side member of the fourth planet carrier (P4) is extended to the outer periphery of the fourth ring gear (R4), and one friction member of the second brake (B2) is locked to the spline formed on the outer periphery. The other friction member of the second brake (B2) is a fourth ring gear (R4) to which a spline for locking the friction member of the first brake (B1) formed in front of the transmission case 1 is extended. A brake case, which is locked on the outer periphery and holds a piston and a return spring that form a hydraulic servo of the second brake (B2), is fixed to the transmission case 1 in front of the friction member with a retaining ring.

  FIG. 14 is a structural diagram of a schematic diagram of the FR specification of FIG. 1, which is 9AT, but connects the third planet carrier (P3) and the fourth ring gear (R4) of the power train to the second connecting member (8 ) Can be braked by the fourth brake (B4), it becomes 11AT of FIG. 2 described later, and the arrangement structure of the fourth brake (B4) is shown for reference. One friction member of the fourth brake (B4) is locked to the spline formed on the outer periphery of the second connecting member (8), and the other friction member is the first brake formed in front of the transmission case 1. It is locked to the spline that is a stretched material that locks the friction member (B1). The brake case, which holds the piston and the return spring that form the hydraulic servo of the fourth brake (B4), is integrated with the brake case of the second brake (B2) behind the friction member of the fourth brake (B4) and the transmission case. Fixed to 1 with a retaining ring. Note that the fourth brake (B4) is not necessary in FIG. 1 showing 9AT.

  The second planetary gear train (20) arranged behind the fourth planetary gear train (40) has the first connection in the first forward speed (1st) to the fourth forward speed (4th) and in the reverse (Rev1). The rotation input from the member (7) and the input shaft 3a via the first planetary gear train (10) is decelerated and output. The second sun gear (S2) of the second planetary gear train (20) is integrally molded with the first sun gear (S1) of the first planetary gear train (10) and is rotatable around the outer periphery of the input shaft 3a by the needle roller roller bearing 4h. And is forwardly extended as a first connecting member (7) to be spline-connected with the fourth sun gear (S4) of the fourth planetary gear train (40). The planetary pinion gear that meshes with the second sun gear (S2) is supported by the second planetary carrier (P2), the right side member is welded to the output shaft 3c, and the output hub is arranged on the outer periphery of the first planetary gear train (10). The second ring gear (R2) meshed with the planetary pinion gear is a plate that is axially fixed to the left end tooth portion by a retaining ring and is spline-connected to the left side of the second planetary carrier (P2). The material is extended to the inner circumference of the side member and is rotatably arranged with its axial direction restricted by a thrust needle bearing between the second member and the second sun gear (S2) of the second planetary gear train (20). The friction member of the third brake (B3) is locked on the outer periphery of the second ring gear (R2), but the center of gravity of both is within the width of the planet pinion gear of the fourth planetary gear train (40). Since the radial load of the second ring gear (R2) and the friction member is received by the two planet carrier (P2), the radial bearing dedicated to the second ring gear (R2) is not required.

  The first planetary gear train (10) arranged behind the second planetary gear train (20) is connected to the first connecting member (7) in the fourth forward speed (4th) to the ninth forward speed (9th). The rotation input from the shaft 3a is input to the first sun gear (S1) and the first planet carrier (P1) and output from the first ring gear (R1). The planetary pinion gear meshing with the first sun gear (S1) integrally formed with the second sun gear (S2) is supported by the first planetary carrier (P1), and the right side member is splined to the input shaft 3a. The first ring gear (R1) meshing with the planetary pinion gear is arranged behind the first planetary gear train (10), is axially supported by the bush 4g on the inner circumference, and locks the friction member of the third clutch (C3). The clutch hub is welded.

  The third clutch (C3) can connect the first ring gear (R1) of the first planetary gear train (10) and the output shaft 3c. One friction member of the third clutch (C3) is locked to the spline formed on the outer circumference of the clutch hub of the third clutch (C3), and formed on the inner circumference of the output hub 9 welded to the output shaft 3c. The other friction member of the third clutch (C3) is locked to the spline. The piston of the third clutch (C3) and the return spring are held in front of the output shaft 3c, and a hydraulic servo of the third clutch (C3) is formed. The output shaft 3c is supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e at the rear of the transmission case 1, and is sealed from the transmission case 1 by a seal ring between the needle roller roller bearing 4f and the deep groove ball bearing 4e. The hydraulic oil of the third clutch (C3) is supplied through the oil passage. A parking gear 6a is formed on the outer circumference of the flange of the output shaft 3c.

  The third brake (B3) enables braking of the second ring gear (R2) of the second planetary gear train (20). The second ring gear (R2) has a spline formed on the outer peripheral portion thereof to lock one friction member of the third brake (B3). A spline is formed on the inner circumference of the transmission case 1, and the other friction member of the third brake (B3) is locked. Further, the piston and the return spring of the third brake (B3) are held in the transmission case 1 behind the flange of the output shaft 3c, and a hydraulic servo of the third brake (B3) is formed. The piston of the third brake (B3) presses the friction member of the third brake (B3) on the outer circumference of the output hub 9.

  The planetary gear train of FIG. 1 is characterized in that the input power of the third planetary gear train (30) and the fourth planetary gear train (40) is input to a ring gear having a large diameter. As a result, the load on the planetary gears of the third planetary gear train (30) and the fourth planetary gear train (40) can be reduced, and the tooth gap can be set to a small value. Combined with the built-in construction, the FRONT GEAR (front gearbox) can be made compact. Further, in the structure diagram of FIG. 14, the arrangement of the constituent parts described in “claim 1” and the structure of the first and second clutches (C1, C2) described in “claim 5” are shown in FIG. It is about 5% simpler and more compact than the structure described in. In FIG. 14, the structure for injecting the cooling oil from the end plates of the first, second and fourth brakes (B1, B2, B4) is the structure proposed by the applicant of the present application in Japanese Patent Laid-Open No. 2009-236234. It is also possible to directly connect the prime mover and the input shaft with a rotation fluctuation absorbing damper without using a converter, and to use it as a starting device for sliding the first brake (B1) to perform creeping or smooth starting of the vehicle. It has a simple structure that produces less heat and has a better cooling effect than a clutch that requires the same torque capacity as the input shaft as the starting device used.

  Fig. 20 "C1-1-1, 9AT (FF)" is a structural diagram conceptually designed with the input power from the prime mover set to 300 Nm, which is a schematic diagram of the schematic diagram on the right of the FF specification suitable for passenger cars in Fig. 1, with a total length of 400 mm. It will be about. In FIG. 20, power is transmitted from a prime mover (not shown) on the left front side of the drawing to a torque converter 200a, which is a fluid transmission device, and is guided to the input shaft 3 of the transmission. The housing that houses the entire transmission is composed of a front torque converter case 1a, a transmission case 1b that serves as the transmission main body, and a transmission case 1c that closes the rear portion. The housing is arranged in parallel with the input shaft 3 and the input shaft 3. It serves as a mother body that pivotally supports the relay shaft 17 and the differential device 9 including the output shaft. The input shaft 3 is rotatably supported at one end by a bearing 4a disposed on the holding members 2a and 2b, and at the other end by a bearing 4b disposed at the transmission case 1c. The holding members 2a and 2b are partition walls separating the dry torque converter case 1a and the wet transmission case 1b, and support the torque converter 200a to hold the charging pump of the transmission. The relay shaft 17 has one end supported by a bearing 4g arranged in the torque converter case 1a and the other end supported by a bearing 4f arranged in the transmission case 1b, and meshes with the output counter gear 5 coaxial with the input shaft 3. The counter gear 6 is connected by a spline, and a pinion gear that transmits power to a differential device 9 including an output shaft is integrally formed. The output shaft portion serves as a carrier of the differential device 9, one end of which is a bearing 4i arranged in the torque converter case 1a, and the other end of which is rotatably supported by a bearing 4h arranged in the transmission case 1b, which is integrated with the pinion gear. A large gear 18 meshing with the relay shaft 17 is fastened with a bolt. As is well known, the differential device 9 comprises a pinion gear and a side gear, and the side gear is connected to the output shaft of the automatic transmission.

  In FIG. 20, the transmission case 1b has a transmission case 1c fastened to the rear by bolts, and an L-shaped cylindrical partition wall is integrally provided in the transmission case 1b at the center in the axial direction. The angular bearing 4e is fixed to the outer periphery of the cylindrical portion by screws with the back faces aligned, and the output counter gear 5 is pivotally supported. A third clutch (C3) and a first planetary gear train (10) are arranged in front of the partition wall from the partition wall side, and behind the output counter gear 5, a second planetary gear train (in order from the output counter gear 5 side ( 20), a fourth planetary gear train (40), first and second clutches (C1, C2), and a third planetary gear train (30). On the outer circumference of the fourth planetary gear train (40), the first and second clutches (C1, C2), and the third planetary gear train (30), the second brake (B2) and the first brake are sequentially arranged from the output counter gear 5 side. (B1) is arranged to form a front transmission mechanism, a third brake (B3) is arranged on the outer periphery of the second planetary gear train (20), and a third clutch (C3) and a first planetary gear train (10) are arranged. Together with this, it constitutes the main transmission mechanism.

  The transmission case 1c for closing the rear part of the transmission has an integral inner peripheral portion protruding forward in a cylindrical shape, and a needle roller roller bearing 4b is arranged at the inner peripheral end portion of the protruding cylindrical member to provide the input shaft 3 To support. Therefore, the shaft support of the input shaft 3 is shortened, and rotational vibration can be suppressed. A third planetary gear train (30) and a first and a second clutch (C1, C2), which serve as a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged on the outer periphery of the protruding cylindrical member from the right rear. It The third planetary gear train (30) decelerates the rotation of the input shaft 3 and selectively transmits it to the fourth planetary gear train (40) serving as a main front transmission mechanism. The third planet carrier (P3) of the third planetary gear train (30) is freely rotatable around the outer circumference of the cylindrical member of the transmission case 1c with a bush press-fitted into the inner diameter of the left side member, which is cylindrically extended. Further, the third sun gear (S3) is rotatably held on the outer periphery which is cylindrically extended on the inner periphery of the left side member of the third planet carrier (P3) by a bush press-fitted on the inner periphery. It A thin plate-shaped brake hub is welded to the rear side of the third sun gear (S3), extended to the outer periphery of the third planetary gear train (30), and the friction member of the first brake (B1) is locked. The right side member of the third planet carrier (P3) passes through the outer periphery of the third planetary gear train (30) and is extended backward as a second connecting member (8). The third ring gear (R3) is welded to the input connecting member (Y) and spline-connected to a clutch hub that locks one friction member of the first clutch (C1). The input connecting member (Y) is spline-connected to the input shaft 3 in front of the cylindrical member of the transmission case 1c, and extends rearward in a cylindrical shape along the outer circumference of the cylindrical member so as to be rotatable around the outer circumference of the cylindrical member by a bush. Retained.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member of the transmission case 1c in front of the third planetary gear train (30). The first clutch (C1) can connect the input shaft 3 and the fourth planet carrier (P4) of the fourth planetary gear train (40), and the second clutch (C2) connects the fourth planet carrier (P4) and the fourth planet carrier (P4). A second connecting member (8) that connects to the ring gear (R4) can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The right side member of the fourth planetary carrier (P4) forming the fourth planetary gear train (40) is welded to the front end of the fourth planetary gear train (P4). The right side member of the fourth planet carrier (P4) is a servo that presses the radial outer periphery of the dual clutch connecting member (X) and the friction member of the second clutch (C2) arranged in the second connecting member (8). A mechanism is arranged, and between the double clutch connecting member (X) and the right side member of the fourth planetary carrier (P4), there is a hydraulic fluid for the piston which is supplied to the servo mechanism arranged on the upper side in the radial direction by a partition plate. The hydraulic oil and the hydraulic cancel oil are supplied from the outer periphery of the cylindrical member of the transmission case 1c through the passage of the hydraulic cancel oil through the oil passage sealed by the piston ring provided in the input coupling member (Y). Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism of the second clutch (C2) of the right side member of the fourth planetary carrier (P4), and seals the hydraulic chamber. Hydraulic oil and hydraulic cancel oil are supplied from the outer circumference of the cylindrical member of the transmission case 1c to the servo mechanism of the second clutch (C2) through an oil passage sealed by a piston ring provided in the input coupling member (Y). To be done. Behind the double clutch connecting member (X), a clutch hub that locks the friction member of the first clutch (C1) is welded to the input connecting member (Y) and serves as a servo mechanism that presses the friction member. The cover, the piston, and the return spring are held by the input connecting member (Y). The servo mechanism of the first clutch (C1) is supplied with working oil and hydraulic cancel oil from the outer circumference of the cylindrical member of the transmission case 1c through an oil passage sealed by a piston ring provided in the input coupling member (Y). It The servo mechanism provided on the input coupling member (Y) of the FF specification first clutch (C1) is the first clutch provided on the FR specification dual clutch coupling member (X) described in paragraph “0043”. Although different from the servo mechanism of (C1), the servo mechanism of the first clutch (C1) may be provided on either side. Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism of the second clutch (C2) of the right side member of the fourth planetary carrier (P4), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 6" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

A fourth planetary gear train (40) serving as a main front transmission mechanism of the FRONT GEAR (front transmission mechanism) is arranged in front of the cylindrical member of the transmission case 1c and the first and second clutches (C1, C2). It The fourth planetary gear train (40) selectively transmits five types of rotations to the first and second planetary gear trains (10, 20) which are main transmission mechanisms. The fourth ring gear (R4) of the fourth planetary gear train (40) is a second connecting member (8) integrated with the right side member of the third planet carrier (P3) of the third planetary gear train (30). The material is extended forward and splined. The right side member of the fourth planet carrier (P4) supporting the planet pinion gear that meshes with the fourth ring gear (R4) serves as a clutch cover for the second clutch (C2) and is welded to the dual clutch connecting member (X). The left side member is extended on the outer periphery of the fourth ring gear (R4). The output fourth sun gear (S4) is spline-coupled to the first coupling member (7) which is axially supported by the bush on the outer periphery of the input shaft 3 on the inner periphery.

  The first brake (B1) enables braking of the third sun gear (S3) of the third planetary gear train (30). A thin plate-shaped brake hub welded to the rear of the third sun gear (S3) is extended to the outer periphery of the third planetary gear train (30), and one side of the friction of the first brake (B1) is applied to the spline formed on the outer periphery. The member is locked. The other friction member of the first brake (B1) is locked to a spline formed in front of the transmission case 1c, and a piston and a return spring are held in a hydraulic chamber formed in the transmission case 1c. A hydraulic servo of the brake (B1) is formed.

  The second brake (B2) enables braking of the fourth planet carrier (P4) of the fourth planetary gear train (40). The left side member of the fourth planetary carrier (P4) forming the fourth planetary gear train (40) is extended to the outer periphery of the fourth ring gear (R4), and the second brake (B2) is attached to the spline formed on the outer periphery. 1) one of the friction members is locked. The other friction member of the second brake (B2) is locked to a spline formed on the inner circumference of the transmission case 1b. A brake case that holds a piston forming a hydraulic servo of the second brake (B2) and a return spring is fixed to the transmission case 1b in front of the friction member of the second brake (B2) with a retaining ring.

The second sun gear (S2) of the second planetary gear train (20) arranged in front of the fourth planetary gear train (40) is spline-coupled to the first coupling member (7). The planet pinion gear that meshes with the second sun gear (S2) is supported by the second planet carrier (P2), and the left side member is spline-connected to the output counter gear 5 and fixed by the retaining ring. The second ring gear (R2) that meshes with the planetary pinion gear has a plate in which the right end tooth portion is axially fixed by a retaining ring and is spline-connected to the second sun gear (S2) of the second planetary gear train (20). It is rotatably arranged between the second sun gear (S2), the fourth sun gear (S4) of the fourth planetary gear train (40), and the first connecting member (7) with its axial direction restricted by the thrust needle bearing. To be done.

  The third brake (B3) enables braking of the second ring gear (R2) of the second planetary gear train (20). The second ring gear (R2) has a spline formed on the outer peripheral portion thereof to lock one friction member of the third brake (B3). A spline is formed on the inner circumference of the transmission case 1b to lock the other friction member of the third brake (B3). Further, behind the friction member of the third brake (B3), there is a brake case which is integrated with the brake case of the second brake (B2) and forms a hydraulic servo of the third brake (B3) and a return spring. Will be distributed.

  The output counter gear 5 arranged in front of the second planetary gear train (20) is fixed to the outer periphery of the cylindrical portion of the L-shaped cylindrical partition wall arranged at the axial center of the transmission case 1b by screws. It is pivotally supported by the back-aligned angular bearing 4e. The left side member of the second planetary carrier (P2) of the second planetary gear train (20) is spline-connected to the right outer periphery of the output counter gear 5 and fixed by a retaining ring.

  A third clutch (C3) arranged in front of a partition wall arranged in the axial center of the transmission case 1b holds a piston and a return spring to form a hydraulic servo of the third clutch (C3). The inner peripheral cylindrical member passes through the inner peripheral surface of the partition wall and extends to the inner periphery of the left side member of the second planetary carrier (P2) of the second planetary gear train (20) and is spline-connected to the left side member, and the output counter It is connected to the gear 5. The inner peripheral cylindrical member of the clutch cover of the third clutch (C3) is pivotally supported by the bush on the outer periphery of the input shaft 3 by the bush on the outer periphery of the first connecting member (7), and the transmission case is also supported. The hydraulic oil of the third clutch (C3) is supplied from the oil passage opened on the inner circumference of the partition wall of 1b through the oil passage sealed by the seal ring. A spline is formed on the inner periphery of the outer peripheral plate of the clutch cover of the third clutch (C3), one friction member of the third clutch (C3) is locked, and the spline is arranged in front of the third clutch (C3). The other friction member is locked to the spline on the outer peripheral portion of the clutch hub welded to the first ring gear (R1) of the first planetary gear train (10).

  In the first planetary gear train (10) arranged in front of the third clutch (C3), the left side member of the first planetary carrier (P1) is extended to the inner peripheral side of the first sun gear (S1) to form an input shaft. The first sun gear (S1) is spline-coupled to the first coupling member (7) axially supported by the bushing on the outer circumference of the input shaft 3 on the inner circumferential side of the third clutch (C3). The first ring gear (R1) is welded to the clutch hub of the third clutch (C3) on the right side, and the clutch hub is pivotally supported by the bush on the outer periphery of the first sun gear (S1). The center of gravity of the clutch hub of the third clutch (C3) that locks the friction member outside the first ring gear (R1) and the first planetary gear train (10) is the planetary pinion gear of the first planetary gear train (10). Since it is out of the width, it requires a dedicated radial bearing.

  In the structure of FIG. 20, the right side member of the fourth planet carrier (P4) of the fourth planetary gear train (40) serves as the hydraulic chamber of the second clutch (C2). And the structure of "claim 2" which defines the arrangement of the input connecting member (Y) and the double clutch connecting member (X), and the first and second clutches (C1, C2) each having a friction member in the radial direction. According to the structure of "Claim 1", the third planetary gear train (30) is arranged so as to be rotatably arranged on the outer periphery of the cylindrical member so as to be overlapped in two stages to supply hydraulic oil from the outer periphery of the cylindrical member. The transmission mechanism is made compact, the first planetary gear train (10) and the second planetary gear train (20) are arranged so as to sandwich the output counter gear 5, and the second ring gear (R2) of the second planetary gear train (20) is arranged. The thrust needle bear between the second sun gear (S2) and the fourth sun gear. The main transmission mechanism is made compact by the structure of "Claim 3" in which the main shaft is supported by a ring and the structure of "Claim 4" in which the hydraulic oil of the third clutch (C3) is supplied from the partition wall, and it is realized like an FF. It is a thing.

<1-TYPE>"C3-1-2,11AT" FIG. 2, (FIG. 14), FIG.
FIG. 2 shows a schematic diagram of two types of C3 type 11AT, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift speed, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 2, the velocity diagram is divided into a MAIN GEAR (main transmission mechanism) and a FRONT GEAR (front transmission mechanism), and the FRONT GEAR (front transmission mechanism) is a main front transmission mechanism and a sub front transmission mechanism. It is divided into a stationary transmission mechanism. The speed diagram of the MAIN GEAR (main transmission mechanism) is such that the first, second, third, and fourth constituent elements are arranged in order from the right side of the figure, and the speed diagram of the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). The fifth, sixth, and seventh constituent elements (A, B, C) are arranged in order from the right in the figure, and the speed diagram of the main front transmission mechanism is the eighth, ninth, and tenth constituent elements in order from the right in the figure. (D, E, F) are arranged, the sixth component (B) and the eighth component (D) are connected by the second connecting member (8), and the first component and the tenth component (F). Are connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed. This arrangement is exactly the same as in FIG.

  In the two types of schematic diagrams in Fig. 2, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. The difference from FIG. 1 is only the main front transmission mechanism having the eighth, ninth and tenth constituent elements (D, E, F), so the main front transmission mechanism will be described. In FIG. 1, the fourth ring gear (R4) of the fourth planetary gear train (40), the fourth planetary carrier (P4), and the fourth sun gear (S4) of the fourth planetary gear train (40) constitute the eighth, ninth, and ninth gears. In contrast to the ten constituent elements (D, E, F), in FIG. 2, the fourth sun gear (S4) and the fourth planet carrier (P4) of the fourth planetary gear train (40) forming the main front transmission mechanism. , The fourth ring gear (R4) is the eighth, ninth, and tenth constituent elements (D, E, F), and the fourth sun gear (S4) and the fourth ring gear (R4) are interchanged. Then, the fourth sun gear (S4), which is the eighth constituent element (D), can be braked by the fourth brake (B4), so that one more speed-up rotation of the input shaft than that in FIG. 1 is obtained. . As a matter of course, like the structure in which the fourth ring gear (R4) described as a reference in FIG. 14 can be braked by the fourth brake (B4), the fourth ring gear (which is the eighth component (D) in FIG. 1 ( R4) may be braked by the fourth brake (B4).

  In Patent Document 2, since the second clutch (C2) plays a role of integrating the fourth planetary gear train (40), any two of the three constituent elements of the fourth planetary gear train (40) are used. It suffices to connect the second clutch (C2). The first clutch (C1) can connect the input shaft to the fourth planetary carrier (P4) of the fourth planetary gear train (40), and the first and second clutches (C1, C2) are two-story double clutches. Therefore, the second clutch (C2) has a simpler and more compact structure in which the fourth planet carrier (P4) and the fourth ring gear (R4) can be connected. Therefore, the arrangement of the FRONT GEAR (front transmission mechanism) is, in order from the MAIN GEAR (main transmission mechanism) side, the fourth planetary gear train (40), the first and second clutches (C1, C2), and the third planetary gear in FIG. The gear train (30) is arranged in this order, but in FIG. 2, the fourth ring gear (R4) of the fourth planetary gear train (40) is connected to the MAIN GEAR (main transmission mechanism), so the first and second clutches are connected. (C1, C2), the fourth planetary gear train (40), and the third planetary gear train (30) are arranged in this order. The cylindrical member that supplies hydraulic oil to the first and second clutches (C1, C2) further approaches the MAIN GEAR (main transmission mechanism), and all the FRONT GEAR (previous transmission mechanism) is provided on the outer periphery of the cylindrical member. The parts are arranged. The first, second and fourth brakes (B1, B2, B4) are provided on the outer circumferences of the first and second clutches (C1, C2), the fourth planetary gear train (40) and the third planetary gear train (30). ) Is arranged.

  In the schematic diagram of the FF specification shown in the right diagram of FIG. 2 and the table showing the gear ratio of each planet gear, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is shown. Is 2.300, and the gear ratio of the ring gear and the sun gear of the second planetary gear train (20) is 3.000, and the gear ratio of the first planetary gear train (10) of FIG. 1 is 3.000 and the second planetary gear. It is significantly different from the gear ratio (2.667) of the gear train (20). This is to make the gear ratios of 9AT of FIG. 1 and 11AT of FIG. 2 appropriate, respectively, and in FIG. 2, since the gear ratio of the first planetary gear train (10) is small at 2.300, the first ring gear ( The outer diameter of R1) becomes smaller. Therefore, the friction member of the third clutch (C3) can be arranged on the outer periphery of the first ring gear (R1). Also, since the first, second, third, and fourth brakes (B1, B2, B3, B4) are arranged behind the partition wall arranged in the center of the transmission case, the output counter gear is It was arranged on the front side of the partition wall on the side opposite to.

  In the table showing the gear ratios of "C3-1-2, 11AT" in FIG. 2, the gear ratios are 6.721 to 0,615, the forward deceleration stage is 8 and the speed increasing stage is 3, which is a practical problem. No gear ratio is obtained. However, the range of the gear ratio (RANGE) becomes slightly smaller for 10.93 and 11AT, and the step value (STEP) of the gear ratio becomes a little cross from the low speed region to the medium speed region. Although the meshing efficiency (GEAR EFF) of the planetary gears is slightly inferior to that of "C3-1-1, 9AT" in Fig. 1, it is the highest level for AT in this class. This is because only the fourth brake (B4) is added to “C3-1-1, 9AT” in FIG. 1 and the shift mode is exactly the same as in FIG.

  In paragraph “0006”, “For a light-weight passenger car, a Gear Range of about 9 is sufficient, and if the step value of the gear ratio, the weight of the transmission, and the transmission efficiency are appropriate, the number of gears is 7, 8 (7, 8). 8AT) is sufficient. ”As described above, this gear train is suitable for commercial vehicles such as trucks and buses because it does not require 11AT except for special use in passenger cars.

    FIG. 21 “C1-1-2, 11AT (FF)” is a structural diagram conceptually designing the schematic diagram on the right side of the FF specification of FIG. 2 with the input power from the prime mover set to 300 Nm, and the total length is 9AT (of FIG. FF), which is about 400 mm. The input structure from the torque converter 200a and the output structure for outputting to the differential device 9 via the counter gear 6 meshing with the output counter gear 5 are the same as those in FIG. The housing that houses the entire transmission is composed of a front torque converter case 1a, a transmission case 1b that serves as a transmission main body, and a transmission case 1c that closes the rear portion. A character-shaped cylindrical partition wall is integrally provided, and an angular bearing 4e is fixed to the outer periphery of the inner peripheral cylindrical portion by screws with their backs aligned, and axially supports the output counter gear 5. The output counter gear 5, the third clutch (C3) and the first planetary gear train (10) are arranged in front of the partition wall from the partition wall side, and behind the partition wall, the second planetary gear train ( 20), first and second clutches (C1, C2), a fourth planetary gear train (40), and a third planetary gear train (30). The second brake (B2) and the fourth brake (B4) are sequentially arranged on the outer circumferences of the first and second clutches (C1 and C2), the fourth planetary gear train (40) and the third planetary gear train (30) from the partition side. And a first brake (B1) are arranged to form a front transmission mechanism, a third brake (B3) is arranged on the outer periphery of the second planetary gear train (20), and a third clutch (C3) and a first planetary gear are arranged. Together with the row (10) constitutes the main transmission mechanism.

  The transmission case 1c for closing the rear part of the transmission has an integral inner peripheral portion protruding forward in a cylindrical shape, and a needle roller roller bearing 4b is arranged at the inner peripheral end portion of the protruding cylindrical member to provide the input shaft 3 To support. Therefore, the shaft support of the input shaft 3 is shortened, and rotational vibration can be suppressed. A third planetary gear train (30), a fourth planetary gear train (40), a first planetary gear train (40), a first planetary gear train (40), and a second planetary gear train (30), which serve as a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged on the outer periphery of the protruding cylindrical member from the right rear. Clutch (C1, C2) is arranged. Since the fourth planetary gear train (40) is arranged on the outer periphery of the cylindrical member as compared with 9AT in FIG. 20, the cylindrical member projects forward by that amount, further shortening the axial support interval of the input shaft 3. The third planetary gear train (30) decelerates the rotation of the input shaft 3 and selectively transmits it to the fourth planetary gear train (40) serving as a main front transmission mechanism. The third planet carrier (P3) of the third planetary gear train (30) is freely rotatable around the outer circumference of the cylindrical member of the transmission case 1c with a bush press-fitted into the inner diameter of the left side member, which is cylindrically extended. Further, the third sun gear (S3) is rotatably held on the outer periphery which is cylindrically extended on the inner periphery of the left side member of the third planet carrier (P3) by a bush press-fitted on the inner periphery. It A thin plate-shaped brake hub is welded to the rear side of the third sun gear (S3), extended to the outer periphery of the third planetary gear train (30), and the friction member of the first brake (B1) is locked. The right side member of the third planet carrier (P3) is extended as a second connecting member (8) around the outer periphery of the third planetary gear train (30). The third ring gear (R3) is spline-connected to the input hub welded to the input connecting member (Y). The input connecting member (Y) is spline-connected to the input shaft 3 in front of the cylindrical member of the transmission case 1c, and extends rearward in a cylindrical shape along the outer circumference of the cylindrical member so as to be rotatable around the outer circumference of the cylindrical member by a bush. Retained.

    The fourth planetary gear train (40) arranged in front of the third planetary gear train (30) selectively rotates six types of rotations into the first and second planetary gear trains (10, 20) which are main transmission mechanisms. Communicate to. The fourth sun gear (S4) of the fourth planetary gear train (40) is axially supported by a bush on the outer circumference of the input coupling member (Y) on the inner circumference, and the brake hub of the fourth brake (B4) is welded to the brake hub. The second connecting member (8) extended to the outer periphery of the third planetary gear train (30) is spline-connected and the friction member of the fourth brake (B4) is locked. The fourth planet carrier (P4) supporting the planet pinion gear that meshes with the fourth sun gear (S4) has a right side member extended to the outer periphery of the fourth planetary gear train (40) and a left side member provided with the second clutch (C2). ), And is welded to the double clutch connecting member (X).

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member of the transmission case 1c in front of the fourth planetary gear train (40). The first clutch (C1) can connect the input shaft 3 and the fourth planet carrier (P4) of the fourth planetary gear train (40), and the second clutch (C2) connects the fourth planet carrier (P4) and the fourth planet carrier (P4). The ring gear (R4) can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The left side member of the fourth planetary carrier (P4) constituting the fourth planetary gear train (40) is welded to the rear end of the fourth planetary gear train (P4). The left side member of the fourth planet carrier (P4) is a servo that presses the radial outer periphery of the dual clutch coupling member (X) and the friction member of the second clutch (C2) arranged in the fourth ring gear (R4). A mechanism is provided, and working oil and hydraulic cancel oil are supplied from the outer periphery of the cylindrical member of the transmission case 1c through an oil passage sealed by a piston ring provided in the input coupling member (Y). A clutch cover, which serves as a servo mechanism for pressing the friction member of the first clutch (C1), a piston, and a return spring are splined to the front end of the input coupling member (Y) in front of the dual clutch coupling member (X). Held by the clutch hub. In the servo mechanism of the first clutch (C1), hydraulic oil and hydraulic cancel oil are supplied from the oil passage sealed from the outer circumference of the cylindrical member of the transmission case 1c by the piston ring provided in the input coupling member (Y). Supplied to the hub. The servo mechanism provided on the input coupling member (Y) of the FF specification first clutch (C1) is the first clutch provided on the FR specification dual clutch coupling member (X) described in paragraph “0043”. Although different from the servo mechanism of (C1), the servo mechanism of the first clutch (C1) may be provided on either side, but the FF specifications of FIGS. 20 and 21 are more compact. Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism for the second clutch (C2) of the left side member of the fourth planetary carrier (P4), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 6" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved. The fourth ring gear (R4) of the fourth planetary gear train (40) is spline-connected to the first connecting member (7) having an inverted L-shaped cross section, and is freely rotatable around the outer circumference of the input shaft 3 with a bush. Is pivotally supported.

  The first brake (B1) enables braking of the third sun gear (S3) of the third planetary gear train (30). A thin plate-shaped brake hub welded to the rear of the third sun gear (S3) is extended to the outer periphery of the third planetary gear train (30), and one side of the friction of the first brake (B1) is applied to the spline formed on the outer periphery. The member is locked. The other friction member of the first brake (B1) is locked to a spline formed in front of the transmission case 1c, and a piston and a return spring are held in a hydraulic chamber formed in the transmission case 1c. A hydraulic servo of the brake (B1) is formed.

  The fourth brake (B4) is arranged in front of the first brake (B1) in opposition to the right side member of the third planet carrier (P3) arranged on the outer periphery of the third planetary gear train (30). The second connecting member (8) that connects the fourth sun gear (S4) of the four-planetary gear train (40) can be braked. The brake hub of the fourth brake (B4) welded to the inner circumference of the fourth sun gear (S4) is spline-connected to the second connecting member (8) extended to the outer circumference of the third planetary gear train (30). One friction member of the fourth brake (B4) is locked, and the other friction member is locked to a spline formed in front of the transmission case 1c. At the rear end of the transmission case 1b, a brake case holding a piston forming a hydraulic servo of the fourth brake (B4) and a return spring is fixed by a retaining ring.

  A second brake (B2) is arranged in front of the fourth brake (B4) to enable braking of the fourth planet carrier (P4) of the fourth planetary gear train (40). The right side member of the fourth planetary carrier (P4) that constitutes the fourth planetary gear train (40) is extended to the outer periphery of the fourth ring gear (R4), and the second brake (B2) is attached to the spline formed on the outer periphery. One of the friction members is locked. The other friction member of the second brake (B2) is locked to a spline formed on the inner circumference of the transmission case 1b. The piston and the return spring that form the hydraulic servo of the second brake (B2) are held in opposition to each other in the brake case that holds the piston and the return spring that form the hydraulic servo of the fourth brake (B4).

  A second planetary gear train (20) is arranged in front of the first connecting member (7) and the first and second clutches (C1, C2), and a second sun gear (S2) is provided on the outer periphery of the input shaft 3. Spline-connected to the supported first connecting member (7). The planetary pinion gear that meshes with the second sun gear (S2) is supported by the second planetary carrier (P2), and the left side member is the third clutch (on the inner peripheral side of the inverted L-shaped partition wall integrated with the transmission case 1b). C3) is spline-connected to the inner cylindrical member of the clutch cover. The second ring gear (R2) that meshes with the planet pinion gear has a plate that is axially fixed to the right end tooth portion with a retaining ring and is spline-connected to the second sun gear (S2). The axial direction is restricted by the thrust needle bearing between the right side member of P2) and the first connecting member (7), and the member is rotatably arranged.

  The third brake (B3) enables braking of the second ring gear (R2) of the second planetary gear train (20). The second ring gear (R2) has a spline formed on the outer peripheral portion thereof to lock one friction member of the third brake (B3). A spline is formed on the inner circumference of the transmission case 1b to lock the other friction member of the third brake (B3). Further, in front of the friction member of the third brake (B3), a piston and a return spring which are integrated with the brake case of the second brake (B2) and form a hydraulic servo of the third brake (B3) are integrated with the transmission case 1b. It is held by the partition wall.

  The output counter gear 5 arranged in front of the partition wall integral with the transmission case 1b is pivotally supported by a back-aligned angular bearing 4e fixed to the outer periphery of the cylindrical portion of the partition wall by screws. A spline hub integral with the clutch cover of the third clutch (C3) is connected to the left outer periphery of the output counter gear 5 and fixed by a retaining ring.

  The third clutch (C3) arranged in front of the output counter gear 5 holds the piston and the return spring to form the hydraulic servo of the third clutch (C3). The left side member extending through the inner side of the second planetary carrier (P2) of the second planetary gear train (20) is spline-coupled, and the bush is attached to the outer periphery of the first coupling member (7). The hydraulic oil of the third clutch (C3) is supplied from the oil passage that is supported and is opened in the inner circumference of the partition wall of the transmission case 1b through the oil passage that is sealed by the seal ring. Further, the outer peripheral plate of the clutch cover of the third clutch (C3) is extended to the outer peripheral side of the first planetary gear train (10) arranged in front of the third clutch (C3), and the third spline of the inner periphery is used as the third member. One friction member of the clutch (C3) is locked, and the other friction is applied to the spline formed on the outer periphery of the first ring gear (R1) of the first planetary gear train (10) of the third clutch (C3). The member is locked. The piston of the third clutch (C3) is arranged on the outer peripheral portion of the clutch cover and passes between a partition plate and a canceller plate for partitioning the hydraulic oil of the clutch extended to the outer peripheral portion and the canceller oil for canceling the centrifugal oil pressure of the piston. Oil is supplied.

  In the first planetary gear train (10) arranged in front of the third clutch (C3), the left side member of the first planetary carrier (P1) is extended to the inner peripheral side of the first sun gear (S1) to form an input shaft. The first sun gear (S1) is spline-coupled to the first coupling member (7) axially supported by the bushing on the outer circumference of the input shaft 3 on the inner circumferential side of the third clutch (C3). The first ring gear (R1) is welded to the clutch hub of the third clutch (C3) on the right side, and the clutch hub is pivotally supported by the bush on the outer periphery of the first sun gear (S1).

  In the structure of 11AT of FIG. 21, compared with the structure of 9AT of FIG. 20, since power is input from the third planetary gear train (30) to the fourth sun gear (S4) of the fourth planetary gear train (40), the fourth ring gear is used. The tooth flank load becomes larger than that of the structure of 9AT of FIG. 20 input to (R4), and the tooth width must be increased. In addition, since the power is output from the fourth ring gear (R4) by the first connecting member (7), the plate of the first connecting member (7) has the structure of 9AT in FIG. 20 which is output from the fourth sun gear (S4). The thickness becomes longer in the axial direction. However, since the tooth ratio between the ring gear and the sun gear of the first planetary gear train (10) can be reduced, the friction member of the third clutch (C3) can be arranged on the outer peripheral side of the first planetary gear train (10). As a result, the third clutch (C3) becomes compact. As a result, the axial directions of the structures of 11AT in FIG. 21 and 9AT in FIG. 20 are the same. The content of the patent claim is the same as that of 9AT in FIG. 20, but the structure of claim 4 in which the hydraulic chamber of the fourth brake (B4) is provided in the partition wall is added.

<1-TYPE> “C3-1-3, 15AT” FIGS. 3 and 15
FIG. 3 is a schematic diagram of two types of C3 type 15AT, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift speed, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 3, the velocity diagram is divided into a MAIN GEAR (main transmission mechanism) and a FRONT GEAR (front transmission mechanism), and the FRONT GEAR (front transmission mechanism) is a main front transmission mechanism and a sub front transmission mechanism. It is divided into a stationary transmission mechanism. The speed diagram of the MAIN GEAR (main transmission mechanism) is such that the first, second, third, and fourth constituent elements are arranged in order from the right side of the figure, and the speed diagram of the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). 5, the fifth, sixth and seventh constituent elements (A, B, C) are arranged in order from the right of the figure, and the speed diagram of the main front transmission mechanism is the eighth, ninth, tenth and eleventh in order from the right of the figure. The constituent elements (D, E, F, G) are arranged, the sixth constituent element (B) and the eighth constituent element (D) are connected by the second connecting member (8), and the first constituent element and the tenth constituent element. The element (F) is connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed. This arrangement is exactly the same as in FIGS. 1 and 2 except that the eleventh component (G) is added.

  In the two types of schematic diagrams in Fig. 3, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. The transmission has a third planetary gear train (30) consisting of simple planetary gears in the axial direction from the left front in the left figure in the FR specification, a two-story fourth planetary gear train (40) and a fifth planetary gear train (50). ), A second planetary gear train (20) and a first planetary gear train (10) are arranged, and the first and second planetary gear trains (10, 20) constitute a MAIN GEAR (main transmission mechanism), and a third planetary gear train (10). , The 4th and 5th planetary gear trains (30, 40, 50) form a FRONT GEAR (front transmission). The layout of each planetary gear is reversed in the right figure of the FF specification. In addition, the third planetary gear train (30) constitutes a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), and the fourth and fifth planetary gear trains (40, 50), which are two stories, are mainly in front. It constitutes a stationary transmission mechanism. The first, second, third, fourth and fifth planetary gear trains (10, 20, 30, 40, 50) include first, second, third, fourth and fifth sun gears (S1, S2, S3, S4, S5), first, second, third, fourth, fifth planetary carriers (P1, P2, P3, P4, P5), first, second, third, fourth, fourth 5 ring gears (R1, R2, R3, R4, R5). The speed change mode is described in Patent Document 2 and will not be described. Compared to the C3 type 11AT of FIG. 2, FIG. 3 only adds the fifth planetary gear train (50) to FIG. Note that the FF specification does not require a torque converter having a torque amplification function at the 15AT, so a fluid coupling whose axial direction is shorter than that of the torque converter by about 20% was used.

  The schematic diagram of FIG. 3 is almost the same as the schematic diagram described in FIG. The difference is only the connection of the third planetary gear train (30) that constitutes the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). In Patent Document 2, the input shaft is connected to the third ring gear (R3) of the third planetary gear train (30) so that the third sun gear (S3) can be braked by the first brake (B1). The input shaft is connected to the sun gear (S3) so that the third ring gear (R3) can be braked by the first brake (B1), and the decelerated rotation transmitted to the eighth component (D) is disclosed in Patent Document 2. I made it bigger. It should be noted that the gear ratios of the ring gear and the sun gear of the first, second, third, fourth, and fifth planetary gear trains (10, 20, 30, 40, 50) are exactly the same as in Patent Document 2, In the present application, the connection of the third planetary gear train (30) is changed so that the Gear Range is slightly smaller than in Patent Document 2. Further, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is 2.000, and the gear ratio of the ring gear and the sun gear of the second planetary gear train (20). Is 3.000, and the tooth ratio of the ring gear and the sun gear of the first planetary gear train (10) is smaller than that of 11AT in FIG. 2, and the third clutch (on the outer periphery of the first ring gear (R1) is the same as in FIG. The friction member of C3) can be arranged. In the schematic diagrams of the FR specifications and the FF specifications on the left and right of FIG. 3, the third planetary gear train (30) forming the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism) and the main front transmission mechanism 2 are formed. The first and second clutches (C1, C2) having a two-story structure are arranged between the shafts of the fourth and fifth planetary gear trains (40, 50) having a one-story structure, and the front transmission mechanism of the transmission case is arranged. The third planetary gear train (30) and the first and second clutches (C1, C2) are arranged on the outer side in the outer circumferential direction of the cylindrical member provided at one end that is the end of the cylindrical member end portion that axially supports the input shaft. The structure of "claim 1" is shown in which the second and fourth planetary gear trains (40, 50) are arranged on the transmission mechanism side. In the schematic diagram of the FF specification on the right side of FIG. 3, the first planetary gear train (10) and the second planetary gear train (20) of the MAIN GEAR (main transmission mechanism) are arranged with the output counter gear interposed therebetween. The structure is shown.

In the table showing the gear ratios of “C3-1-3, 15AT” in FIG. 3, the gear ratios are 11.199 to 0,604, 10 forward deceleration stages and 5 speed increasing stages. No gear ratio is obtained. The range of the gear ratio (RANGE) is 18.54, which is so large that there is no comparison target for passenger cars, but the step value (STEP) of the gear ratio becomes a little cross from the low speed region to the medium speed region. The gearing efficiency (GEAR EFF) of the planetary gears is 96.2% when the gear ratio is 11.199 at 1st and 97.4% when it is 7.064 at 2nd, but exceeds 98% at other gears. In particular, since it exceeds 99% in the speed increasing stage, the efficiency is close to the highest level as a whole. Generally, since a multi-speed transmission such as 15AT starts at the third speed, the inefficiency of 1st and 2nd does not pose a problem. Needless to say, 15AT as shown in the schematic diagram of the FF specification of FIG. 3 is not required for a passenger car, but only the fifth planetary gear train (50) is added to the C3 type 11AT of FIG. 2, and the fifth planetary gear train ( Since 50) is arranged on the first floor of the fourth planetary gear train (40), the gear width is slightly widened, and is substantially the same as the axial length of 11AT.
Therefore, it can be imagined that the total length will be around 410 mm even if the input power from the prime mover is set to 300 Nm and the torque converter is used. If the fluid joint is used instead of the torque converter, it will be about 400 mm, which is the same as C3 type 9AT, 11AT. FIG. 3 is for determining the limit of the number of gear stages that can be installed in a passenger car, because the gear ratio and the meshing efficiency of the planetary gears are appropriate, and it can be said that it is possible up to the 15th forward speed.

  FIG. 15 “C1-1-3, 15AT (FR)” is a structural diagram conceptually designing the schematic diagram on the right side of the FR specification of FIG. 3 with the input torque from the prime mover set to 300 Nm. It is almost the same as 11AT (FR), only about 1.5% longer. In FIG. 15, a prime mover (not shown) is arranged on the left front side of the transmission, and power is input to the transmission via the torque converter (200a). The transmission case 1 is integrally arranged, and the transmission case 1 is arranged in the left front from the third planetary gear train (30), the first and second clutches (C1 and C2), and the fourth floor of a two-story structure. , A fifth planetary gear train (40, 50) and front transmissions composed of first, second, and fourth brakes (B1, B2, B4) arranged on the outer periphery thereof, and a second planetary gear train. (20), a first planetary gear train (10), a third clutch (C3) in which friction members are arranged on the outer periphery of the first planetary gear train (10), and a friction member on the outer periphery of the second planetary gear train (20) A main speed change mechanism including a third brake (B3) in which is arranged is sequentially arranged.

  A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened to the front portion of the transmission case 1 with bolts, and a gear shift for fixing the wheel stator of the torque converter (200a) to the holding member 2a. The cylindrical member 2b, which is cylindrically extended in the machine inward direction, is fastened with a bolt. Bushings 4a and needle roller roller bearings 4b are arranged on the inner circumferences of both ends of the cylindrical member 2b to support the input shaft 3a. An output shaft 3c rotatably supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e is arranged at the rear portion of the transmission case 1, and an input shaft 3a is arranged by a needle roller roller bearing 4d arranged on the inner circumference of the output shaft 3c. To support. Here, since the integrated input shaft 3a is pivotally supported at three points, the distance between the respective pivots is short, and the input shaft 3a is less susceptible to vibration due to unbalance in the circumferential direction, so that the torque transmission capacity is reduced. The diameter of the portion arranged around the input shaft 3a can be reduced, which leads to the weight reduction of the transmission.

  On the outer periphery of the cylindrical member 2b, a third planetary gear train (30) and a first and second clutches (C1, C2), which serve as a sub front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged from the left front. . The third planetary gear train (30) decelerates the rotation of the input shaft 3a and selectively transmits the rotation to the fourth and fifth planetary gear trains (40, 50) serving as the main front transmission mechanism. The third ring gear (R3) of the third planetary gear train (30) is extended to the outer periphery of the third planetary gear train (30) to lock the friction member of the first brake (B1) and to the inner periphery. A thin plate-shaped brake hub having a bush 4k press-fitted into the inner diameter of a tubular member is welded and rotatably held on the outer periphery of the cylindrical member 2b. The left side member of the third planetary carrier (P3) is cylindrically extended on the inner circumference, and is freely rotatable around the outer circumference of the brake hub welded to the third ring gear (R3) by the bush 4j press-fitted on the inner diameter. Further, the third sun gear (S3) is rotatably held by the bush 4c press-fitted on the inner periphery of the third planet carrier (P3) on the outer periphery which is cylindrically extended on the inner periphery of the left side member. It The right side member of the third planet carrier (P3) extends to the outer periphery of the third planetary gear train (30) and extends rearward along the inner side of the transmission case 1 as the second connecting member (8). The third sun gear (S3) is welded to the clutch covers of the first and second clutches (C1, C2) welded to the input coupling member (Y). The input coupling member (Y) is spline-coupled to the input shaft 3a behind the cylindrical member 2b, and extends forward in a cylindrical shape along the outer periphery of the cylindrical member 2b to form the first and second clutches (C1, C2). The third sun gear (S3) is welded to the third sun gear (S3) via the clutch cover, and the third sun gear (S3) is rotatably held by a bush 4c press-fitted on the inner circumference.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member 2b behind the third planetary gear train (30). The first clutch (C1) can connect the input shaft 3a to the fifth planet carrier (P5) of the fifth planetary gear train (50), and the second clutch (C2) connects the input shaft 3a and the fifth sun gear (S5). Can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in the radial direction is integrated with the input connecting member (Y). A servo mechanism for pressing the friction member of the two-clutch (C2) is arranged. A servo mechanism for pressing the friction member of the first clutch (C1) is arranged on the clutch covers of the first and second clutches (C1, C2) welded to the input connecting member (Y). Between the clutch covers of the first and second clutches (C1, C2) and the double clutch coupling member (X), a partition plate supplies the servo mechanism of the first clutch (C1) arranged on the upper side in the radial direction. A passage for the hydraulic oil of the piston and a hydraulic cancel oil is formed. The servo mechanism of the second clutch (C2) and the servo mechanism of the first clutch (C1) arranged on the dual clutch connecting member (X) are provided on the input connecting member (Y) from the outer circumference of the cylindrical member 2b. Hydraulic oil and hydraulic cancel oil are supplied through an oil passage sealed by a piston ring. The structure of the first and second clutches (C1, C2) is the "claim 2" of the present application, which is extremely simple and compact, and the oil passages to the first and second clutches (C1, C2) are also line resistance. Is small and the responsiveness of the clutch is improved.

  Behind the cylindrical member 2b and the first and second clutches (C1, C2), there are two-story fourth and fifth planetary gear trains (40) which are the main front transmission mechanism of the FRONT GEAR (front transmission mechanism). , 50) are arranged. The fourth and fifth planetary gear trains (40, 50) selectively transmit eight types of rotations to the first and second planetary gear trains (10, 20) which are main transmission mechanisms. The fifth ring gear (R5) of the fifth planetary gear train (50) and the fourth sun gear (S4) of the fourth planetary gear train (40) integrated with each other include the right side member of the third planetary carrier (P3). The second connecting member (8) extended along the inner side of the transmission case 1 is spline-connected to the fourth sun gear (S4). A fifth planetary gear train (50) is arranged on the first floor of the fourth and fifth planetary gear trains (40, 50), which are two-story, and one friction member of the second clutch (C2) is provided on the outer peripheral spline. A fifth sun gear (S5) to which a clutch hub to be locked is welded is rotatably arranged on the outer periphery of the input shaft 3a by a bush 4i. The other friction member of the second clutch (C2) is locked to the inner peripheral spline of the double clutch connecting member (X). The fifth planetary carrier (P5) of the fifth planetary gear train (50) has a left side member that is a bush 41 and is axially supported on the outer periphery of the fifth sun gear (S5), and the fifth planetary carrier (P5) of the fourth planetary gear train (40) The right side member of the fifth planet carrier (P5) of the fifth planetary gear train (50), which is integral with the right side member of the carrier (P4), is rotatably arranged around the input shaft 3a by the bush 4m. Here, the fourth and fifth planetary carriers (P4, P5) of the two-story fourth and fifth planetary gear trains (40, 50) are firmly supported on both sides. Further, a clutch hub in which one friction member of the first clutch (C1) is locked to the inner peripheral spline is spline-connected to the left side member of the fifth planetary carrier (P5) of the fifth planetary gear train (50). The other friction member of the first clutch (C1) is locked to the outer peripheral spline of the dual clutch connecting member (X). A brake hub of a second brake (B2) extended to the outer periphery of the fourth planetary gear train (40) is welded to the left side member of the fourth planetary gear train (40) on the fourth planetary carrier (P4). A first connecting member (7) is spline-connected to the rear inner periphery of the output fourth ring gear (R4). The arrangement and structure of the first, second and fourth brakes (B1, B2, B4) are the same as 11AT in FIG.

  The second planetary gear train (20) arranged behind the fourth planetary gear train (40) has the first connection in the first forward speed (1st) to the seventh forward speed (7th) and the reverse (Rev1). The rotation input from the member (7) and the input shaft 3a via the first planetary gear train (10) is decelerated and output. The second sun gear (S2) of the second planetary gear train (20) is integrally molded with the first sun gear (S1) of the first planetary gear train (10) and is rotatable around the outer periphery of the input shaft 3a by the needle roller roller bearing 4h. And is spline-connected to the first connecting member (7) at the front. The planetary pinion gear that meshes with the second sun gear (S2) is supported by the second planetary carrier (P2), the right side member is welded to the output shaft 3c, and the output hub is arranged on the outer periphery of the first planetary gear train (10). 9 is splined. Further, the second ring gear (R2) that meshes with the planetary pinion gear has a plate in which the left end tooth portion is axially fixed by a retaining ring and is spline-connected to the first connecting member (7) to extend the inner periphery of the plate. The two planetary gear trains (20) and the left side member of the second planetary carrier (P2) are rotatably arranged with their axial directions restricted by thrust needle bearings. The friction member of the third brake (B3) is locked on the outer periphery of the second ring gear (R2), but the center of gravity of both is within the width of the planet pinion gear of the fourth planetary gear train (40). Since the radial load of the second ring gear (R2) and the friction member is received by the two planet carrier (P2), the radial bearing dedicated to the second ring gear (R2) is not required.

  The first planetary gear train (10) arranged behind the second planetary gear train (20) has the first connecting member (7) and the input shaft 3a in the fourth forward speed (7th) to the ninth forward speed (15th). The rotation input from is input to the first sun gear (S1) and the first planet carrier (P1), and output from the first ring gear (R1). The planetary pinion gear meshing with the first sun gear (S1) integrally formed with the second sun gear (S2) is supported by the first planetary carrier (P1), and the right side member is splined to the input shaft 3a. The first ring gear (R1) that meshes with the planetary pinion gear is arranged behind the first planetary gear train (10), is axially supported by the bush 4g on the inner circumference, and is a spline formed on the outer circumference and is the third clutch (C3). Lock the friction member of.

  The third clutch (C3) can connect the first ring gear (R1) of the first planetary gear train (10) and the output shaft 3c. One of the friction members of the third clutch (C3) is locked to the spline formed on the outer periphery of the first planetary gear train (10), and the spline formed on the inner periphery of the output hub 9 welded to the output shaft 3c. The other friction member of the third clutch (C3) is locked at. The piston of the third clutch (C3) and the return spring are held in front of the output shaft 3c, and a hydraulic servo of the third clutch (C3) is formed. The output shaft 3c is supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e at the rear of the transmission case 1, and is sealed from the transmission case 1 by a seal ring between the needle roller roller bearing 4f and the deep groove ball bearing 4e. The hydraulic oil of the third clutch (C3) is supplied through the oil passage. A parking gear 6a is formed on the outer circumference of the flange of the output shaft 3c. The arrangement and structure of the third brake (B3) are the same as in FIG.

  FIG. 4 of Patent Document 2 is a structural diagram conceptually designed for a commercial vehicle with an input torque from a prime mover of 1000 Nm. The commercial vehicle is ten times as harsh as a load frequency compared to a passenger vehicle, and has a durability life of 10 times, and is completely strong. It ’s different. Therefore, in FIG. 4 of Patent Document 2, the gear widths of the fourth and fifth planetary gear trains (40, 50) and the second planetary gear train (20), which are two-story, are made extremely large. So you don't have to make it that big. However, since the reduction ratio is larger than 9AT and 11AT in FIG. 14, the gear width is increased. As a result, the axial length is increased, but the diameter of the first planetary gear train (10) can be reduced, so that the friction member of the third clutch (C3) can be arranged on the outer periphery of the first planetary gear train (10) to be compact. Therefore, the total length is only 1.5% longer than 9AT and 11AT in FIG. In addition, although three types of 9AT, 11AT, and 15AT are described as examples of "1-TYPE" according to Patent Document 2, the A type 6AT of the present application is more efficient than the 15AT used in the FRONT GEAR (forward shift mechanism). However, the B type 6AT or C type 5AT described in “Paragraph 4” may be used for the FRONT GEAR (frontward transmission mechanism) that transmits one type of speed change rotation to the main frontward transmission mechanism. The A type 6AT used for the 15AT of the present application has a power train and a structure which is more compact in the axial direction than the currently commercialized 4AT and can be commercialized as 6AT.

<2-TYPE>
In the multi-stage transmission described in paragraph "0014", the power train of Patent Document 1 proposed by Daimler is modified to have the structure of the invention of the present application. The power train of Patent Document 1 is a C3 type 9AT and has been put to practical use by Daimler, and the arrangement of each part that has been put to practical use is different from that of Patent Document 1, and the applicant of the present application described it in FIG. 6 of Patent Document 2. . Since the layout proposed by Daimler and the layout put to practical use are only FR specifications, and are not simple and compact, the FR specifications and FF are shown in FIG. 4, FIG. 16, and FIG. 22 of “C3-2-1, 9AT” of the present application. Claim the structure of the specification as a patent and describe its example. Since the power train of Patent Document 1 is only 9AT, 11AT based on this 9AT is described as "C3-2-2, 11AT" in FIG. 5 as an example of the present application.

<2-TYPE> “C3-2-1, 9AT” FIGS. 4, 16 and 22
FIG. 4 shows a schematic diagram of two types of C3 type 9AT, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift speed, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 4, the velocity diagram is divided into MAIN GEAR (main transmission mechanism) and FRONT GEAR (previous transmission mechanism), and FRONT GEAR (previous transmission mechanism) is compared with “1-TYPE”. , In order to simplify understanding, it is divided into a main front transmission mechanism and a sub front transmission mechanism. The speed diagram of the MAIN GEAR (main transmission mechanism) is such that the first, second, third, and fourth constituent elements are arranged in order from the right side of the figure, and the speed diagram of the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). The fifth, sixth, and seventh constituent elements (A, B, C) are arranged in order from the right in the figure, and the speed diagram of the main front transmission mechanism is the sixth, seventh, and eighth constituent elements in order from the right in the figure. (B, C, D) are arranged, the first component and the sixth component (B) are connected by the first connecting member (7), and the third component serves as the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  Since the gearshift mode of the MAIN GEAR (main gearshift mechanism) is the same as that in FIG. 1, the description thereof is omitted, and the gearshift mode of the FRONT GEAR (front gear shift mechanism) will be described. In the velocity diagram of FIG. 4, the third planetary gear train (30) serving as the auxiliary front speed change mechanism has the third sun gear (S3) serving as the fifth constituent element (A) connected to the input shaft, and the sixth constituent element. The third planet carrier (P3), which is (B), can be braked by the first brake (B1), can be connected to the input shaft by the first clutch (C1), and is the fourth main forward transmission mechanism. The fourth ring gear (R4), which is the output component of the planetary gear train (40), can be connected by the second clutch (C2), and the third ring gear (R3), which is the seventh component (C), is the fourth component. The fourth sun gear (S4) is connected to the fourth planet carrier (P4) of the planetary gear train (40) and can be braked by the second brake (B2). Here, the third planetary gear train (30) serving as the auxiliary front speed change mechanism outputs the reverse rotation of the input shaft to the fourth planet carrier (P4) by engaging the first brake (B1) and the second clutch ( C2) outputs 0 rotation to the fourth ring gear (R4), and first clutch (C1) outputs the rotation of the input shaft to the fourth planet carrier (P4) and the second clutch (C2). The rotation of the input shaft is output to the fourth ring gear (R4) by fastening. That is, the third planetary gear train (30) serving as the auxiliary front speed change mechanism provides the fourth planetary gear train (40) serving as the main front speed change mechanism with three types of rotations of the input shaft, the reverse rotation, and the zero rotation. It is output. Further, the fourth planetary gear train (40), which serves as the main front shifting mechanism, brakes the fourth sun gear (S4) with the second brake (B2) so that the third planetary gear (R3) moves to the fourth planetary carrier (40). The rotation and the reverse rotation of the input shaft output in P4) are accelerated and output. That is, the fourth planetary gear train (40) serving as the main front transmission mechanism outputs four kinds of rotations of the input shaft, namely, rotation of the input shaft, reverse rotation, zero rotation, and accelerated rotation. Further, the speed diagram of FIG. 4 shows the third planetary gear train (30) which becomes the auxiliary front speed change mechanism and the fourth planetary gear train (40) which becomes the main front speed change mechanism by engaging the second clutch (C2). There are four constituent elements (A, B, C, D), and decelerated rotation of the input shaft is obtained by braking with the second brake (B2), and a total of five types of rotation are output.

  In the two types of schematic diagrams in Fig. 4, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. A prime mover is located on the front left side (not shown), and power is input to the input shaft of the transmission via the torque converter. In the left diagram of the FR specification, the third planetary gear train (30), the fourth planetary gear train (40), the second planetary gear train (20), the first planetary gear train (20), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40) In the right figure of the FF specification in which the planetary gear train (10) is arranged, the first planetary gear train (10), the second planetary gear train (20), the second planetary gear train (20), the second planetary gear train (20) A four planetary gear train (40) and a third planetary gear train (30) are arranged, and the first and second planetary gear trains (10, 20) constitute a MAIN GEAR (main transmission mechanism), and third and fourth The planetary gear train (30, 40) constitutes a FRONT GEAR (forward gear change mechanism). Further, the third planetary gear train (30) constitutes a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), and the fourth planetary gear train (40) constitutes a main front transmission mechanism. In the right diagram of the FF specification, an output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20) and is axially supported by a partition wall integrated with the transmission case. The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The input shaft is rotatably supported at the other end of the transmission case, and in the FR specification, the input shaft is rotatably supported at the output shaft rotatably supported at the other end of the transmission case. C1 and C2) are arranged between the third planetary gear train (30) and the fourth planetary gear train (40) on the outer peripheral side of the cylindrical member so that the friction members are superposed in two stages in the radial direction. A fourth planetary gear train (S4, P4, R4) of the front transmission mechanism (FRONT GEAR) is arranged in the axial direction from the inner peripheral end portion toward the main transmission mechanism (MAIN GEAR).

  The table of FIG. 4 shows the engagement elements (SHIFT) and the gear ratios (RATIO) and the meshing efficiencies (GEAR EFF) of the planetary gears at each gear stage that Daimler has commercialized. Regarding the gear ratio (RATIO), the gear ratio step value (STEP), and the gear ratio range (RANGE), which show the performance, the gear ratio range (RANGE) becomes a little smaller value for 9.08 and 9AT. The step value (STEP) of is almost appropriate. The meshing efficiency (GEAR EFF) of the planetary gears is higher than 99%, which is almost the same as that of FIG. 1, except that the first forward speed (1st) is 98.1% and the third forward speed (3rd) is 98.6%. The performance is considerably superior to the conventional 8AT.

  FIG. 16 “C3-2-1, 9AT (FR)” is a structural diagram conceptually designed for passenger cars with the FR power schematic diagram on the left side of FIG. 4 as the input power from the prime mover is 300 Nm. In FIG. 16, a prime mover (not shown) is arranged on the left front side of the transmission, and power is input to the transmission via the torque converter (200a). The transmission case 1 is arranged as one body, and the third planetary gear train (30), the first and second clutches (C1, C2), and the fourth planetary gear train (40 ) And first and second brakes (B1, B2) arranged on the outer circumferences thereof, a second planetary gear train (20), a first planetary gear train (10), a third A main transmission mechanism including a clutch (C3) and a third brake (B3) having a friction member arranged on the outer periphery of the second planetary gear train (20) is sequentially arranged.

  A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened to the front portion of the transmission case 1 with bolts, and a gear shift for fixing the wheel stator of the torque converter (200a) to the holding member 2a. The cylindrical member 2b, which is cylindrically extended in the machine inward direction, is fastened with a bolt. Bushings 4a and needle roller roller bearings 4b are arranged on the inner circumferences of both ends of the cylindrical member 2b to support the input shaft 3a. An output shaft 3c rotatably supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e is arranged at the rear portion of the transmission case 1, and an input shaft 3a is arranged by a needle roller roller bearing 4d arranged on the inner circumference of the output shaft 3c. To support. Here, since the integrated input shaft 3a is pivotally supported at three points, the distance between the respective pivots is short, and the input shaft 3a is less susceptible to vibration due to unbalance in the circumferential direction, so that the torque transmission capacity is reduced. The diameter of the portion arranged around the input shaft 3a can be reduced, which leads to the weight reduction of the transmission.

  On the outer periphery of the cylindrical member 2b, a third planetary gear train (30) and a first and second clutches (C1, C2), which serve as a sub front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged from the left front. . The third sun gear (S3) of the third planetary gear train (30) is rotatably held on the outer periphery of the cylindrical member 2b by a bush 4c press-fitted on the inner periphery thereof, and is splined to the input shaft 3a behind the cylindrical member 2b. The input connecting member (Y), which is a cylindrical member extending forward along the outer periphery of the cylindrical member 2b, is welded. The right side member of the third planetary carrier (P3) of the third planetary gear train (30) is welded to the double clutch connecting member (X) at the inner circumference, and the thin plate-shaped brake hub is welded to the left side member. The material is extended to the outer periphery of the third planetary gear train (30) to lock the friction member of the first brake (B1). The third ring gear (R3) has an outer periphery extending backward along the inner side of the transmission case 1 and is spline-connected with a connecting member welded to the fourth planet carrier (P4) of the fourth planetary gear train (40). To be done.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member 2b behind the third planetary gear train (30). The first clutch (C1) can connect the input shaft 3a and the third planet carrier (P3) of the third planetary gear train (30), and the second clutch (C2) connects the third planet carrier (P3) and the fourth planet carrier (P3). It can be connected to the ring gear (R4). In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The right side member of the third planetary carrier (P3) constituting the third planetary gear train (30) is welded to the front end of the third planetary gear train (30). On the inner side of the rear side of the dual clutch connecting member (X), the first clutch (locked to the radial inner circumference of the dual clutch connecting member (X) and the clutch hub splined to the input connecting member (Y) ( A servo mechanism for pressing the friction member of C1) is arranged, and the right side member of the third planetary carrier (P3) is arranged on the outer circumference in the radial direction of the dual clutch coupling member (X) and the fourth ring gear (R4). A servo mechanism for pressing the friction member of the second clutch (C2) is provided. Between the dual clutch coupling member (X) and the right side member of the third planet carrier (P3), a partition plate is provided to the servo mechanism arranged on the upper side in the radial direction of the right side member of the third planet carrier (P3). A passage for the supplied hydraulic oil for the piston and the hydraulic cancel oil is formed. The servo mechanism of the second clutch (C2) arranged on the upper side in the radial direction of the dual clutch coupling member (X) and the servo mechanism of the first clutch (C1) arranged on the lower side in the radial direction include the outer circumference of the cylindrical member 2b. Is supplied with hydraulic oil and hydraulic cancel oil through an oil passage sealed by a piston ring provided in the input connecting member (Y). Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism of the second clutch (C2) of the right side member of the third planetary carrier (P3), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 7" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

  A fourth planetary gear train (40) serving as a main front transmission mechanism of the FRONT GEAR (front transmission mechanism) is arranged behind the cylindrical member 2b and the first and second clutches (C1, C2). The fourth planetary gear train (40) selectively transmits five types of rotations to the first and second planetary gear trains (10, 20) which are main transmission mechanisms. The right side member of the fourth planetary carrier (P4) of the fourth planetary gear train (40) passes through the outer periphery of the first and second clutches (C1, C2) along the inner periphery of the transmission case 1 at the outer periphery. A connecting member that is spline-connected to the third ring gear (R3) of the three planetary gear train (30) is welded. The fourth ring gear (R4) of the fourth planetary gear train (40), which is the output, is extended to the outer circumference of the front first and second clutches (C1, C2) and serves as a friction member for the second clutch (C2). The second planetary gear train (20) and the first planetary gear train (10) that are locked and are spline-coupled to the front side of the first connecting member (7) and extended to the inner circumference of the fourth planetary gear train (40). ) Is integrated with the second sun gear (S2) and the first sun gear (S1), the first connecting member (7) is spline-connected. The left side member of the fourth planetary carrier (P4) of the fourth planetary gear train (40) is extended to the inner circumference of the fourth planetary gear train (40) and is axially supported by the first connecting member (7) by the bush 4i. , The fourth sun gear (S4) of the fourth planetary gear train (40) is pivotally supported by the second sun gear (S2) extended by the bush 4j on the inner circumference, and the fourth sun gear (S4) of the fourth planetary gear train (40) is arranged rearward. A brake hub that extends to the outer periphery and locks the friction member of the second brake (B2) is welded.

  The first brake (B1) enables braking of the third planet carrier (P3) of the third planetary gear train (30). A thin plate-shaped brake hub is welded to the left side member of the third planetary carrier (P3), is extended to the outer periphery of the third planetary gear train (30), and the first brake (B1) is attached to the spline formed on the outer periphery. One of the friction members is locked. The other friction member of the first brake (B1) is locked to a spline formed in front of the transmission case 1, and a piston is provided in a hydraulic chamber of a holding member 2a fixed to the front of the transmission case 1 by a bolt. The return spring is held and the hydraulic servo of the first brake (B1) is formed.

  The second brake (B2) enables braking of the fourth sun gear (S4) of the fourth planetary gear train (40). The brake hub welded to the rear portion of the fourth sun gear (S4) is extended to the outer circumference of the fourth ring gear (R4), and one friction member of the second brake (B2) is locked to the spline formed on the outer circumference. To be done. The other friction member of the second brake (B2) is locked to the spline formed at the center of the transmission case 1 at the outer circumference of the fourth ring gear (R4) to form a hydraulic servo for the second brake (B2). The brake case that holds the piston and the return spring is fixed to the transmission case 1 in front of the friction member with a retaining ring. The spline formed in the center of the transmission case 1 is extended backward to lock the friction member of the third brake (B3) on the outer periphery of the second planetary gear train (20).

  Since the structure of the main transmission mechanism (MAIN GEAR) arranged behind the fourth planetary gear train (40) is the same as that in FIG. 14 of "1-TYPE", description thereof will be omitted. Comparing the 9AT of "2-TYPE" that Daimler has put into practical use with the 9AT of FIG. 16 of the present application, the 9AT of FIG. 16 of the present application is about 15% more compact due to the arrangement structure of parts even with the same power train, Weight is reduced. Further, comparing the “2-TYPE” 9AT of FIG. 16 with the “1-TYPE” 9AT of FIG. 14, the third planetary gear train (30) and the fourth planetary gear train (40) of the front transmission (FRONT GEAR) are compared. ), The arrangement of each site is the same, only the connection is different. However, in the third planetary gear train (30) and the fourth planetary gear train (40) of “1-TYPE” 9AT of FIG. 14, since power is input to the ring gear having a large diameter, the tooth surface load is input to the sun gear. It is smaller than "2-TYPE" 9AT in FIG. 16, and the tooth width can be reduced accordingly. Therefore, the total length of the transmission can be shortened by about 1% in the "1-TYPE" 9AT in FIG. 14 than in the "2-TYPE" 9AT in FIG. 16, but it can be said that they are almost the same. Compared with the commercialized ZF "D-TYPE" 8AT and TOYOTA "B-TYPE" 8AT, the "1-TYPE" 9AT and "2-TYPE" 9AT have the same overall length of the transmission. Below, the gear ratio becomes extremely excellent. The meshing efficiency of the planetary gear is extremely poor for TOYOTA's "B-TYPE" 8AT, which is comparable to this class's best ZF "D-TYPE" 8AT.

  FIG. 22 “C3-2-1, 9AT (FF)” is a structural diagram conceptually designing the schematic diagram on the right of the FF specification suitable for passenger cars of FIG. 4 with the input power from the prime mover set to 300 Nm. The total length of "C3-1-1, 9AT (FF)" is slightly shorter than 400 mm, but slightly longer than 400 mm. The difference between FIG. 20 and FIG. 22 is only that the connection of the third planetary gear train (30) and the fourth planetary gear train (40) of the front transmission mechanism (FRONT GEAR) is different, and the arrangement of each part is the same and the big difference is made. Does not appear. Therefore, the structure of the front transmission mechanism (FRONT GEAR) of FIG. 22 will be described, and the others will be omitted. The structure diagrams of the "1-TYPE" 9AT of FIG. 20 and the "2-TYPE" 9AT of FIG. Although it does not reach, it is fully established.

  In FIG. 22, the transmission case 1c for closing the rear portion of the transmission has an integrally formed inner peripheral portion that protrudes forward in a cylindrical shape, and the needle roller roller bearing 4b is arranged at the inner peripheral end portion of the protruding cylindrical member. And supports the input shaft 3. Therefore, the shaft support of the input shaft 3 is shortened, and rotational vibration can be suppressed. A third planetary gear train (30) and a first and a second clutch (C1, C2), which serve as a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), are arranged on the outer periphery of the protruding cylindrical member from the right rear. It The third sun gear (S3) of the third planetary gear train (30) is rotatably held on the outer periphery of the cylindrical member of the transmission case 1c by a bush press-fitted on the inner periphery thereof, and is splined to the input shaft 3 in front of the cylindrical member. The input connecting member (Y) that is connected and extends rearward in a cylindrical shape along the outer periphery of the cylindrical member 2 is welded. The left side member of the third planetary carrier (P3) of the third planetary gear train (30) is welded to the double clutch connecting member (X) at the inner circumference, and the thin plate brake hub is welded to the right side member. The material is extended to the outer periphery of the third planetary gear train (30) to lock the friction member of the first brake (B1). The third ring gear (R3) has an outer periphery extending forward along the inner side of the transmission case 1 and is spline-coupled to the fourth planet carrier (P4) of the fourth planetary gear train (40). To be done.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member of the transmission case 1c in front of the third planetary gear train (30). The first clutch (C1) can connect the input shaft 3 and the third planet carrier (P3) of the third planetary gear train (30), and the second clutch (C2) connects the third planet carrier (P3) and the fourth planet carrier (P3). The ring gear (R4) can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The left side member of the third planetary carrier (P3) forming the third planetary gear train (30) is welded to the rear end of the third planetary gear train (30). The left side member of the third planet carrier (P3) is a servo that presses the radial outer circumference of the dual clutch coupling member (X) and the friction member of the second clutch (C2) arranged in the fourth ring gear (R4). A mechanism is provided, and a partition plate is provided between the dual clutch connecting member (X) and the left side member of the third planet carrier (P3), and is arranged above the left side member of the third planet carrier (P3) in the radial direction. It operates from the outer periphery of the cylindrical member of the transmission case 1c through the passage of the hydraulic oil of the piston and the hydraulic cancel oil supplied to the servo mechanism through the oil passage sealed by the piston ring provided in the input coupling member (Y). Oil and hydraulic cancel oil are supplied. Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism for the second clutch (C2) of the left side member of the third planetary carrier (P3), and seals the hydraulic chamber. A clutch hub that locks the friction member of the first clutch (C1) is spline-connected to the input coupling member (Y) inside the front side of the dual clutch coupling member (X), and a servo mechanism that presses the friction member. The clutch cover, the piston, and the return spring are held by the input connecting member (Y). The servo mechanism of the first clutch (C1) is supplied with working oil and hydraulic cancel oil from the outer circumference of the cylindrical member of the transmission case 1c through an oil passage sealed by a piston ring provided in the input coupling member (Y). It The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 7" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

  A fourth planetary gear train (40) serving as a main front transmission mechanism of the FRONT GEAR (front transmission mechanism) is arranged in front of the cylindrical member of the transmission case 1c and the first and second clutches (C1, C2). It The left side member of the fourth planetary carrier (P4) of the fourth planetary gear train (40) passes through the outer periphery of the first and second clutches (C1, C2) along the inner periphery of the transmission case 1c. A connecting member that is spline-connected to the third ring gear (R3) of the three planetary gear train (30) is welded. The fourth ring gear (R4) of the fourth planetary gear train (40), which is the output, is extended to the outer periphery of the rear first and second clutches (C1, C2) and serves as a friction member for the second clutch (C2). The output hub is splined at the rear side by being locked and extended to the inner circumference of the fourth planetary gear train (40), and splined with the first connecting member (7) axially supported by the bushing on the outer circumference of the input shaft 3. Be connected. The right side member of the fourth planetary gear train (P4) of the fourth planetary gear train (40) is extended to the inner circumference of the fourth planetary gear train (40) and is axially supported by the output hub by the bushing. The fourth sun gear (S4) of (40) is axially supported by the first connecting member (7) by the bush 4 of the inner circumference, and extends forward to the outer circumference of the fourth planetary gear train (40), and the second brake (S4). The brake hub that locks the friction member of B2) is welded.

  A thin plate brake hub is welded to the right side member of the third planetary carrier (P3), extended to the outer periphery of the third planetary gear train (30), and the first brake (B1) is attached to the spline formed on the outer periphery. One of the friction members is locked. The other friction member of the first brake (B1) is engaged with a spline formed in the transmission case 1c, and a piston and a return spring are held in a hydraulic chamber formed behind the transmission case 1c. A hydraulic servo of the brake (B1) is formed.

  The brake hub welded to the front part of the fourth sun gear (S4) is extended to the outer periphery of the fourth ring gear (R4), and one friction member of the second brake (B2) is engaged with the spline formed on the outer periphery. Be stopped. The other friction member of the second brake (B2) is locked to the spline of the brake case fixed to the rear part of the transmission case 1b by a retaining ring, and the hydraulic servo of the second brake (B2) is attached to the brake case. The forming piston and return spring are retained. The brake case holds a piston and a return spring that form a hydraulic servo for the third brake (B3) in opposition to each other.

<2-TYPE> “C3-2-2, 11AT” FIG. 5
FIG. 5 shows that the seventh component (C) of the FRONT GEAR (front transmission mechanism) of FIG. 4 showing the C3 type 9AT can be braked by the fourth brake (B4), and a larger decelerated rotation of the input shaft can be achieved in the MAIN GEAR. The main transmission mechanism can be transmitted. Therefore, six types of rotation of the input shaft, two types of decelerated rotation of the input shaft, increased rotation of the input shaft, zero rotation, and reverse rotation can be transmitted to the MAIN GEAR (main transmission mechanism). That is, as is clear from FIG. 5, regarding the arrangement of the constituent parts, the fourth brake (B4) is simply added to the schematic view of the “2-TYPE” 9AT in FIG. The seventh component (C) includes a fourth brake (P4) at a connecting portion between the third ring gear (R3) of the third planetary gear train (30) and the fourth planetary carrier (P4) of the fourth planetary gear train (40). Since B4) is arranged, the structure can be as simple and compact as "1-TYPE" 11AT in FIG. However, if a good gear ratio of 11AT is obtained, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is increased to 2.800, and the second planetary gear train (20). The gear ratio of the ring gear and sun gear in () must be reduced to 1.727. Therefore, the outer diameter of the first planetary gear train (10) is increased, and the friction member of the third clutch (C3) is attached to the outer periphery of the first planetary gear train (10) as shown by "1-TYPE" 11AT in FIG. Cannot be placed, and the axial direction becomes long. In addition, since the reduction ratio becomes larger, the capacities of the second clutch (C2) and the second planetary gear train (20) must be increased, and the shaft length becomes longer, which makes it unsuitable for FF specifications. The gear ratio is 7.317 to 0.602, and the deceleration stage is 7 and the speed increasing stage is 4, and the forward deceleration stage of "1-TYPE" 11AT in FIG. 2 is 8 and the speed increasing stage is 3. It is better, but there is no practical problem. The range of the gear ratio (RANGE) may be 12.24, which is wider than 10.93 of "1-TYPE" 11AT in FIG. Although the meshing efficiency of the planetary gears (GEAR EFF) is inferior to 9 AT of "1-TYPE" or "2-TYPE", it is the highest level as AT of this class like 11 AT of "1-TYPE" in FIG. As a comprehensive judgment, "1-TYPE" 11AT of FIG. 2 is more suitable for passenger cars, and it can be said that it is a power train suitable for commercial vehicles (Truck Bus) as FR specifications. The "2-TYPE" 11AT of FIG. 5 is a schematic diagram only and does not show a structural diagram, and the structure of the FR specification can be imagined more fully than the "2-TYPE" 9AT of FIG. 16 and becomes a compact structure.

<3-TYPE>
In the multi-stage transmission described in paragraph “0015”, the four ATs that are conventionally considered by two planetary gear trains are controlled by two clutches and three brakes, or by four planetary gear trains. It is 11AT which used 4AT which controlled four constituent elements with three clutches and two brakes for FRONT GEAR (front transmission mechanism). From FRONT GEAR (previous gear shift mechanism), there are 6 types of rotation of the input shaft, 2 types of deceleration rotation of the input shaft, 2 types of deceleration rotation of the input shaft, 0 rotation, and reverse rotation of the same as the "2-TYPE" 11AT of FIG. It becomes possible to transmit to MAIN GEAR (main transmission mechanism). As described in the paragraph “0015”, several patents have been filed as schematic diagrams, and the present application is a proposal to make them an optimal structure. Since there are planetary gear trains of this type that have not yet been applied for patents, "C3-3-1, 11AT" in Fig. 6, Fig. 17, "C3-3-2, 11AT" in Fig. 7, "C3-" of the present application. 4-1, 11AT ”FIG. 8 claims the structure of FR specifications and FF specifications as a patent, and an example thereof will be described. In addition, if the "fourth and fifth planetary gear trains (40, 50) having two floors" used in "1-TYPE" 15AT of FIG. 3 can be applied to the four constituent elements of the two planetary gear trains. The "3-TYPE" is compact, but the two-story system is not described in the embodiment because the gear ratio of the ring gear and the sun gear is limited and an appropriate gear ratio of 11AT cannot be obtained.

<3-TYPE> “C3-3-1, 11AT” FIGS. 6 and 17
"C3-3, 11AT" is a FRONT GEAR (forward gear change mechanism) that uses 4AT in which four constituent elements consisting of two planetary gear trains are controlled by two clutches and three brakes. FIG. 6 is a diagram showing two types of schematic diagrams, a velocity diagram showing a shift mode, a fastening element (SHIFT) at each shift stage, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. is there. In the velocity diagram of FIG. 6, the velocity diagram is divided into MAIN GEAR (main transmission mechanism) and FRONT GEAR (previous transmission mechanism), and the MAIN GEAR (main transmission mechanism) The first, second, third, and fourth constituent elements are arranged in order, and the speed diagram of the sub front transmission mechanism of FRONT GEAR (front transmission mechanism) is the fifth, sixth, seventh, and eighth constituent elements in order from the right of the figure. (A, B, C, D) are arranged, the first component and the seventh component (C) are connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  Since the gearshift mode of the MAIN GEAR (main gearshift mechanism) is the same as that in FIG. 1, the description thereof is omitted, and the gearshift mode of the FRONT GEAR (front gear shift mechanism) will be described. In the velocity diagram of FIG. 6, the third ring gear (R3) of the third planetary gear train (30) and the fourth sun gear (S4) of the fourth planetary gear train (40) are connected to each other to form the fifth component (A). ), The third planet carrier (P3) of the third planetary gear train (30) and the fourth planetary carrier (P4) of the fourth planetary gear train (40) are connected to form the sixth component (B). , The fourth ring gear (R4) of the fourth planetary gear train (40) constitutes the seventh component (C), and the third sun gear (S3) of the third planetary gear train (30) is the eighth component (D). ). The input shaft and the fifth and sixth constituent elements (A, B) can be connected by the first and second clutches (C1, C2) respectively, and the eighth, fifth and sixth constituent elements (D, A, B) Can be braked by the first, second, and fourth brakes (B1, B2, B4), respectively. When the first clutch (C1) and the first brake (B1) are engaged, the largest decelerated rotation of the input shaft is output, and when the second clutch (C2) and the first brake (B1) are engaged, the next input shaft is decelerated. Outputs rotation, outputs the rotation of the input shaft by engaging the first clutch (C1) and the second clutch (C2), and speeds up the input shaft by engaging the second clutch (C2) and the second brake (B2). Rotation is output, 0 rotation is output by engaging the first brake (B1) and the fourth brake (B4), and reverse rotation of the input shaft is output by engaging the first clutch (C1) and the fourth brake (B4). To do. In other words, 6 types of input shaft rotation, input shaft deceleration rotation 2 types, input shaft speedup rotation, 0 rotation, reverse rotation can be input to MAIN GEAR (main transmission mechanism) from FRONT GEAR (front transmission mechanism). become.

  In the two types of schematic diagrams in Fig. 6, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. Here, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is increased to 3.000, and the gear ratio of the ring gear and the sun gear of the second planetary gear train (20) is increased. A good gear ratio cannot be obtained unless the gear ratio is set to 2.222. Therefore, the outer diameter of the first planetary gear train (10) becomes large, and the friction member of the third clutch (C3) is provided on the outer periphery of the first planetary gear train (10) as shown in "1-TYPE" 11AT in FIG. It cannot be placed and the axial direction becomes long. In addition, since the reduction ratio becomes larger, the capacity of the second planetary gear train (20) must be increased, and the shaft length becomes longer, which makes it unsuitable for FF specifications. A fluid coupling that shortens the axial direction by about 20% was used. A prime mover is located on the left front (not shown), and power is input to the input shaft of the transmission via a torque converter (fluid coupling). In the left diagram of the FR specification, the third planetary gear train (30), the fourth planetary gear train (40), the second planetary gear train (20), the first planetary gear train (20), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40) In the right figure of the FF specification in which the planetary gear train (10) is arranged, the first planetary gear train (10), the second planetary gear train (20), the second planetary gear train (20), the second planetary gear train (20) A four planetary gear train (40) and a third planetary gear train (30) are arranged, and the first and second planetary gear trains (10, 20) constitute a MAIN GEAR (main transmission mechanism), and third and fourth The planetary gear train (30, 40) constitutes a FRONT GEAR (forward gear change mechanism). In the right diagram of the FF specification, an output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20) and is axially supported by a partition wall integrated with the transmission case. The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The input shaft is rotatably supported at the other end of the transmission case, and in the FR specification, the input shaft is rotatably supported at the output shaft rotatably supported at the other end of the transmission case. C1 and C2) are arranged so that the friction members are radially stacked in two stages, and the fourth planetary gear train (40) of the front gearbox (FRONT GEAR) and the main gearbox side (MAIN GEAR) are located outside the cylindrical member in the outer circumferential direction. Of the second planetary gear train (20). The three first, second, and fourth brakes (B1, B2, B4) arranged in the front transmission mechanism (FRONT GEAR) brake the components approaching in the axial direction, so that "1-TYPE" or It becomes difficult to arrange it on the outer periphery of the third and fourth planetary gear trains (30, 40) and the first and second clutches (C1, C2) like 11 AT of "2-TYPE", and the axial direction becomes long. .

  The table of FIG. 6 is a table in which the applicant selects the gear ratios of the ring gear and the sun gear of each planetary gear train so that an appropriate gear ratio of 11AT can be obtained. The fastening element (SHIFT) and the gear ratio at each gear (RATIO) and the meshing efficiency of the planetary gears (GEAR EFF) are shown. Regarding the gear ratio (RATIO) indicating the performance, the gear ratio step value (STEP), and the gear ratio width (RANGE), the gear ratio width (RANGE) is 11.27, and the gear ratio step value (STEP) is also ideal. It is not appropriate, but it is almost appropriate. The meshing efficiency (GEAR EFF) of the planetary gears is excellent except for the first forward speed (1st) 97.2% and the second forward speed (3rd) 97.7%, and there is no significant problem in terms of performance. However, like the "2-TYPE" 11AT, it is better that the deceleration stage is 7 and the speed-up stage is 4, and the forward deceleration stage of the "1-TYPE" 11AT is 8 and the speed-up stage is 3 stages. .

  FIG. 17 “C3-3-1, 11AT (FR)” is a structural diagram conceptually designing a schematic view of the FR specification shown on the left side of FIG. 6 for a passenger vehicle with an input power from a prime mover of 300 Nm. In FIG. 17, a prime mover (not shown) is arranged on the left front side of the transmission, and power is input to the transmission via the torque converter (200a). The transmission case 1 is arranged as one body, and the second brake (B2), the third planetary gear train (30), the fourth planetary gear train (40), the first, A front transmission including a second clutch (C1, C2) and first and fourth brakes (B1, B4) arranged on the outer periphery thereof, a second planetary gear train (20), and a first planetary gear. A main transmission mechanism including a train (10), a third clutch (C3), and a third brake (B3) having friction members arranged on the outer periphery of the second planetary gear train (20) is sequentially arranged.

  A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened to the front portion of the transmission case 1 with bolts, and a gear shift for fixing the wheel stator of the torque converter (200a) to the holding member 2a. The cylindrical member 2b, which is cylindrically extended in the machine inward direction, is fastened with a bolt. Bushings 4a and needle roller roller bearings 4b are arranged on the inner circumferences of both ends of the cylindrical member 2b to support the input shaft 3a. An output shaft 3c rotatably supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e is arranged at the rear portion of the transmission case 1, and an input shaft 3a is arranged by a needle roller roller bearing 4d arranged on the inner circumference of the output shaft 3c. To support. Here, since the integrated input shaft 3a is pivotally supported at three points, the distance between the respective pivots is short, and the input shaft 3a is less susceptible to vibration due to unbalance in the circumferential direction, so that the torque transmission capacity is reduced. The diameter of the portion arranged around the input shaft 3a can be reduced, which leads to the weight reduction of the transmission.

  On the outer circumference of the cylindrical member 2b, from the left front, the second brake (B2) of the FRONT GEAR (previous transmission mechanism), the third planetary gear train (30), the fourth planetary gear train (40), the first and second clutches. (C1, C2) are arranged. The third sun gear (S3) of the third planetary gear train (30) is welded to the brake hub of the first brake (B1) in front of the third sun gear (S3) and is extended to the outer periphery of the third planetary gear train (30) to form the first brake (S3). The friction member (B1) is locked, and the bush 4j is press-fitted on the inner circumference, and is axially supported on the outer circumference of the brake hub of the second brake (B2). The right side member of the third planetary carrier (P3) of the third planetary gear train (30) is extended to the inner circumference of the third planetary gear train (30), and is connected to the bush 4j of the brake hub of the second brake (B2). The left side member of the third planet carrier (P3) is pivotally supported on the outer periphery and extends to the rear of the outer periphery of the third planetary gear train (30) to lock the friction member of the fourth brake (B4) and the fourth member. It is splined to the fourth planet carrier (P4) of the planetary gear train (40). The third ring gear (R3) of the third planetary gear train (30) has a connecting member welded to the rear side, and is spline-coupled to the fourth sun gear (S4) of the fourth planetary gear train (40).

  A fourth planetary gear train (40) is arranged behind the third planetary gear train (30). The left side member of the fourth planetary carrier (P4) of the fourth fourth planetary gear train (40) has a third planetary carrier (P3) extended to the rear of the outer periphery of the third planetary gear train (30) on the outer periphery. The right side member is spline-connected to the left side member, and the right side member is axially supported on the outer circumference of the clutch hub of the first clutch (C1) spline-connected to the fourth sun gear (S4) by the bush 4k press-fitted on the inner circumference. A clutch hub that is extended to the rear of the outer periphery and that locks the friction member of the second clutch (C2) is welded. The fourth sun gear (S4) of the fourth planetary gear train (40) is spline-connected to the connecting member welded to the third ring gear (R3) on the outer circumference and is extended forward on the inner circumference. The brake hub of (B2) and the clutch hub of the first clutch (C1) extending rearward are spline-connected, and the brake hub of the second brake (B2) and the clutch hub of the first clutch (C1) are cylindrical members. A bush 4i and a bush 4c are pivotally supported on the outer circumference of 2b. The fourth ring gear (R4) of the fourth planetary gear train (40) is spline-connected to the first connecting member (7) that extends rearward on the outer circumference.

  First and second clutches (C1, C2) are arranged at the outer peripheral end of the cylindrical member 2b behind the fourth planetary gear train (40). In the first and second clutches (C1 and C2), a dual clutch connecting member (X) in which friction members are coaxially overlapped in two radial steps is formed integrally with the input connecting member (Y), The member (Y) is spline-connected to the input shaft 3a behind the cylindrical member 2b. A clutch cover is welded to the rear of the input connecting member (Y), and a servo mechanism for pressing the friction member of the second clutch (C2) locked to the clutch hub of the third planet carrier (P3) is arranged in front of the outer periphery. A servo mechanism that presses the friction member of the first clutch (C1) locked to the clutch hub that is spline-connected to the fourth sun gear (S4) is provided inside the front side of the dual clutch connecting member (X). It Between the dual clutch connecting member (X) and the clutch cover, a partition plate forms a passage for the working oil of the piston and the hydraulic cancel oil supplied to the servo mechanism of the two clutch (C2). In the servo mechanism of the first and second clutches (C1, C2), the working oil and the hydraulic cancel oil are provided from the outer circumference of the cylindrical member 2b through an oil passage sealed by a piston ring provided in the input coupling member (Y). Is supplied. The structure of the first and second clutches (C1, C2) is the "claim 2" of the present application, which is extremely simple and compact, and the oil passages to the first and second clutches (C1, C2) are also line resistance. Is small and the responsiveness of the clutch is improved.

    In front of the transmission case 1, a first brake (B1) and a second brake (B2) are radially arranged in two stages with the first brake (B1) as an upper part. The first brake (B1) enables braking of the third sun gear (S3) of the third planetary gear train (30), and the brake hub of the first brake (B1) is welded in front of the third sun gear (S3). One friction member is locked by being extended on the outer periphery of the third planetary gear train (30), and the other friction member is locked by a spline formed in front of the transmission case 1. A piston and a return spring are held in a hydraulic chamber on the outer peripheral side, which is two stages of a holding member 2a fixed to the front of the transmission case 1 by a bolt, and a hydraulic servo of the first brake (B1) is formed.

  The second brake (B2) enables braking of the fourth sun gear (S4) of the fourth planetary gear train (40) connected to the third ring gear (R3) of the third planetary gear train (30). The fourth sun gear (S4) is spline-connected to the brake hub of the second brake (B2) extended to the front of the third planetary gear train (30) at the inner circumference, and is connected to the brake hub of the second brake (B2). Is engaged with one friction member in front of the third planetary gear train (30), and the other friction member is engaged with a spline formed on a holding member 2a fixed to the transmission case 1 with a bolt. . The piston and the return spring are held in the hydraulic chamber on the inner peripheral side, which is the second stage of the holding member 2a, and the hydraulic servo of the second brake (B2) is formed.

  The fourth brake (B4) enables braking of the third planet carrier (P3) of the third planetary gear train (30) connected to the fourth planetary carrier (P4) of the fourth planetary gear train (40). One friction member is locked to the left side member of the third planet carrier (P3) extended to the rear of the outer periphery of the third planetary gear train (30), and the other friction member is in front of the transmission case 1. The first brake (B1) formed in the above is engaged with a spline extended to the rear. A brake case that holds a piston and a return spring that form a hydraulic servo of the fourth brake (B4) is fixed to the transmission case 1 behind the friction member by a retaining ring.

  The third planetary gear train (20), the first planetary gear train (10), the third clutch (C3) and the third planetary gear train (20) shown in FIG. The arrangement structure of the main transmission mechanism including the brake (B3) is the same as that of the “1-TYPE” 9AT of FIG. 14 or the “2-TYPE” 9AT of FIG. Since the tooth ratio of the ring gear and the sun gear of the row (20) is as small as 2.222, the diameter of the second sun gear (S2) is increased to increase the diameter of the second sun gear (S2) to the first connecting member (7). It has a structure that is connected to the spline.

  The structure of Fig. 17 "3-TYPE" 11AT is conceptually designed with the input power from the prime mover set to 300 Nm in the schematic diagram of the FR specification on the left side of Fig. 6 with the input power from the same prime mover set to 300 Nm. The total length of the transmission is about 8 to 10% longer than that of "1-TYPE" 9AT and 11AT of 14 and "1-TYPE" 15AT of FIG. This is a structure that connects the two constituent elements of the third and fourth planetary gear trains (30, 40) used in the front transmission mechanism, and the first, second, and fourth brakes (B1, B2, B4). ) Is due to the arrangement structure.

<3-TYPE> “C3-3-2, 11AT” FIG. 7
The FRONT GEAR (previous gear shift mechanism) of the same speed change mode as in FIG. 6 uses four ATs in which four constituent elements of two planetary gear trains are controlled by two clutches and three brakes, and two planetary gear trains are used. The combination is different from that of JP-A-2005-161311 (Honda). However, JP-A-2015-161311 (Honda) has the same shift mode as JP-A-2012-247057 (Hyundai), and can be easily assumed from FIGS. 1 and 4 of JP-A 2012-247057 (Hyundai). In both cases, a schematic diagram of the FF specifications is described, but this arrangement is not practical because the axial direction becomes too long. In addition, although the fastening elements of the third clutch (C3) are different between these patents and the present application, the capacity of the clutch is smaller in the present application. FIG. 7 is a schematic diagram of two types of C3 type 11AT, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift speed, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 7, the velocity diagram is divided into MAIN GEAR (main transmission mechanism) and FRONT GEAR (front transmission mechanism), and the velocity diagram of MAIN GEAR (main transmission mechanism) is from the right of the figure. The first, second, third, and fourth constituent elements are arranged in order, and the speed diagram of the sub front transmission mechanism of FRONT GEAR (front transmission mechanism) is the fifth, sixth, seventh, and eighth constituent elements in order from the right of the figure. (A, B, C, D) are arranged, the first component and the seventh component (C) are connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  Since the gearshift mode of the MAIN GEAR (main gearshift mechanism) is the same as that in FIG. 1, the description thereof is omitted, and the gearshift mode of the FRONT GEAR (front gear shift mechanism) will be described. In the velocity diagram of FIG. 7, the third sun gear (S3) of the third planetary gear train (30) constitutes the fifth component (A), and the third planetary carrier (P3) of the third planetary gear train (30). And a fourth ring gear (R4) of the fourth planetary gear train (40) are connected to form a sixth component (B), and the fourth ring gear (R3) and the fourth ring gear (R3) of the third planetary gear train (30) are connected to each other. The fourth planet carrier (P4) of the planetary gear train (40) is connected to form a seventh component (C), and the fourth sun gear (S4) of the fourth planetary gear train (40) is the eighth component (D). ). The input shaft and the fifth and sixth constituent elements (A, B) can be connected by the first and second clutches (C1, C2) respectively, and the eighth, fifth and sixth constituent elements (D, A, B) Can be braked by the first, second, and fourth brakes (B1, B2, B4), respectively. When the first clutch (C1) and the first brake (B1) are engaged, the largest decelerated rotation of the input shaft is output, and when the second clutch (C2) and the first brake (B1) are engaged, the next input shaft is decelerated. Outputs rotation, outputs the rotation of the input shaft by engaging the first clutch (C1) and the second clutch (C2), and speeds up the input shaft by engaging the second clutch (C2) and the second brake (B2). Rotation is output, 0 rotation is output by engaging the first brake (B1) and the fourth brake (B4), and reverse rotation of the input shaft is output by engaging the first clutch (C1) and the fourth brake (B4). To do. In other words, 6 types of input shaft rotation, input shaft deceleration rotation 2 types, input shaft speedup rotation, 0 rotation, reverse rotation can be input to MAIN GEAR (main transmission mechanism) from FRONT GEAR (front transmission mechanism). become.

  In the two types of schematic diagrams in Fig. 7, the left diagram shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right diagram shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. Here, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is increased to 3.333, and the gear ratio of the ring gear and the sun gear of the second planetary gear train (20) is increased. A good gear ratio cannot be obtained unless the gear ratio is reduced to 1.818. Therefore, the outer diameter of the first planetary gear train (10) becomes large, and the friction member of the third clutch (C3) is provided on the outer periphery of the first planetary gear train (10) as shown in "1-TYPE" 11AT in FIG. It cannot be placed and the axial direction becomes long. In addition, since the reduction ratio becomes larger, the capacity of the second planetary gear train (20) must be increased, and the shaft length becomes longer, which makes it unsuitable for FF specifications. A fluid coupling that shortens the axial direction by about 20% was used. A prime mover is located on the left front (not shown), and power is input to the input shaft of the transmission via a torque converter (fluid coupling). In the left diagram of the FR specification, the third planetary gear train (30), the fourth planetary gear train (40), the second planetary gear train (20), the first planetary gear train (20), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40), the first planetary gear train (40) In the right figure of the FF specification in which the planetary gear train (10) is arranged, the first planetary gear train (10), the second planetary gear train (20), the second planetary gear train (20), the second planetary gear train (20) A four planetary gear train (40) and a third planetary gear train (30) are arranged, and the first and second planetary gear trains (10, 20) constitute a MAIN GEAR (main transmission mechanism), and third and fourth The planetary gear train (30, 40) constitutes a FRONT GEAR (forward gear change mechanism). In the right diagram of the FF specification, an output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20) and is axially supported by a partition wall integrated with the transmission case. The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The input shaft is rotatably supported at the other end of the transmission case, and in the FR specification, the input shaft is rotatably supported at the output shaft rotatably supported at the other end of the transmission case. (C1, C2) so that the friction members are stacked in two stages in the radial direction, the third planetary gear train (30) and the fourth planetary gear train (40) of the front transmission mechanism (FRONT GEAR) on the outer peripheral direction of the cylindrical member. Distributed between the two. The three first, second, and fourth brakes (B1, B2, B4) arranged in the front transmission mechanism (FRONT GEAR) brake the components approaching in the axial direction, but the third and fourth planetary gears are used. The gear trains (30, 40) and the first and second clutches (C1, C2) are arranged on the outer circumference.

  The gear ratios shown in the table of FIG. 7 are almost the same as those of JP-A-2015-161311 (Honda), and the gear ratios of the ring gear and the sun gear of each planetary gear train are those selected by the applicant of the present application. It was devised in consideration of the gear ratio of the ring gear and sun gear of each planetary gear train. Regarding the gear ratio (RATIO), the gear ratio step value (STEP), and the gear ratio width (RANGE) indicating the performance, the gear ratio width (RANGE) is 11.07, and the gear ratio step value (STEP) is also ideal. Although it is not the target, it becomes almost appropriate as in FIG. The meshing efficiency (GEAR EFF) of the planetary gears is slightly better than that of FIG. 6, and there is no major problem in terms of performance. However, as in FIG. 6, it is better that the number of deceleration stages is seven and the number of speed-increasing stages is four, and the number of forward decelerating stages of “1-TYPE” 11AT is eight and three speed-increasing stages are three.

  The schematic diagrams of FIGS. 6 and 7 and the structural diagram of FIG. 7 represent “claim 1” and “claim 3” of the present application in “3-TYPE” and “C3-3, 11AT”. As for the combination of four constituent elements with two planetary gear trains in the FRONT GEAR (front gearbox) of FIG. 7, FIG. 7 is better than FIG. 6 in terms of the meshing efficiency and the axial length of the planetary gears. FIG. 7 is preferable for the "C3-3, 11AT" type shift mode of "TYPE". However, "1-TYPE" 11AT is better for passenger cars, and its application needs to be considered before application.

<3-TYPE> “C3-4-1, 11AT” FIG. 8
It is an 11AT that uses a 4AT in which four constituent elements consisting of two planetary gear trains are controlled by three clutches and two brakes in a FRONT GEAR (front transmission mechanism). Although the shift mode of the FRONT GEAR (frontward shift mechanism) is different from the "C3-3, 11AT" of FIGS. 6 and 7, it has been practically used for a general 4AT, and a similar gear ratio can be obtained. The one proposed in Japanese Patent Laid-Open No. 2014-77535 (Modern) is different in the engagement element of the clutch C1, but is used generally in the present application. In the arrangement of JP-A-2014-77535 (modern), the axial direction becomes too long, which is not practical. FIG. 8 is a diagram showing two types of schematic diagrams, a speed diagram showing a shift mode, a fastening element (SHIFT) at each shift stage, and a gear ratio (RATIO). In the velocity diagram of FIG. 8, the velocity diagram is divided into a MAIN GEAR (main transmission mechanism) and a FRONT GEAR (front transmission mechanism), and the MAIN GEAR (main transmission mechanism) speed diagram is from the right of the figure. The first, second, third, and fourth constituent elements are arranged in order, and the speed diagram of the sub front transmission mechanism of FRONT GEAR (front transmission mechanism) is the fifth, sixth, seventh, and eighth constituent elements in order from the right of the figure. (A, B, C, D) are arranged, the first component and the sixth component (B) are connected by the first connecting member (7), and the third component becomes the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  Since the gearshift mode of the MAIN GEAR (main gearshift mechanism) is the same as that in FIG. 1, the description thereof is omitted, and the gearshift mode of the FRONT GEAR (front gear shift mechanism) will be described. The four constituent elements of the FRONT GEAR (front transmission mechanism) are the same as those in FIG. 7, and the input shaft and the fifth, seventh, and eighth constituent elements (A, C, D) are respectively the first, second, and The fourth clutch (C1, C2, C4) can be connected, and the seventh and eighth constituent elements (C, D) can be braked by the first and second brakes (B1, B2), respectively. When the first clutch (C1) and the first brake (B1) are engaged, the largest decelerated rotation of the input shaft is output, and when the first clutch (C1) and the second brake (B2) are engaged, the next input shaft is decelerated. Outputs rotation, outputs rotation of input shaft by engaging first clutch (C1) and second clutch (C2), and speeds up input shaft by engaging second clutch (C2) and second brake (B2). Rotation is output, 0 rotation is output by engaging the first brake (B1) and the second brake (B2), and reverse rotation of the input shaft is output by engaging the fourth clutch (C4) and the first brake (B1). To do. In other words, 6 types of input shaft rotation, input shaft deceleration rotation 2 types, input shaft speedup rotation, 0 rotation, reverse rotation can be input to MAIN GEAR (main transmission mechanism) from FRONT GEAR (front transmission mechanism). become.

  In the two types of schematic diagrams in Fig. 8, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. Here, the tooth ratio of the ring gear and the sun gear of the third planetary gear train (30) and the fourth planetary gear train (40) is replaced with that in FIG. 7, and the distance between the constituent elements A and B of the velocity diagram and C- FIG. 8 shows that the distance between D is reversed. If the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) of the MAIN GEAR (main transmission mechanism) is the same as in FIG. 7, the arrangement of the planetary gears is the third planetary gear train (30) and the fourth planetary gear. The gear train (40) is only the reverse of FIG. In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The other end of the transmission case pivotally supports the input shaft, and in the FR specification, the output shaft pivotally supported at the other end of the transmission case pivotally supports the input shaft. The first, second, and fourth clutches (C1, C2, C4) are arranged such that the second and fourth clutches (C2, C4) are arranged so that the friction members are radially stacked in two stages and the first clutch ( C1) are arranged in the axial direction, and are arranged between the fourth planetary gear train (40) and the third planetary gear train (30) of the front speed change mechanism on the outer peripheral direction outer side of the cylindrical member. The two first and second brakes (B1, B2) arranged in the front transmission are the outer circumferences of the fourth planetary gear train (40) and the first, second and fourth clutches (C1, C2, C4). Will be distributed to. Since three clutches are arranged in the front transmission mechanism, the axial direction becomes longer than 11 AT of "1-TYPE" or "2-TYPE".

  The schematic diagram of FIG. 8 represents "claim 1" and "claim 3" of the present application. The gear ratio shown in the table of FIG. 8 is exactly the same as that of FIG. 7, and there is no major problem in terms of performance. However, since three clutches are used in the front transmission mechanism, the cost is high and the axial direction is long. Therefore, the embodiment is proposed as a reference.

<4-TYPE>
In the multi-stage transmission described in paragraph “0016”, the shift mode is similar to “1-TYPE” 9AT. In "1-TYPE" 9AT, one type of deceleration rotation speed can be input from the auxiliary front speed change mechanism to the main front speed change mechanism of the front speed change mechanism, but in "4-TYPE", two speeds can be input from the auxiliary front speed change mechanism. The difference is that it is possible to input the deceleration rotation speed of the seed. Since the main front speed change mechanism has the same transmission mode, one brake and one planetary gear train are provided for each of the same main front speed change mechanism and sub front speed change mechanism as in "1-TYPE" 9AT. In addition to 2 types of deceleration rotation speeds that can be input to the main front speed change mechanism, there are 8 kinds of acceleration rotation, deceleration rotation, reverse rotation and input shaft rotation and 0 rotation of each 2 types of input shafts. The rotation of is input to the main transmission mechanism to be 14AT. This method has not been disclosed yet, and the structure of FR specification and FF specification in "C3-5-1, 14AT" of FIG. 9, FIG. 18, and FIG. An example is given. Note that FIG. 10 shows a plurality of planetary gear trains that can be used for the auxiliary front transmission mechanism. Since "4-TYPE" has a gear position of 14AT which is not necessary for passenger cars, FIG. 18 which is a structural diagram of FR specification is the only commercial vehicle for trucks and buses in the present embodiment. The structural diagram of FIG. 15 of "1-TYPE" 15AT for passenger cars of FR specification is already the concept for trucks and buses in Patent Document 2, and the size and size of one embodiment of other passenger cars 9AT-11AT. If the concept of the "4-TYPE" 14AT is for passenger cars, the size is equal to or smaller than that of the "1-TYPE" 15AT, and these 14AT and 15AT are extremely compact. Since the structural drawing of "1-TYPE" 15AT is not shown as FF specification, the structural drawing of "4-TYPE" 14AT for passenger car FF is shown in FIG.

<4-TYPE> “C3-5-1, 14AT” FIGS. 9, 18, and 23.
FIG. 9 shows two schematic diagrams of 14AT, a velocity diagram showing a shift mode, a fastening element (SHIFT) at each shift stage, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is a thing. In the speed diagram of FIG. 9, the speed diagram is divided into MAIN GEAR (main speed change mechanism) and FRONT GEAR (front speed change mechanism). The speed map of MAIN GEAR (main speed change mechanism) is from the right of the figure. The 1st, 2nd, 3rd, and 4th constituent elements are arranged in order, and the speed diagram of the sub-front transmission mechanism of FRONT GEAR (previous transmission mechanism) is the 5th, 6th, 7th, and 11th constituent elements in order from the right of the figure. (A, B, C, G) are arranged, and the velocity diagram of the main front transmission mechanism is arranged such that the eighth, ninth and tenth constituent elements (D, E, F) are arranged in order from the right side of the drawing, and the sixth diagram is shown. The component (B) and the eighth component (D) are connected by the second connecting member (8), the first component and the tenth component (F) are connected by the first connecting member (7), and The three components are the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed. This arrangement is exactly the same as in FIG.

  In the two types of schematic diagrams in Fig. 9, the left figure shows an FR specification gear train suitable for passenger cars (Passenger Car) and commercial vehicles (Truck Bus), and the right figure shows an FF specification gear train suitable for passenger cars (Passenger Car). is there. A prime mover is located on the front left side (not shown), and power is input to the input shaft of the transmission via the torque converter. In the left diagram of the FR specification, the fourth planetary gear train (40), the fifth planetary gear train (50), and the third planetary gear train, which are two-story structures consisting of simple planetary gears, are axially arranged from the front left of the transmission. The second planetary gear train (20) and the first planetary gear train (10) are arranged, and in the right diagram of the FF specification, the first planetary gear train (10) including the simple planetary gears in the axial direction from the left front of the transmission. A second planetary gear train (20), a third planetary gear train (30), a four-storey fourth planetary gear train (40), and a fifth planetary gear train (50), and the first and second planetary gear trains (50) are arranged. The planetary gear trains (10, 20) form a MAIN GEAR (main transmission mechanism), and the third planetary gear train (30), the fourth planetary gear train (40) and the fifth planetary gear train (50) are two-story. ) Constitutes a FRONT GEAR (front transmission). Also, the second planetary gear train (40) and the fifth planetary gear train (50), which are two-story structures, constitute a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), and the third planetary gear train (30). ) Constitutes the main front shifting mechanism. In the right diagram of the FF specification, an output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20) and is axially supported by a partition wall integrated with the transmission case. The first, second, third, fourth and fifth planetary gear trains (10, 20, 30, 40, 50) include first, second, third, fourth and fifth sun gears (S1, S2, S3, S4, S5), first, second, third, fourth, fifth planetary carriers (P1, P2, P3, P4, P5), first, second, third, fourth, fourth 5 ring gears (R1, R2, R3, R4, R5). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end on the front transmission mechanism side (FRONT GEAR) of the transmission case. The input shaft is rotatably supported at the other end of the transmission case, and in the FR specification, the input shaft is rotatably supported at the output shaft rotatably supported at the other end of the transmission case. C1 and C2) are the second floor of the outer peripheral direction of the cylindrical member so that the friction members are stacked in two steps in the radial direction. The fourth planetary gear train (40), the fifth planetary gear train (50) and the third planetary gear train (50). The third planetary gear train (30) of the front transmission mechanism (FRONT GEAR) is arranged between the gear trains (30) in the axial direction of the main transmission mechanism side (MAIN GEAR) from the inner peripheral end of the cylindrical member. Will be distributed.

  In the two types of schematic diagrams and velocity diagrams of FIG. 9, the first and second sun gears (S1, S2) of the first and second planetary gear trains (10, 20) forming the MAIN GEAR (main transmission mechanism) are shown. The first component, the first planet carrier (P1) is the second component, the first ring gear (R1) and the second planet carrier (P2) are the third component, and the second ring gear (R2) is As a fourth component, the first sun gear (S1) of the fourth planetary gear train (40) and the fifth of the fifth planetary gear train (50) that constitute the sub-front gearbox of FRONT GEAR (front gearbox). The ring gear (R5) is the fifth component (A), and the fourth and fifth planetary carriers (P4, P5) of the fourth and fifth planetary gear trains (40, 50) are the sixth component (B). , The fourth ring gear (R4) of the fourth planetary gear train (40) is connected to the seventh component (C ), The fifth sun gear (S5) of the fifth planetary gear train (50) is the eleventh component (G), and the third ring gear (of the third planetary gear train (30) that constitutes the main front transmission mechanism (30). R3), the third planet carrier (P3), and the third sun gear (S3) are the eighth, ninth, and tenth constituent elements (D, E, F).

  Here, the first and second sun gears (S1, S2) forming the first constituent element of the MAIN GEAR (main transmission mechanism) are connected and also connected to the first connecting member (7) to form the second constituent element. The first planet carrier (P1) that constitutes the first planetary carrier (P1) is connected to the input shaft, the second planetary carrier (P2) that constitutes the third component is connected to the output shaft, and the third clutch (C3) forms the first ring gear. The second ring gear (R2), which is connectable to (R1) and constitutes the fourth component, can be braked by the third brake (B3). The fifth component (A), which is a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism), is connected to the input shaft, and the sixth component (B) is connected to the second connecting member (8). The seventh constituent element (C) can be braked by the fourth brake (B4), the eleventh constituent element (G) can be braked by the first brake (B1), and the eighth constituent element serves as a main front shifting mechanism. (D) is connected to the second connecting member (8), and the ninth component (E) can be connected to the input shaft by the first clutch (C1) and can be braked by the second brake (B2). The tenth constituent element (F) is connected to the first connecting member (7), and the eighth constituent element (D) and the ninth constituent element (E) can be connected by the second clutch (C2).

  A shift mode of the FRONT GEAR (front shift mechanism) will be described. In the velocity diagram of the fifth, sixth, seventh, and eleventh constituent elements (A, B, C, G) configuring the auxiliary front speed change mechanism, one of the first and fourth brakes (B1, B4) By one fastening, two kinds of decelerated rotations of the input shaft are input from the sixth component (B) to the eighth component (D) serving as the main pre-shift mechanism via the second connecting member (8). It The decelerated rotation input to the eighth component (D) is output as it is as the third planetary gear train (30) is integrated by engaging the second clutch (C2) of the main front transmission mechanism, and the ninth component ( The tenth constituent element (F) is accelerated and output by engaging the first clutch (C1) of E), and the tenth constituent element (F) is engaged by engaging the second brake (B2) of the ninth constituent element (E). ) Is reversed and output, and a total of three types of rotations can be output from the tenth component (F) to the first component of the MAIN GEAR (main transmission mechanism) via the first connecting member (7). Since two types of decelerated rotations are input to the main front transmission from the sub front transmission, a total of six types of rotation can be output to the MAIN GEAR (main transmission). In addition, since the third planetary gear train (30) is integrated by engaging the second clutch (C2) of the main front transmission mechanism, any one of the first clutch (C1) and the second brake (B2) With the engagement of, the two types of rotation of the input shaft and 0 rotation can be output to the MAIN GEAR (main transmission mechanism). In other words, each of the two of the first and second clutches (C1, C2), the first, second, and fourth brakes (B1, B2, B4) of the FRONT GEAR (front shifting mechanism) can be engaged to obtain two. It is possible to output eight kinds of rotations, namely, speed increase rotation, deceleration rotation, reverse rotation, input shaft rotation and zero rotation of the input shaft to the MAIN GEAR (main transmission mechanism).

  The MAIN GEAR (main transmission mechanism) has the same shift mode as that of the "1-TYPE" 9AT shown in FIG. The table of FIG. 9 shows the engagement element (SHIFT) and the gear ratio (RATIO), and the meshing efficiency (GEAR EFF) of the planetary gears at each gear. Compared with "1-TYPE" 15AT shown in FIG. 3, the range of the gear ratio (RANGE) is 18.40, which is the same as 18.54 of "1-TYPE" 15AT, and the step value of the gear ratio (STEP). Is slightly better. The meshing efficiency (GEAR EFF) of the planetary gears is also improved, but the gear ratio (RATIO) is 9.4 to 0.511, which is not so good as the 1-TYPE 15AT's 11.199 to 0.604. . As a comprehensive evaluation, a gear ratio suitable for a multi-stage transmission can be obtained, the meshing efficiency (GEAR EFF) of the planetary gears is good, and it can be said that the transmission has excellent performance like "1-TYPE" 15AT.

  Fig. 18 "C3-5-1, 14AT (FR)" is a structural diagram that conceptually designed the FR specification diagram on the left of Fig. 9 for commercial vehicles such as trucks and buses with an input power of 1000 Nm from the prime mover Is. In the field of heavy-duty commercial vehicles, vehicle manufacturers, engine manufacturers, and transmission manufacturers are different in the world, and therefore mounting and mounting are standardized by SAE. In FIG. 18, SAE NO1 Housing is used for the transmission case. A torque converter is generally used for this joint, but a fluid coupling having no torque amplification effect, a hydro damper for absorbing torque fluctuation, or a motor generator for HEV may be used. A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened by a bolt to the front of the integrated main case 1a, and a dry joint side is provided on the left front side of the charging pump. A partition wall 5a that separates the wet transmission side is fastened to the main case 1a. The charging pump is directly driven by a gear from a prime mover via a joint, and this gear drives a gear for a PTO (Power Take Off) not shown. The PTO is an indispensable work device for a specially equipped vehicle, and it must be directly driven by a prime mover, and such a transmission must be equipped with the PTO device. The holding member 2a has an L-shaped cross section, and a cylindrical holding member 2b is fastened to a cylindrical inner peripheral cylindrical portion, and both ends of the holding member 2b hold the needle roller roller bearings 4a and 4b. The input shaft 3a, to which power is input from the prime mover through the joint, is pivotally supported. Further, taper roller bearings 4f and 4e are mounted on the rear case 1b so as to be back-to-back, and support the output shaft 3c. An input shaft 3a is arranged at the center of rotation of the transmission, and its rear end is rotatably supported by a needle roller roller bearing 4d arranged on the inner peripheral portion of the output shaft 3c. Here, the input shaft 3a is pivotally supported at three points.

  On the outer periphery of the cylindrical portion of the holding member 2a, from the left front, a two-story fourth planetary gear train (40) and a fifth planetary gear train (50) that serve as a sub-front transmission mechanism of FRONT GEAR (front transmission mechanism). And the first and second clutches (C1, C2) are arranged. The second-floor fourth planetary gear train (40) and the fifth planetary gear train (50) selectively reduce the rotation of the input shaft 3a into two types and serve as a main front transmission mechanism. 30). In the two-story fourth planetary gear train (40) and fifth planetary gear train (50), the fifth planetary gear train (50) is located on the first floor and the fourth planetary gear train (40) is located on the second floor. The fifth ring gear (R5) and the first sun gear (S1) are integrated, and the left side members of the fifth planet carrier (P5) and the fourth planet carrier (P4) are integrated and arranged compactly. , The right side member of the fifth planetary carrier (P5) has its inner peripheral portion extended backward and is axially supported by the needle roller roller bearing 4j on the outer periphery of the cylindrical portion of the holding member 2a. A second connecting member (8) extending rearward along the inner circumference of the main case 1a is welded to the side member. The fifth ring gear (R5) and the first sun gear (S1), which are integrated between the right side member of the fifth planet carrier (P5) and the fourth planet carrier (P4), are connected to the input shaft 3a by an input splined. The connecting member (Y) is spline-connected. The fifth sun gear (S5) of the fifth planetary gear train (50) is rotatably supported on the outer circumference of the cylindrical portion of the holding member 2a by a needle roller roller bearing 4l arranged on the inner circumference, and at the same time, on the first brake (B1). The brake hub extends to the outer circumference of the fourth planetary gear train (40) on the second floor and is spline-connected. Further, the fourth ring gear (R4) of the fourth planetary gear train (40) is extended rearward to form a spline for locking the friction member of the fourth brake (B4) on the outer circumference, and at the front inner circumference. The plates that regulate the axial direction are spline-connected by the thrust needle bearings.

  First and second clutches (C1, C2) are arranged on the outer peripheral end of the cylindrical portion of the holding member 2a behind the fourth planetary gear train (40) and the fifth planetary gear train (50), which are two stories. It The first clutch (C1) can connect the input shaft 3a and the third planet carrier (P3) of the third planetary gear train (30), and the second clutch (C2) connects the third planet carrier (P3) and the third planet carrier (P3). The second connecting member (8), which is connected to the ring gear (R3), can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The left side member of the third planetary carrier (P3) forming the third planetary gear train (30) is welded to the rear end of the third planetary gear train (30). A servo mechanism for pressing the friction member of the first clutch (C1), which is arranged on the clutch hub integral with the input coupling member (Y) at the inner circumference, is disposed in front of the dual clutch coupling member (X). The left side member of the planet carrier (P4) has a servo mechanism for pressing the radial outer periphery of the double clutch connecting member (X) and the friction member of the second clutch (C2) arranged on the second connecting member (8). Will be distributed. The input connecting member (Y) is spline-connected to the input shaft 3a behind the outer peripheral end of the cylindrical portion of the holding member 2a, and is provided on the outer periphery of the right side member of the fifth planetary carrier (P5) of the fifth planetary gear train (50). It is pivotally supported by the bush 4c. A partition plate is provided between the double clutch connecting member (X) and the left side member of the third planet carrier (P3) to supply the servomechanism arranged above the left side member of the third planet carrier (P3) in the radial direction. A passage for the working oil of the piston and the hydraulic cancel oil is formed. The servo mechanism of the second clutch (C2) arranged on the upper side in the radial direction of the dual clutch coupling member (X) and the servo mechanism of the first clutch (C1) arranged on the lower side in the radial direction include the outer circumference of the cylindrical member 2b. Is supplied with hydraulic oil and hydraulic cancel oil through an oil passage sealed by a piston ring provided in the input connecting member (Y). Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism for the second clutch (C2) of the left side member of the third planetary carrier (P3), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 6" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

  A third planetary gear train (30) serving as a main front transmission mechanism of the FRONT GEAR (front transmission mechanism) is arranged behind the cylindrical portion of the holding member 2a and the first and second clutches (C1, C2). . The third planetary gear train (30) selectively transmits eight kinds of rotations to the first and second planetary gear trains (10, 20) which are main transmission mechanisms. The third ring gear (R3) of the third planetary gear train (30) is connected to the second connecting member (8) welded to the right side member of the fourth planetary carrier (P4) of the fourth planetary gear train (40). Splined. The left side member of the third planet carrier (P3) supporting the planet pinion gear that meshes with the third ring gear (R3) serves as a clutch cover for the second clutch (C2) and is welded to the dual clutch connecting member (X). The right side member is extended on the outer circumference of the third ring gear (R3). The output third sun gear (S3) is spline-connected to the first connecting member (7) at the inner circumference. Here, the structures of the first and second clutches (C1, C2) and the third planetary gear train (30) are shown in FIG. 14 "C3-1-1, 9AT, C3-1-2," of "1-TYPE". 11AT (FR) ".

  The first brake (B1) enables braking of the fifth sun gear (S5) of the fifth planetary gear train (50). A thin plate brake hub welded in front of the fifth sun gear (S5) is extended to the outer circumference of the two-story fourth planetary gear train (40), and the first brake (B1) is attached to the spline formed on the outer circumference. One of the friction members is locked. The other friction member of the first brake (B1) is locked to a spline formed in front of the main case 1a, and a piston and a return spring are provided in a hydraulic chamber of a holding member 2a fixed to the front of the main case 1a by a bolt. Is held, and the hydraulic servo of the first brake (B1) is formed.

  The fourth brake (B4) enables braking of the fourth ring gear (R4) of the fourth planetary gear train (40). One of the friction members of the fourth brake (B4) is locked to the outer peripheral spline extending to the rear of the fourth ring gear (R4). The other friction member of the fourth brake (B4) is locked to the extended spline that locks the friction member of the first brake (B1) formed in front of the main case 1a, and the fourth brake (B4) is locked. The brake case, which holds the piston and the return spring forming the hydraulic servo, is fixed to the main case 1a behind the friction member with a retaining ring.

  The second brake (B2) enables braking of the third planet carrier (P3) of the third planetary gear train (30). The right side member of the third planet carrier (P3) is extended to the outer circumference of the third ring gear (R3), and one friction member of the second brake (B2) is locked to the spline formed on the outer circumference. The other friction member of the second brake (B2) is locked to the spline formed in the main case 1a, and the piston forming the hydraulic servo of the second brake (B2) and the return spring are connected to the fourth brake (B4). It is arranged in a brake case that holds the hydraulic servo of the fourth brake (B4) as opposed to the hydraulic servo.

  In the second planetary gear train (20) arranged behind the third planetary gear train (30), the second sun gear (S2) extends to the inner circumference of the first sun gear (S1) of the first planetary gear train (10). The rolled material is rotatably arranged on the outer periphery of the input shaft 3a by the needle roller roller bearing 4h, and is stretched forward as the first connecting member (7) to the fourth sun gear (S4) of the fourth planetary gear train (40). Splined. The planetary pinion gear that meshes with the second sun gear (S2) is supported by the second planetary carrier (P2), and the right side member is arranged on the outer periphery of the first planetary gear train (10) spline-coupled to the output shaft 3c. The output hub 9 is splined. The second ring gear (R2) meshing with the planetary pinion gear has a plate in which the left end tooth portion is axially fixed by a retaining ring and is spline-connected to the inner periphery of the left side member of the second planetary carrier (P2). Along with the second sun gear (S2) of the second planetary gear train (20) that has been extended, it is rotatably arranged with its axial direction restricted by a thrust needle bearing. The friction member of the third brake (B3) is locked to the outer periphery of the second ring gear (R2), but the center of gravity of both is within the width of the planet pinion gear of the second planetary gear train (20). Since the radial load of the second ring gear (R2) and the friction member is received by the two planet carrier (P2), the radial bearing dedicated to the second ring gear (R2) is not required.

  The first planetary gear train (10) arranged behind the second planetary gear train (20) is the second sun gear of the second planetary gear train (20) in which the first sun gear (S1) is extended to the inner circumference. (S2) is spline-connected. The planetary pinion gear that meshes with the first sun gear (S1) is supported by the first planetary carrier (P1), and the right side member is spline-connected to the input shaft 3a at the inner circumference of the first sun gear (S1). The first ring gear (R1) that meshes with the planetary pinion gear has a spline formed on the outer periphery for locking the friction member of the third clutch, and a plate whose axial direction is regulated by a thrust needle bearing is welded.

  The third clutch (C3) can connect the first ring gear (R1) of the first planetary gear train (10) and the output shaft 3c. One friction member is locked to a spline formed on the outer periphery of the first ring gear (R1) of the first planetary gear train (10), and a third clutch (C3) is attached to the spline formed on the inner periphery of the output hub 9. ) The other friction member is locked. The output shaft 3c and the right side member of the second planetary carrier (P2) of the second planetary gear train (20) are connected to this spline. The piston of the third clutch (C3) and the return spring are held in front of the output shaft 3c, and a hydraulic servo of the third clutch (C3) is formed. The output shaft 3c is rotatably supported by tapered roller bearings 4e and 4f that are back-to-back in the rear portion of the main case 1a. It receives the supply of hydraulic oil from the clutch (C3).

  The third brake (B3) enables braking of the second ring gear (R2) of the second planetary gear train (20). The second ring gear (R2) has a spline formed on the outer peripheral portion thereof to lock one friction member of the third brake (B3). A spline is formed on the inner circumference of the transmission case 1, and the other friction member of the third brake (B3) is locked. Further, the piston and the return spring of the third brake (B3) are held in the main case 1a behind the flange of the output shaft 3c, and a hydraulic servo of the third brake (B3) is formed. The piston of the third brake (B3) presses the friction member of the third brake (B3) on the outer circumference of the output hub 9. Since the capacity of the third brake (B3) is large, the area of the hydraulic chamber that presses the friction member is increased to generate a large pressing force, and the hydraulic chamber is doubled to improve the response, and the hydraulic chamber having a small area is used. When oil is supplied to the hydraulic chamber, a check valve is provided in the hydraulic chamber having a large area to prevent the hydraulic chamber having a large area from becoming a vacuum and hindering the movement of the piston.

  Since FIG. 18 is a structural diagram conceptually designed for commercial vehicles such as trucks and buses with an input power from a prime mover of 1000 Nm, the same 1000 Nm truck described in FIG. It must be compared with the C3-15AT, which was designed conceptually for commercial vehicles such as buses. In the C3-14AT of FIG. 18, the input power of the third planetary gear train (30) and the fifth planetary gear train (50) is input to the ring gear having a large diameter, and "load = torque / radius" is satisfied. The load on the planetary gear trains of the planetary gear train (30) and the fifth planetary gear train (50) can be reduced, and the tooth width can be set to a small value. The first and second clutches (C1, C2) according to claim 4 can be set. Combined with the structure and the reduction ratio, the axial direction becomes compact by about 5% compared to C3-15AT. In the world, a tractor is generally used as a heavy-duty truck, and an MT having 10 to 16 forward steps is used as a tractor head. In Japan, a cargo truck called a 10-ton car is often used, and an MT having seven gear stages and a gear ratio width of about 10 is mainly used. Therefore, the "1-TYPE" 9AT, 11AT of the present application is sufficient for this cargo truck, but if the 14AT having a large speed ratio of 1.8 times in FIG. 18 is used, which is not so large as 9AT, 11AT, The rotation speed of can be reduced to 1 / 1.8. In addition, since the gear efficiency is 0.5 to 1% better than MT, the fuel consumption is better than MT even if the loss of the charging pump is taken into consideration. This is because the engine control is an all-speed governor control that produces full torque regardless of the accelerator pedal opening, so the charging pump loss is less than 1%. ATs are required for large commercial vehicles that require automatic driving rather than passenger vehicles.

  FIG. 23 “C3-5-1, 14AT (FF)” is a structural diagram conceptually designing the schematic diagram of the FF specification shown on the right side of FIG. 9 with the input power from the prime mover set to 300 Nm. FF) and 11AT (FF) shown in FIG. 21, almost no more than 400 mm. From the main transmission mechanism including the first planetary gear train (10), the second planetary gear train (20), the third clutch (C3), and the third brake (B3), and the output counter gear 5 axially supported by the partition wall. The output structure and the like are all the same as 11AT (FF) in FIG.

  The transmission case 1c for closing the rear part of the transmission has an integral inner peripheral portion protruding forward in a cylindrical shape, and a needle roller roller bearing 4b is arranged at the inner peripheral end portion of the protruding cylindrical member to provide the input shaft 3 To support. Therefore, the shaft support of the input shaft 3 is shortened, and rotational vibration can be suppressed. On the outer circumference of the projecting cylindrical member, from the right rear, a two-story fourth planetary gear train (40), a fifth planetary gear train (5), and a second front gear train (40) that serve as a sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). First and second clutches (C1, C2) are arranged, and a third planetary gear train (30) is arranged in front of the protruding cylindrical member and the first and second clutches (C1, C2). The two-story fourth planetary gear train (40) and the fifth planetary gear train (5) selectively use two kinds of decelerated rotations of the input shaft 3 for the third planetary gear train (30) serving as a main front transmission mechanism. Communicate to. The fifth planetary carrier (P5) of the fifth planetary gear train (50), which is arranged on the first floor of the two-story building, has an inner diameter that is cylindrically extended on the inner circumference of the left side member and is speed-shifted by the needle roller roller bearing 4j. The fifth sun gear (S5) is rotatably held on the outer circumference of the cylindrical member of the machine case 1c, and further, the fifth sun gear (S5) is rotatably held on the outer circumference of the cylindrical member of the transmission case 1c by a bush 4l pressed into the inner circumference. A thin plate-shaped brake hub is spline-connected to the rear side of the fifth sun gear (S5), and is extended to the outer periphery of the fourth planetary gear train (40) arranged on the second floor of the two-story building to extend the first brake (B1). ) The friction member is locked. The right side member of the fifth planet carrier (P5) is integrated with the right side member of the fourth planet carrier (P4) of the fourth planetary gear train (40) and is forward of the left side member of the fourth planet carrier (P4). The second connecting member (8) that has been rolled into a sheet is welded. The fifth ring gear (R5) and the first sun gear (S1), which are integrated between the left side member of the fifth planet carrier (P5) and the fourth planet carrier (P4), are connected to the input shaft 3 by splines. The connecting member (Y) is spline-connected. The fourth ring gear (R4) of the fourth planetary gear train (40) has a spline formed on the outer periphery for locking the friction member of the fourth brake (B4), and an axial direction formed by a thrust needle bearing on the inner periphery at the rear. The plates that regulate the are connected by splines.

  First and second clutches (C1, C2) are arranged at the outer peripheral ends of the cylindrical members in front of the fourth planetary gear train (40) and the fifth planetary gear train (50), which are two stories. The first clutch (C1) enables the input shaft 3 and the third planet carrier (P3) of the third planetary gear train (30) to be connected, and the second clutch (C2) connects the third planet carrier (P3) and the third planet carrier (P3). The second connecting member (8), which is connected to the ring gear (R3), can be connected. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which each friction member is coaxially superposed in two steps in a radial direction is rotatable by a bush 4k on the outer periphery of the input connecting member (Y). The right side member of the third planetary carrier (P3) constituting the third planetary gear train (30) is welded to the front end of the third planetary gear train (30). A servo mechanism for pressing the friction member of the first clutch (C1), which is arranged on the clutch hub integrated with the input coupling member (Y) on the inner periphery, is disposed behind the dual clutch coupling member (X). The right side member of the planet carrier (P3) has a servo mechanism for pressing the radial outer periphery of the double clutch connecting member (X) and the friction member of the second clutch (C2) arranged on the second connecting member (8). Will be distributed. The input coupling member (Y) is spline-coupled to the input shaft 3 in front of the cylindrical member of the transmission case 1c, and is bushed on the outer periphery of the left side member of the fifth planetary carrier (P5) of the fifth planetary gear train (50). It is pivotally supported by 4c. A partition plate is provided between the double clutch connecting member (X) and the right side member of the third planet carrier (P3) to supply the servo mechanism arranged on the radial upper side of the right side member of the third planet carrier (P3). A passage for the working oil of the piston and the hydraulic cancel oil is formed. The servo mechanism of the second clutch (C2) arranged on the upper side in the radial direction of the dual clutch coupling member (X) and the servo mechanism of the first clutch (C1) arranged on the lower side in the radial direction include the outer circumference of the cylindrical member 2b. Is supplied with hydraulic oil and hydraulic cancel oil through an oil passage sealed by a piston ring provided in the input connecting member (Y). Further, an O-ring is arranged on the axial support portion of the planetary pinion gear of the hydraulic chamber that serves as the servo mechanism of the second clutch (C2) of the right side member of the third planetary carrier (P3), and seals the hydraulic chamber. The structures of the first and second clutches (C1, C2) are "claim 2", "claim 5", and "claim 6" of the present application, which are extremely simple and compact, and the first and second clutches ( The oil passage to C1 and C2) also has a small resistance to the pipeline, and the responsiveness of the clutch is improved.

  A third planetary gear train (30) serving as a main front transmission mechanism of FRONT GEAR (front transmission mechanism) is arranged in front of the cylindrical member of the transmission case 1c and the first and second clutches (C1, C2). It The third planetary gear train (30) selectively transmits eight kinds of rotations to the first and second planetary gear trains (10, 20) which are main transmission mechanisms. The third ring gear (R3) of the third planetary gear train (30) is attached to the second connecting member (8) welded to the left side member of the fourth planetary carrier (P4) of the fourth planetary gear train (40). Splined. The right side member of the third planet carrier (P3) supporting the planet pinion gear meshing with the third ring gear (R3) is welded to the dual clutch connecting member (X) by the right side member serving as the clutch cover of the second clutch (C2). A brake hub of the second brake (B2) extended to the outer periphery of the third ring gear (R3) is welded to the left side member. The output third sun gear (S3) is spline-connected to the first connecting member (7) at the inner circumference.

  The first brake (B1) enables braking of the fifth sun gear (S5) of the fifth planetary gear train (50). A thin plate brake hub welded to the rear of the fifth sun gear (S5) is extended to the outer periphery of the two-story fourth planetary gear train (40), and the first brake (B1) is attached to the spline formed on the outer periphery. One of the friction members is locked. The other friction member of the first brake (B1) is locked to a spline formed in the transmission case 1c, and a piston and a return spring are held in a hydraulic chamber provided behind the transmission case 1c. A hydraulic servo of the brake (B1) is formed.

  The fourth brake (B4) enables braking of the fourth ring gear (R4) of the fourth planetary gear train (40). One friction member of the fourth brake (B4) is locked to the outer peripheral spline of the fourth ring gear (R4). The other friction member of the fourth brake (B4) is locked to the extended spline that locks the friction member of the first brake (B1) formed in the transmission case 1c, and the fourth brake (B4). The brake case that holds the piston and the return spring forming the hydraulic servo is fixed to the transmission case 1b in front of the friction member with a retaining ring.

  The second brake (B2) enables braking of the third planet carrier (P3) of the third planetary gear train (40). The left side member of the third planet carrier (P3) is welded with the brake hub of the second brake (B2) extended to the outer periphery of the third ring gear (R3), and the spline is formed on the outer periphery of the brake hub. One friction member of the second brake (B2) is locked. The other friction member of the second brake (B2) is locked to the spline formed in the transmission case 1b, and the piston and the return spring forming the hydraulic servo of the second brake (B2) are connected to the fourth brake (B4). It is arranged in a brake case that holds the hydraulic servo of the fourth brake (B4) as opposed to the hydraulic servo of.

  The structure of 14AT (FF) in FIG. 23 is almost the same as 11AT (FF) in FIG. 21, and is a size that can be realized as FF including 9AT (FF) in FIG. Normally, a passenger car does not require such a multi-stage transmission, but 14AT (FF) in FIG. 23 has a gear ratio range of 18.40, which is 3 times that of 6AT and 2 times that of 9AT, which is an existing concept. It has different characteristics. Therefore, the engine speed should be 1000 to 2000 RPM, and if an engine with good fuel consumption can be developed with this specification, it will bring a great change.

  FIG. 10 shows the configuration of a sub-front transmission mechanism including four constituent elements A, B, C, and G of the FRONT GEAR (front transmission mechanism) of FIG. 9 in the "4-TYPE" 14AT. There are many other methods of outputting two kinds of decelerated rotations of the input shaft by using two planetary gear trains and two fastening elements, and any one may be used. "4-TYPE" is an auxiliary front speed change mechanism, and if the deceleration rotation of the input shaft is increased by one type, the main front speed change mechanism can output the rotation increased by three types to the MAIN GEAR (main speed change mechanism). The GEAR (main transmission mechanism) can have five types of gears. In other words, this is a system that can easily be made to be "4-TYPE" 19AT and is easy to be multistage.

<5-TYPE>
In the multi-stage transmission described in paragraph “0017”, there is an original patent of US20090054196 (GM), which is described in FIG. 11 of the present application. The C3 type is used as a base for the shift mode, and a 10-speed AT is added to the C3 type, which is one planetary gear train and the same number of fastening elements as the C3 type 9AT. The arrangement structure of the first and second planetary gear trains (10, 20), the third clutch (C3), and the third brake (B3) of the MAIN GEAR (main transmission mechanism) is the same as "1 to 4-TYPE". . The "5-TYPE" has 10 planetary gear trains and the same number of fastening elements as 9AT, which is one more gear stage, and there are many combinations of FRONT GEAR (previous gear mechanism), and the gear stages are also speed-up stages. At first glance, it looks excellent because it has a good balance with 3 steps. The AT of "1 to 4-TYPE" described as the C3 type AT has the MAIN GEAR (main transmission mechanism) in which the transmission mode is the third brake (B3) engaged state and the third clutch (C3) engaged state. It is possible to distinguish between two states, and it means that there is only one transformation point at which the shift mode changes. On the other hand, in the AT of "5-TYPE", in addition to the two states of the MAIN GEAR (main transmission mechanism), that is, the B3 brake is engaged and the C3 clutch is engaged, the C1 clutch can be engaged. There are two transformation points. Since the shift mode changes depending on the transformation point, the possibility that the chain of gear ratios deteriorates increases, and the D type described in paragraph “0006” corresponds to this example, and it is difficult to obtain a good gear ratio. FRONT GEAR (front gearbox) starts the gear from the rotation and speedup of the input shaft, while MAIN GEAR (main gearbox) uses a large gear ratio of the ring gear and sun gear that is not used in a normal AT, to improve efficiency. Has slowed down a lot. Since a large step value is desired in the low speed stage of the transmission, the speed increasing ratio becomes large and the efficiency also deteriorates. By the way, the large speed increasing ratio of the second speed of the FRONT GEAR (front transmission mechanism) is 0.45, and the large speed increasing ratio of the second speed used in 11AT, 14AT, 15AT of "1-4-TYPE" (0.57). ) Is 1.25 times. Therefore, the gear efficiency is extremely deteriorated in the third forward speed and the fifth forward speed using the second large speed increasing ratio. Despite these major drawbacks, Aisin Seiki, AW, and Toyota have applied for many patents including GM. Then, regarding the Japanese Patent Application Laid-Open No. 2005-105721 to 105726 (AW) for which the most structural patents have been applied among them, a reference example when applied to the present application is shown in FIG. 12, and its structural diagram is shown in FIG.

<5-TYPE> “C4-1-1, 10AT” FIG. 11
FIG. 11 shows an original patent of “5-TYPE”, US20090054196 (GM) 10AT, as a schematic diagram of FR specifications to which the structure of the present application is applied, and a speed diagram showing a shift mode and engagement at each shift stage. It shows an element (SHIFT) and a gear ratio (RATIO). The operating position of the third clutch (C3) of the present application is different from that of US20090054196 (GM) 10AT, but the capacity of the present application is smaller, and this method is used throughout the present application, and the gear shifting action is the same. is there. In the speed diagram of FIG. 11, the speed diagram is divided into MAIN GEAR (main speed change mechanism) and FRONT GEAR (front speed change mechanism), and the speed map of MAIN GEAR (main speed change mechanism) is from the right of the figure. The 1st, 2nd, 3rd, and 4th constituent elements are arranged in order, and the speed diagram of the FRONT GEAR (previous transmission mechanism) shows the 5th, 6th, 7th, 8th constituent elements (A, B, C , D) are arranged, and the seventh constituent element (C) has the first constituent element via the first coupling member (7) and the first clutch (C1), the second coupling member (8) and the fourth clutch (C). C4) is connected to the fourth component, the eighth component (D) is connected to the first component via the first connecting member (7) and the second clutch (C2), and the third component is connected. Is the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  In the schematic diagram of the FR specifications of FIG. 11, a prime mover is located on the left front side (not shown), and power is input to the input shaft of the transmission via the torque converter. A fourth planetary gear train (40), a third planetary gear train, a second planetary gear train (20), and a first planetary gear train (10), which are simple planetary gears and are axially arranged from the left front of the transmission, are arranged. The first and second planetary gear trains (10, 20) form a MAIN GEAR (main transmission mechanism), and the third and fourth planetary gear trains (30, 40) form a FRONT GEAR (previous transmission mechanism). . The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the main transmission mechanism side (MAIN GEAR) of the cylindrical member provided at one end of the transmission case on the front transmission mechanism side (FRONT GEAR), and the transmission case The input shaft is rotatably supported by the output shaft which is rotatably supported at the other end thereof, and the first and second clutches (C1, C2) of the pre-shift mechanism which are connectable to the first connecting member (7) are friction members. Are arranged on the outer side in the outer peripheral direction of the cylindrical member behind the third planetary gear train so as to be stacked in two stages in the radial direction. This is the arrangement according to "claim 1" of the present application. Although a clear effect cannot be seen in the schematic diagram, the effect will be described with reference to FIG. 19 described later.

  The fifth component (A) of the FRONT GEAR (front gearbox) is composed of the third planetary gear train (30) and the third and fourth sun gears (S3, S4) of the fourth planetary gear train (40) connected to each other. , The sixth component (B) consists of the third planet carrier (P3) of the third planetary gear train (30), and the seventh component (C) is the connected third planetary gear train (30) and fourth planetary gear (30). It consists of the third ring gear (R3) of the gear train (40) and the fourth planet carrier (P4), and the eighth component (D) consists of the fourth ring gear (R4) of the fourth planet gear train (40). . Although the speed change mode has been described in the original patent and detailed description is omitted, in the velocity diagram, the sixth component (B) of the FRONT GEAR (previous transmission mechanism) is connected to the input shaft and the fifth component By braking (A) with the first brake (B1), the seventh constituent element (C) is accelerated, and further the eighth constituent element (D) is greatly accelerated, and the first and second clutches ( C1 and C2) are selectively transmitted to the first component of the MAIN GEAR (main transmission). Also, by engaging both the first and second clutches (C1, C2), the planetary gear train of the MAIN GEAR (main transmission mechanism) becomes integral and the rotation of the input shaft is selectively transmitted to the first component. . By engaging the third brake (B3) of the MAIN GEAR (main transmission mechanism), the three kinds of rotations of the first component are further decelerated, and by engaging the third clutch (C3). The three types of rotations are further reduced in speed, and the six types of reduced rotations are output from the third component. In addition, by engaging the two engaging elements of the third brake (B3) and the third clutch (C3) of the MAIN GEAR (main transmission mechanism), one kind of decelerated rotation can be output from the third constituent element. Seven types of decelerated rotations can be output. The shift mode up to this point is the same as that of "1 to 4-TYPE". By making it possible to connect the seventh constituent element (C) of the FRONT GEAR (front transmission mechanism) and the fourth constituent element of the MAIN GEAR (main transmission mechanism) by the fourth clutch (C4) in the speed increasing stage, Accelerated rotation is input to the fourth component, and three types of accelerated rotation can be output from the third component by engaging the third clutch (C3) and engaging each of the first and second clutches (C1, C2). Becomes As for the reverse gear, the fourth component of the MAIN GEAR (main transmission) is engaged by the third brake (B3), and the seventh component (C of the FRONT GEAR) is engaged by engaging the fourth clutch (C4). ) Is braked to reverse the eighth component (D) to be transmitted to the first component of the MAIN GEAR (main transmission mechanism), further decelerated, and output from the third component. As a result, 10 AT for forward speed and 1 for reverse speed are created.

  The gear ratios shown in the table of FIG. 11 are the gear ratios described in US20090054196 (GM) 10AT. The range of the gear ratio (RANGE) is too small for 7.21 and 10AT, and the step value from the third forward speed to the ninth forward speed is a geometric progression of 1.2, which is not good for a vehicle transmission. This is due to the shift mode having two transformation points, and even if the number of constituent parts is reduced and the number of shift stages is increased, the performance cannot be obtained unless the performance is improved. Originally, the two-story planetary gear of the FRONT GEAR (previous transmission mechanism), which has the most efficient meshing efficiency of the planetary gears used in the "1-TYPE" 15AT of FIG. 3 and the "4-TYPE" 14AT of FIG. If a row is used, "5-TYPE" is also made compact and the meshing efficiency of the planetary gears is slightly improved, but it cannot be used because the gear ratio is deteriorated. In FIG. 2 of JP-A-2015-183853 (GM), a two-story planetary gear train is used for the FRONT GEAR (front transmission mechanism), but the meshing efficiency of the planetary gears is poor and the connection is difficult to adopt. .

<5-TYPE> “C4-1-2, 10AT” FIGS. 12 and 19
FIG. 12 shows two planetary gear trains composed of four constituent elements of FRONT GEAR (previous transmission mechanism) with a number of “5-TYPE”, and JP-A-2015-105721 to 2015-105726 (AW ) And the most commonly used Ravigneaux planetary gear train, the structure is compared using a Ravigneaux planetary gear train with the same gear ratio for comparison with the applied construction patent. However, the meshing efficiency (GEAR EFF) of the planetary gears was calculated and described for comparison with "1-4-TYPE". The velocity diagram is the same as in FIG. 11, except that the fifth, sixth, seventh and eighth components (A, B, C, D) of the FRONT GEAR (front gearbox) replace the Ravigneaux planetary gear train. The fifth component (A) of the third and fourth planetary gear trains (30, 40), which is the Ravigneaux planetary gear train of FRONT GEAR (previous transmission mechanism), is the third sun gear (30) of the third planetary gear train (30). S3), the sixth component (B) is a common planet carrier (P) of the third and fourth planetary gear trains (30, 40), and the seventh component (C) is the third and fourth. It consists of a common ring gear (R) of the planetary gear train (30, 40) and the eighth component (D) consists of the fourth sun gear (S4) of the fourth planetary gear train (40). It is a little compact because it uses the Ravigneaux planetary gear train, but it is not as compact as a two-story building.

  In the two types of schematic diagrams in FIG. 12, the left diagram is an FR specification gear train, and the right diagram is an FF specification gear train. The schematic diagram of the FR specification in the left diagram is the same as the schematic diagram of the FR specification in FIG. 11 except that the third and fourth planetary gear trains (30, 40) are changed. In the schematic diagram of the FF specification in the right diagram, the first planetary gear train (10), the second planetary gear train (20), which are simple planetary gears, and the Ravigneaux planetary gear train, which are simple planetary gears, are arranged axially in order from the left front of the transmission. , A fourth planetary gear train (30, 40) is arranged. A cylindrical member is extended from the rear end of the transmission case to the inner periphery, and the input shaft is supported by the inner periphery of the tip of the extended cylindrical member. A Ravigneaux planetary gear train and first and second clutches (C1, C2) are arranged from the rear on the outer circumference of the rolled cylindrical member, and a fourth clutch (C4) is arranged on the outer circumference or rear of the first clutch (C1). Are arranged. Here, hydraulic oil is supplied to the first, second, and fourth clutches (C1, C2, C4) from the outer circumference of the cylindrical member. This is the arrangement according to the "claim 1" of the present application, and is a somewhat simple and compact arrangement.

  The gear ratios shown in the table of FIG. 12 are the same as the gear ratios described in JP-A-2015-105721 to 2015-105726 (AW), although there are slight differences in calculation. Here, the gear ratio of the ring gear and the sun gear of the first planetary gear train (10) is 3.610, and the gear ratio of the ring gear and the sun gear of the second planetary gear train (20) is 4.098, which is large. . Normally, when the gear ratio is large, the meshing efficiency of the planetary gears is deteriorated, so that such a gear ratio is not taken in a vehicle transmission. The “1,4-TYPE” proposed by the applicant of the present application has a maximum tooth number ratio of 3.000. This is because the speed is greatly increased by the FRONT GEAR (frontward transmission mechanism), and this large speed increase also causes deterioration of the meshing efficiency of the planetary gears. Therefore, the meshing efficiency (GEAR EFF) of the planetary gears at the third forward speed, which is greatly increased by the FRONT GEAR (front transmission mechanism) and greatly reduced by the MAIN GEAR (main transmission mechanism), is 97.2%, and the fifth forward speed is The stairs are extremely bad at 97.8%. The other gears are fairly good, but the reason why the gear ratio is large and the starting gear is bad is unavoidable, and it is possible to cover it due to slipping and starting, but there are three forward gears and five gears. Bad things affect fuel economy. The range of the gear ratio (RANGE) is 8.67, which is better than 7.21 in Fig. 11, but it is small as 10AT, and the step value from the third forward speed to the ninth forward speed is also a geometric series of 1.25 and the vehicle. Not a good transmission for cars.

  FIG. 19 “C4-1-2, 10AT (FR)” is a structural diagram conceptually designed for a passenger vehicle based on the FR specification schematic diagram of the left diagram of FIG. 12 with the input power from the prime mover set to 300 Nm. In FIG. 19, a prime mover (not shown) is arranged on the left front side of the transmission, and power is input to the transmission via the torque converter (200a). The transmission case 1 is arranged integrally, and is arranged inside the transmission case 1 from the left front to the third planetary gear train (30), the fourth planetary gear train (40) and the outer periphery thereof as a Ravigneaux planetary gear train. The front shift mechanism including the first brake (B1), the first and second clutches (C1, C2) and the fourth clutch (C4), and the fourth clutch (C4) disposed on the outer circumference. A main transmission mechanism including three brakes (B3), a partition wall, a second planetary gear train (20), a first planetary gear train (10), and a third clutch (C3) is sequentially arranged.

  A holding member 2a for holding a charging pump for hydraulically controlling the transmission is fastened to the front portion of the transmission case 1 with bolts, and a gear shift for fixing the wheel stator of the torque converter (200a) to the holding member 2a. The cylindrical member 2b, which is cylindrically extended in the machine inward direction, is fastened with a bolt. Bushings 4a and needle roller roller bearings 4b are arranged on the inner circumferences of both ends of the cylindrical member 2b to support the input shaft 3a. An output shaft 3c rotatably supported by a needle roller roller bearing 4f and a deep groove ball bearing 4e is arranged at the rear portion of the transmission case 1, and an input shaft 3a is arranged by a needle roller roller bearing 4d arranged on the inner circumference of the output shaft 3c. To support. Here, since the integrated input shaft 3a is pivotally supported at three points, the distance between the respective pivots is short, and the input shaft 3a is less susceptible to vibration due to unbalance in the circumferential direction, so that the torque transmission capacity is reduced. The diameter of the portion arranged around the input shaft 3a can be reduced, which leads to the weight reduction of the transmission.

  On the outer circumference of the cylindrical member 2b, from the left front, a third planetary gear train (30), a fourth planetary gear train (40), a first and a second clutch (as a RAVINOG planetary gear train of FRONT GEAR). C1, C2) are arranged. In the Ravigneaux planetary gear train, a long pinion gear and a short pinion gear that are axially supported by a common planet carrier (P) are meshed with each other, and the long pinion gear is the third sun gear (S3) of the third planetary gear train (30). The common ring gear (R) meshes with the short pinion gear, and the fourth sun gear (S4) of the fourth planetary gear train (40) meshes therewith. The third sun gear (S3) of the third planetary gear train (30) is welded to the brake hub of the first brake (B1) in front of the third sun gear (S3) and is extended to the outer periphery of the third planetary gear train (30) to form the first brake (S3). The friction member B1) is locked and is axially supported on the outer circumference of the cylindrical member 2b by the bush 4i press-fitted on the inner circumference. The common planet carrier (P), which is extended from the central portion in the axial direction of the long pinion gear to the inner peripheral side, is welded to the input connecting member (Y) axially supported by the bush 4c on the outer periphery of the cylindrical member 2b, and the fourth planet The fourth sun gear (S4) of the gear train (40) has the clutch hub of the second clutch (C2) welded to the rear part thereof and is axially supported by the bush 4k on the outer periphery of the input coupling member (Y). A second connecting member (8) serving as a clutch hub of the first clutch (C1) is spline-connected to the outer circumference of the common ring gear (R).

  At the outer peripheral end of the cylindrical member 2b behind the Ravigneaux planetary gear train, the first and second planetary gear trains (10, 10, which are the first constituent elements of the FRONT GEAR (front gearbox) to the MAIN GEAR (main gearbox) are provided. 20) The first and second clutches (C1, C2) connected to the first and second sun gears (S1, S2) are arranged. In the first and second clutches (C1, C2), a dual clutch connecting member (X) in which friction members are coaxially superposed in two steps in the radial direction is axially supported by a bush 4j on the outer periphery of the input connecting member (Y). To be done. The input connecting member (Y) is spline-connected to the input shaft 3a behind the cylindrical member 2b, the outer periphery of the cylindrical member 2b is extended to the center of the Ravigneaux planetary gear train, and the outer periphery of the cylindrical member 2b is rotatably supported by the bush 4c. Is pivotally supported. The double clutch connecting member (X) is integrally formed with a first connecting member (7) for connecting to a rear MAIN GEAR (main transmission mechanism), and the first connecting member (7) is provided on the input shaft 3a with a needle roller roller. It is rotatably supported by the bearing 4n. Further, a clutch cover is welded to the rear of the dual clutch connecting member (X), and a clutch hub that serves as a second connecting member (8) spline-connected to the ring gear (R) is locked to the front of the outer periphery. A servo mechanism for pressing the friction member of the first clutch (C1) is arranged, and the second clutch, which is locked to the clutch hub welded to the fourth sun gear (S4), is located inside the front of the double clutch connecting member (X). A servo mechanism for pressing the friction member (C2) is provided. Between the dual clutch connecting member (X) and the clutch cover, a partition plate forms a passage for the working oil of the piston and the hydraulic cancel oil supplied to the servo mechanism of the two clutch (C2). In the servo mechanism of the first and second clutches (C1, C2), the working oil and the hydraulic cancel oil are provided from the outer circumference of the cylindrical member 2b through an oil passage sealed by a piston ring provided in the input coupling member (Y). Is supplied. The structure of the first and second clutches (C1, C2) is the "claim 2" of the present application, which is extremely simple and compact, and the oil passages to the first and second clutches (C1, C2) are also line resistance. Is small and the responsiveness of the clutch is improved.

  The first brake (B1) arranged in front of the transmission case 1 enables the third sun gear (S3) of the third planetary gear train (30) to be braked, and the first brake (B1) is arranged in front of the third sun gear (S3). The brake hub of the brake (B1) is welded and extended on the outer periphery of the long pinion gear to lock one friction member, and the other friction member is locked to a spline formed in front of the transmission case 1. It The piston and the return spring are held in the outer peripheral hydraulic chamber of the holding member 2a fixed to the front of the transmission case 1 with a bolt, and the hydraulic servo of the first brake (B1) is formed.

  Behind the first and second clutches (C1, C2) is a common ring gear (R) of the Ravigneaux planetary gear train and a second planetary gear train (20) which is the fourth component of the MAIN GEAR (main transmission mechanism). A fourth clutch (C4) that connects the second ring gear (R2) is arranged. The second connecting member (8) spline-connected to the common ring gear (R) is formed with a spline that locks one friction member of the fourth clutch (C4), and the first and second clutches (C1, It is extended to the rear of C2) and is axially supported by bushes 4l and 4m on the outer periphery of a first connecting member (7) integrally formed with the dual clutch connecting member (X). A clutch cover of the fourth clutch (C4) is arranged behind the second connecting member (8) to lock the other friction member of the fourth clutch (C4) and to arrange a servo mechanism for pressing the friction portion. To be done. Behind the clutch cover of the fourth clutch (C4), a retaining ring is fixed to the transmission case 1 and a partition wall extended to the inner periphery is arranged, and a seal ring is sealed from an oil passage provided in the partition wall. Oil is supplied to the hydraulic servo of the fourth clutch (C4) via the oil passage. Further, the plate spline-connected to the second ring gear (R2) of the second planetary gear train (20) is extended to the inner peripheral side of the partition wall, and is axially supported by the bush 4o on the outer periphery of the first connecting member (7). At the same time, the inner peripheral side of the partition wall is spline-connected to the clutch cover of the fourth clutch (C4).

  A second planetary gear train (20) is arranged behind the partition wall. Since the second planetary gear train (20) has a large gear ratio of 4.098 between the ring gear and the sun gear, the second planetary gear train (20) is arranged up to the inner circumference of the transmission case 1. The third brake (B3) for braking the second ring gear (R2) of the second planetary gear train (20) is arranged on the outer periphery of the clutch cover of the fourth clutch (C4) and is connected to the second ring gear (R2). One friction member of the third brake (B3) is locked to the clutch cover, and the other friction member is locked by the spline formed on the transmission case 1. The piston and the return spring are held in the outer peripheral hydraulic chamber of the partition wall to form the hydraulic servo of the third brake (B3). First and second sun gears (S1, S2), which are integrated with the first and second planetary gear trains (10, 20), are spline-connected to the first connecting member (7) on the inner peripheral side of the partition wall. The structures of the first and second planetary gear trains (10, 20), the third clutch (C3), and the output shaft 3c are shown in Fig. 14 "1-TYPE" 9AT, Fig. 16 "2-TYPE" 9AT, 17 "3-TYPE. The description is omitted because it is the same as 11AT.

  Comparing FIG. 19 of the present application with JP-A-2015-105721 to 105726 (AW), JP-A-2015-105721 to 105726 (AW) has a larger waist circumference. This is because the clutch operating oil passage is provided on the input shaft, and the input shaft has to be doubled and made thicker, which leads to an increase in weight. Further, regarding the oil passage between the transmission case and the input shaft, the oil passage to the first and second clutches (C1, C2) must be sealed with a sheet ring, which increases the axial length. Naturally, the line resistance also increases, which is disadvantageous for control. Further, since this transmission mainly uses the increased speed rotation of the input shaft, the internal parts rotate at a higher speed than a normal transmission. In FIG. 19 of the present application, half of the high-speed rotating parts are pivotally supported by the cylindrical member 2b that is integrated with the transmission case 1, whereas in JP-A-2015-105721 to 105726 (AW), all of the high-speed rotating parts are input shafts. It is supported by. Therefore, it is susceptible to vibration. Of course, the present application is simpler and less expensive.

  Comparing FIG. 19 and JP-A-2005-105721 to 105726 (AW), there is a merit in FIG. 19 of the present application, but as FIG. 19 “5-TYPE” 10AT with the same input torque of 300 Nm, FIG. TYPE ”9AT, FIG. 15“ 1-TYPE ”15AT, FIG. 16“ 2-TYPE ”9AT, and FIG. 17“ 3-TYPE ”11AT. Simply comparing the axial lengths of the transmission units, based on FIG. 14 “1-TYPE” 9AT, FIG. 16 “2-TYPE” 9AT is 2%, FIG. 15 “1-TYPE” 15AT is 6%, and FIG. "3-TYPE" 11AT becomes 20% longer, and "5-TYPE" 10AT in FIG. 19 becomes 28% longer. In the "5-TYPE" 10AT, the arrangement of the fourth clutch (C4) lengthens the axial direction by 28%, and in the "3-TYPE" 11AT in Fig. 17, the two planets that make up the FRONT GEAR (previous gear shift mechanism). The connection of the gear train and the five fastening elements makes the axial direction as long as 20%. FIG. 15 "1-TYPE" 15AT, which has a wider transmission ratio range (RANGE) than those of 10AT and 11AT, has three planetary gear trains and five planetary gear trains that form a FRONT GEAR (previous transmission mechanism). There are many fastening elements, but the connection is compact and conversely 15-20% shorter.

<6-TYPE>
The multi-stage transmission described in paragraph "0018" has the same speed change mode as the C type 7AT according to Daimler's Japanese Patent Laid-Open No. 2000-266138, and one type other than two types of input shaft deceleration rotation for FRONT GEAR Since the fourth brake (B4) is arranged in the second component of the MAIN GEAR (main transmission mechanism), it can be applied only to the FF. The front transmission is composed of two planetary gear trains of four constituent elements and four engagement elements including first and second clutches (C1, C2), and is composed of four constituent elements. There are a plurality of combinations of two planetary gear trains, and there are preceding JP2012-225506 (HUNDAI), JP2015-161312 (Honda), and JP2015-161312 (Honda). Since the arrays are parallel, it is difficult to achieve FF specifications. The present application describes FIGS. 13 and 24 as an embodiment in which a combination of two planetary gear trains is a two-story structure, and the output gears are arranged in a characteristic manner to be compact.

<6-TYPE> “C5-1-1, 10AT” FIGS. 13 and 24
FIG. 13 is a schematic diagram of FF specifications of C5 type 10AT, a speed diagram showing a shift form, a fastening element (SHIFT) at each shift stage, and a gear ratio (RATIO) and a meshing efficiency (GEAR EFF) of a planetary gear. It is shown. In the velocity diagram of FIG. 13, the velocity diagram is divided into MAIN GEAR (main transmission mechanism) and FRONT GEAR (front transmission mechanism). The speed diagram of the MAIN GEAR (main transmission mechanism) is such that the first, second, third, and fourth constituent elements are arranged in order from the right side of the figure, and the speed diagram of the sub-front transmission mechanism of the FRONT GEAR (front transmission mechanism). 5, 5th, 6th, 7th, 8th constituent elements (A, B, C, D) are arranged in order from the right side of the drawing, and the 6th constituent element (B) and the 1st constituent element are the first connecting member (7). ) And the third component is the output of the transmission. The vertical direction of the velocity diagram represents the velocity, the value entered as 1 indicates the rotation speed of the input shaft, and the value entered as 0 indicates zero speed.

  In the schematic view of the FF specification in FIG. 13, a prime mover is located on the left front side (not shown), and power is input to the input shaft of the transmission via the torque converter. A first planetary gear train (10), a second planetary gear train (20), a two-story third planetary gear train (30), and a fourth planetary gear train, which are simple planetary gears arranged axially from the left front of the transmission. (40) is arranged, the first and second planetary gear trains (10, 20) form a MAIN GEAR (main transmission mechanism), and the two-story third and fourth planetary gear trains (30, 40) are arranged. It constitutes a FRONT GEAR. An output counter gear is provided between the first planetary gear train (10) and the second planetary gear train (20), which is pivotally supported by a partition wall integrated with the transmission case. The first, second, third and fourth planetary gear trains (10, 20, 30, 40) include first, second, third and fourth sun gears (S1, S2, S3, S4) and a first , Second, third and fourth planetary carriers (P1, P2, P3, P4) and first, second, third and fourth ring gears (R1, R2, R3, R4). In addition, the input shaft is axially supported by the inner peripheral end of the cylindrical member provided at the rear of the two-story third and fourth planetary gear trains (30, 40) behind the transmission case, and The other end in front supports the input shaft. The first and second clutches (C1, C2) of the front transmission mechanism are arranged outside the cylindrical member in the outer peripheral direction so that the friction members are radially stacked in two stages.

In the schematic diagram and the velocity diagram of FIG. 13, the first and second planetary gear trains (10, 20) forming the MAIN GEAR (main transmission mechanism) are the same as those described in “1-5-TYPE”. Therefore, the description is omitted.
In the two-story third and fourth planetary gear trains (30, 40) forming the FRONT GEAR (previous transmission), the fourth planetary gear train (40) is on the first floor and the third planetary gear train (30) is Located on the 2nd floor. The third sun gear (S3) of the third planetary gear train (30) and the fourth ring gear (R4) of the fourth planetary gear train (40) are integrated into a fifth component (A), and the third planetary gear is formed. The third planetary carrier (P3) of the train (30) and the fourth planetary carrier (P4) of the fourth planetary gear train (40) are connected to form the sixth component (B), and the third planetary gear train (30). And the fourth sun gear (S4) of the fourth planetary gear train (40) becomes the eighth component (D).

  Here, the first and second sun gears (S1, S2) forming the first constituent element of the MAIN GEAR (main transmission mechanism) are connected and also connected to the first connecting member (7) to form the second constituent element. The first planetary carrier (P1) that constitutes the second planetary carrier (P1) that can be connected to the input shaft by the third clutch (C3) and can be braked by the fourth brake (B4), and that constitutes the third constituent element. The first ring gear (R1) connected to (P2) is connected to the output shaft, and the second ring gear (R2) constituting the fourth component can be braked by the third brake (B3). The third sun gear (S3) and the fourth ring gear (R4), which form the fifth component (A) of the FRONT GEAR (frontward transmission), can be connected to the input shaft by the first clutch (C1). And the connected third and fourth planetary carriers (P3, P4) constituting the sixth component (B) are coupled to the first coupling member (7) and constitute the seventh component (C). The third ring gear (R3) can be connected to the input shaft by the second clutch (C2) and can be braked by the first brake (B1), and the fourth sun gear (S4) that constitutes the eighth component (D). ) Can be braked by the second brake (B2). As described in "1-5-TYPE", the second planetary carrier (P1) of the MAIN GEAR (main transmission mechanism) can be connected to the input shaft by the third clutch (C3). It is advantageous to be able to connect the carrier (P2) and the first ring gear (R1) with the third clutch (C3) because the capacity of the third clutch (C3) can be small, but in "6-TYPE". Since the first planetary carrier (P1) must be able to be braked by the fourth brake (B4), it is preferable that the same first planetary carrier (P1) can be connected to the input shaft by the third clutch (C3). Becomes compact.

  The table of FIG. 13 shows the engagement element (SHIFT) and the gear ratio (RATIO) at each shift speed, and the meshing efficiency (GEAR EFF) of the planetary gears. Regarding the gear ratio (RATIO) showing the performance, the gear ratio step value (STEP), and the gear ratio width (RANGE), the gear ratio width (RANGE) is suitable for 10.933 and 10AT of 7,833 to 0.714. As a result, an appropriate speed change step can be achieved with 7 forward deceleration speeds and 3 speed up speeds, and the step value (STEP) of the gear ratio becomes almost appropriate. Regarding the meshing efficiency (GEAR EFF) of the planetary gears, the fourth forward speed (4th) is 97.3%, which is extremely poor as compared with the other gears, which is a major performance drawback. This is because the meshing efficiency of the planetary gears becomes poor in the speed change of FRONT GEAR (front speed change mechanism), but there are two positions of the third forward speed (3rd) and the fifth forward speed (5th) like "5-TYPE". Since the middle speed stage is not bad and the performance is good except at the forward fourth speed (4th), the structure was examined in FIG. 24 for practical use. A description of the shift mode such as the flow of power in each shift stage is omitted because it has already been disclosed in Japanese Patent Laid-Open No. 2012-225506 (HUNDAI).

  FIG. 24 “C5-1-1, 10AT (FF)” is a structural diagram conceptually designing the schematic view of the FF specification of FIG. 13 with the input power from the prime mover set to 300 Nm, and the total length is 9AT (FF) of FIG. The same is about 400 mm. The input structure from the torque converter 200a and the output structure for outputting to the differential device 9 via the counter gear 6 meshing with the output counter gear 5 are the same as those in FIG. The housing that houses the entire transmission is composed of a front torque converter case 1a, a transmission case 1b that serves as the transmission main body, and a transmission case 1c that closes the rear portion. The transmission case 1b has an L-shaped central portion in the axial direction. A mold-shaped cylindrical partition is integrally provided, and an angular bearing 4e is fixed to the outer periphery of the inner peripheral cylindrical portion by screws with the back surfaces aligned, and axially supports the output counter gear 5. The first planetary gear train (10) and the third clutch (C3) are arranged in front of the partition wall from the partition wall side, and behind the partition wall, the output counter gear 5 and the second planetary gear train ( 20), two-story third and fourth planetary gear trains (30, 40), and first and second clutches (C1, C2). A first brake (B1) is arranged on the outer periphery of the first and second clutches (C1, C2), and a second brake (B1) is arranged on the outer periphery of the two-story third and fourth planetary gear trains (30, 40). B2) is arranged to constitute a front speed change mechanism, a third brake (B3) is arranged on the outer periphery of the second planetary gear train (20), and an outer periphery of the first planetary gear train (10) is arranged in front of the partition wall. A fourth brake (B4) is arranged and constitutes a main transmission mechanism.

  The transmission case 1c for closing the rear part of the transmission has an integral inner peripheral portion protruding forward in a cylindrical shape, and a needle roller roller bearing 4b is arranged at the inner peripheral end portion of the protruding cylindrical member to provide the input shaft 3 To support. First and second clutches (C1, C2) are arranged on the outer circumference of the protruding cylindrical member. The first clutch (C1) can connect a third sun gear (S3) integrated with the input shaft 3 and a fourth ring gear (R4), and the second clutch (C2) can connect the input shaft 3 and the third ring gear (R3). R3) can be connected. In the first and second clutches (C1 and C2), a dual clutch connecting member (X) in which friction members are coaxially overlapped in two radial steps is formed integrally with the input connecting member (Y), The member (Y) is spline-connected to the input shaft 3 in front of the cylindrical member protruding to the inner circumference of the transmission case 1c, and is extended rearward along the outer circumference of the cylindrical member. A clutch cover is welded to the rear end of the input coupling member (Y), and a second clutch (C2) is arranged on a clutch hub that couples the radial outer circumference of the dual clutch coupling member (X) and the third ring gear. A servo mechanism that presses the friction member is arranged, and in front of the dual clutch connecting member (X), a radially inner circumference of the dual clutch connecting member (X), an integrated third sun gear (S3), and a fourth sun gear (S3). A servo mechanism for pressing the friction member of the first clutch (C1) arranged on the clutch hub connected to the ring gear (R4) is arranged. Between the dual clutch connecting member (X) and the clutch cover, a partition plate forms a passage for piston working oil and hydraulic cancel oil supplied to a servo mechanism arranged on the radial upper side of the clutch cover. The servo mechanism of the second clutch (C2) arranged on the upper side in the radial direction of the dual clutch connecting member (X) and the servo mechanism of the first clutch (C1) arranged on the lower side in the radial direction include the transmission case 1c. The working oil and the hydraulic cancel oil are supplied from the outer circumference of the cylindrical member protruding to the inner circumference through an oil passage sealed by a piston ring provided in the input coupling member (Y).

  Two-story third and fourth planetary gear trains (30, 40) are arranged in front of the first and second clutches (C1, C2). In the fourth planetary gear train (P4) of the fourth planetary gear train (40) arranged on the first floor of the two-story building, the right side member is pivotally supported by the bush on the outer periphery of the input shaft 3 and the fourth sun gear (S4). The first connecting member (7) extends through the inner circumference to the second planetary gear train (20) in front and is spline-connected to the second sun gear (S2), and the left side member is arranged on the second floor. It is formed integrally with the left side member of the third planetary carrier (P3) of the three planetary gear train (30). The fourth sun gear (S4) is axially supported by a bush on the outer periphery of the right side member of the third planet carrier (P3) extended to the second planetary gear train (20), and the second brake (B2) is forwardly mounted. A thin plate-shaped brake hub that locks the friction member is welded. The outer periphery of the fourth ring gear (R4) is integrated with the third sun gear (S3) of the third planetary gear train (30), and the plate splined at the rear is extended to the inner periphery to extend the outer periphery of the input shaft 3. A clutch hub that is axially supported by the bush 4x and that locks the friction member of the first clutch (C1) is spline-connected. The third ring gear (R3) of the third planetary gear train (30) arranged on the second floor of the two-story is internally supported by the bush 4y on the outer periphery of the clutch hub of the first clutch (C1). In addition, a plate to which a clutch hub for locking the friction member of the second clutch (C2) is welded is spline-connected to the side surface.

  The first brake (B1) arranged on the outer periphery of the first and second clutches (C1, C2) enables braking of the third ring gear (R3) of the third planetary gear train (30). The clutch hub of the second clutch (C2) welded to the plate that is spline-connected to the third ring gear (R3) locks one friction member of the second clutch (C2) at the inner circumference and the first at the outer circumference. One friction member of the brake (B1) is locked. The other friction member of the first brake (B1) is locked to a spline formed on the inner circumference of the transmission case 1b, and a piston and a return spring are held in a hydraulic chamber formed in the transmission case 1c. A hydraulic servo for one brake (B1) is formed.

  The second brake (B2) arranged on the outer periphery of the two-story third and fourth planetary gear trains (30, 40) is capable of braking the fourth sun gear (S4) of the fourth planetary gear train (40). To do. The brake hub welded to the front of the fourth sun gear (S4) extends to the outer periphery of the third planetary gear train (30) and locks one friction member of the second brake (B2). The other friction member of the second brake (B2) is locked to the extended spline of the one brake (B1) formed on the inner circumference of the transmission case 1b. A brake case that holds a piston forming a hydraulic servo of the second brake (B2) and a return spring is fixed to the transmission case 1b in front of the friction member of the second brake (B2) with a retaining ring.

  The second sun gear (S2) of the second planetary gear train (20) arranged in front of the two-story third and fourth planetary gear trains (30, 40) is the right side of the fourth planetary carrier (P4). The first connecting member (7) integrated with the member is spline-connected on the inner circumference. The planet pinion gear that meshes with the second sun gear (S2) is supported by the second planet carrier (P2), and the left side member is spline-connected to the output counter gear 5 and fixed by the retaining ring. The second ring gear (R2) that meshes with the planet pinion gear has a plate in which the right end tooth portion is axially fixed by a retaining ring and is spline-connected to the fourth sun gear (S4) of the fourth planetary gear train (40). Between the second sun gear (S2) and the brake hub welded to the fourth sun gear (S4) of the fourth planetary gear train (40), the axial direction is regulated by the thrust needle bearing and the rotatable hub is rotatably arranged. .

  The third brake (B3) enables braking of the second ring gear (R2) of the second planetary gear train (20). The second ring gear (R2) has a spline formed on the outer peripheral portion thereof to lock one friction member of the third brake (B3). A spline is formed on the inner circumference of the transmission case 1b to lock the other friction member of the third brake (B3). Further, behind the friction member of the third brake (B3), there is a brake case which is integrated with the brake case of the second brake (B2) and forms a hydraulic servo of the third brake (B3) and a return spring. Will be distributed.

  The output counter gear 5 arranged in front of the second planetary gear train (20) is fixed to the outer periphery of the cylindrical portion of the L-shaped cylindrical partition wall arranged at the axial center of the transmission case 1b by screws. It is pivotally supported by the back-aligned angular bearing 4e. The left side member of the second planetary carrier (P2) of the second planetary gear train (20) is spline-connected to the right outer periphery of the output counter gear 5 and fixed by a retaining ring.

  In the first planetary gear train (10) arranged in front of the partition wall arranged in the axial center of the transmission case 1b, the first sun gear (S1) is pivotally supported on the outer circumference of the input shaft 3 by the second bush. The material is extended to the planetary gear train (20) and spline-connected to the inner circumference of the second sun gear (S2). The left side member of the first planetary carrier (P1) is extended to the outer periphery of the first ring gear (R1), and one friction member of the third clutch (C3) is locked in front of the spline formed on the outer periphery. At the same time, one of the friction members of the fourth brake (B4) is locked at the rear side, and the right side member is pivotally supported by the extended material portion of the first sun gear (S1) with a bush. The first ring gear (R1) is axially supported by a bush at the inner peripheral portion of the partition wall by the extension of the first sun gear (S1), and is on the left side of the second planet carrier (P2) of the second planetary gear train (20). The side member is spline-connected to the connecting member.

  A third clutch (C3) arranged in front of the first planetary gear train (10) enables the input shaft 3 and the first planetary carrier (P1) of the first planetary gear train (10) to be connected. The clutch cover of the third clutch (C3) welded to the input shaft 3 includes the outer periphery of the cylindrical portion protruding rearward from the inner periphery of the holding members 2a and 2b that are bolted to the transmission cases 1a and 1b, and the holding member 2a. 2b is extended along the side surface to the outer periphery of the first planetary gear train (10) to lock the other friction member of the third clutch (C3), and holds a piston and a return spring forming a hydraulic servo. . Hydraulic oil is supplied to the hydraulic servo of the third clutch (C3) from the outer circumferences of the cylindrical portions of the holding members 2a and 2b via an oil passage sealed with a seal ring.

  The fourth brake (B4) arranged behind the third clutch (C3) on the outer periphery of the first planetary gear train (10) can brake the first planetary carrier (P1) of the first planetary gear train (10). And The other friction member of the fourth brake (B4) is arranged in opposition to the inner periphery of the transmission case 1b behind the friction member of the third clutch (C3), and is a partition wall integrated with the transmission case 1b. A piston and a return spring that form a hydraulic servo are retained in the.

  The "6-TYPE" 10AT is a power train that can be established only in the FF specification, and although there is a drawback that the meshing efficiency of the planetary gears in the fourth forward speed (4th) is deteriorated, the efficiency and the gear ratio of other gear stages are not improved. The performance is good, and it is established by arranging the two planetary gear trains of the front transmission mechanism in a two-story structure or by arranging the two planetary gear trains of the main transmission mechanism with the output gear sandwiched therebetween, as in the present application. The possibility of doing so increases.

  The present invention combines six types of "1-6-TYPE" front transmissions with a main transmission having the same specifications to control two planetary gear trains consisting of four constituent elements with two fastening elements. The present invention can be arranged in a simpler and more compact structure than the conventional structure. Comprehensively evaluating these six types, "1-TYPE" 9AT and "2-TYPE" 9AT are gear ratio (RATIO), gear ratio step value (STEP), gear ratio width (RANGE), and planetary gear ratio. It is the best in terms of meshing efficiency (GEAR EFF), and the structure of the present invention makes it the simplest and most compact. Each of the other 10 to 11 ATs has its own drawbacks and requires careful usage. Among them, "5-TYPE" 10AT has drawbacks in both performance and structure, and there is a doubt in practical use. The 8AT currently in practical use has drawbacks in both performance and structure, and the "6-TYPE" 10AT also has drawbacks in efficiency, but since it is superior to these 8ATs, there is a possibility of practical application. As for 11AT, the range of gear ratio (RANGE) is larger than that of "1-TYPE" 9AT or "2-TYPE" 9AT, so it is worth it, but passenger cars do not need such a range (RANGE) and commercial Suitable for cars (Truck Bus). Comparing the 11ATs, it is better for the commercial vehicle to be compact in the axial direction, and "1-TYPE" 11AT is the best in terms of performance and structure, and "2-TYPE" 11AT becomes "3-TYPE" 11AT in structure. Is slightly inferior. The "1-TYPE" 15AT and "4-TYPE" 14AT can be made much more compact than the "5-TYPE" 10AT and "3-TYPE" 11AT, and can also be realized as an FF, including a prime mover for passenger cars and commercial vehicles. It is an AT that has the potential to revolutionize the future.

1, 1a, 1b, 1c Cases 2a, 2b Holding members 3, 3a Input shafts 3c Output shafts 4a-4y Bearings 7, 8 Connecting members 10, 20, 30, 40, 50 Planetary gear train 200a Torque converter
C1, C2, C3, C4 Clutch B1, B2, B3, B4 Brake S1, S2, S3, S4, S5 Sun gear P1, P2, P3, P4, P5 Planet carrier R1, R2, R3, R4, R5 Ring gear

Claims (7)

A first planetary gear train (10) having the first sun gear (S1), a first planetary carrier (P1), and a first ring gear (R1), which are simple planetary gears, and a second planetary gear train, which is a simple planetary gear. The first sun gear (S1) and the second sun gear (S2) of the second planetary gear train (20) having the components of the sun gear (S2), the second planet carrier (P2), and the second ring gear (R2). ) As a first component, and a first planetary carrier (P1) whose input shaft is connectable with a third clutch (C3) as a second component and a first ring gear (R1) and a second planet The carrier (P2) is used as a third component, or the first planet carrier (P1) connected to the input shaft is used as a second component and a first ring gear (C3) is connectable to the first ring carrier (P3). R1) and second planet key The rear (P2) is the third component, the second ring gear (R2) that can be braked by the third brake (B3) is the fourth component, and the second planet carrier (P2) is connected to the output shaft, The first constituent element of the main transmission mechanism having a common speed diagram in which the first, second, third, and fourth constituent elements are arranged side by side in the axial direction,
Includes at least two planetary gear trains and first and second clutches (C1, C2) for selectively inputting a plurality of variable speed rotations including rotation of the input shaft and at least accelerated rotation of the input shaft. Providing a front transmission with at least four fastening elements,
By selectively restricting the rotation of any two of the first, second and fourth constituent elements of the main transmission mechanism, the third constituent element achieves at least a forward speed of 9th speed and a reverse speed of 1st speed. A multi-stage transmission that is adapted to obtain
Inside the transmission case of the multi-stage transmission, the front transmission mechanism and the main transmission mechanism are arranged axially in order, and the input shaft is arranged integrally at the center of rotation of the front transmission mechanism and the main transmission mechanism, The input shaft is pivotally supported by the inner peripheral end of the cylindrical member provided on one end of the transmission case on the front transmission mechanism side on the main transmission mechanism side, and the other end of the transmission case is described above. The input shaft is pivotally supported, or the input shaft is pivotally supported by the output shaft pivotally supported at the other end of the transmission case,
The first and second clutches (C1, C2) of the front speed change mechanism are rotatably arranged on the outer side in the outer peripheral direction of the cylindrical member so that the friction members overlap in the radial direction in two steps, and the outer periphery of the cylindrical member. To supply hydraulic oil to the first and second clutches (C1, C2) from
At least one planetary gear train or at least two planetary gear trains of at least two planetary gear trains of the front transmission mechanism are arranged axially from the inner peripheral end of the cylindrical member toward the main transmission mechanism. A multi-stage transmission device in which a pair of planetary gear trains are arranged in a two-storied structure in which the planetary gear trains are radially stacked.   A cylindrical input connecting member (Y) that is connected to the input shaft is rotatably arranged on the outer peripheral direction outer side of the cylindrical member, and each friction is applied to the first and second clutches (C1, C2) of the front transmission mechanism. A dual clutch connecting member (X) that is coaxially overlapped in two stages in the radial direction is provided, and the input connecting member (Y) and the dual clutch connecting member (X) are integrally connected, or the input connecting member is connected. The second clutch connecting member (X) is rotatably arranged outside the member (Y) in the outer circumferential direction, and the first and second input clutch connecting members (Y) or the second clutch connecting member (X) are connected to each other. A hydraulic chamber that serves as a servo mechanism for two clutches (C1, C2) is provided, and hydraulic oil is passed from the outer circumference of the cylindrical member to the hydraulic chambers of the first and second clutches (C1, C2) through the input connecting member (Y). The multi-stage transmission according to claim 1, wherein the multi-stage transmission is adapted to be supplied.   A partition wall integrated with the transmission case is provided at an axial center of the transmission case, an output counter gear axially supported by the partition wall is arranged, and the output counter gear is sandwiched on one side in the axial direction. The first planetary gear train (10) of the main transmission mechanism is arranged, and the second planetary gear train (20) of the main transmission mechanism and the front transmission mechanism are arranged on the other side. Multi-speed transmission. The first planetary carrier (P1) of the main transmission mechanism, which is connected to the input shaft, is a second component, and the first planetary carrier (P1) is connectable by the third clutch (C3) and the second planetary gear (R1). In the configuration in which the carrier (P2) is the third constituent element, the third clutch (which enables the first ring gear (R1) and the second planetary carrier (P2) to be coupled to a partition wall integrated with the transmission case ( A passage for supplying hydraulic oil of C3) is provided,
Alternatively, in the configuration in which the multi-stage transmission has four brakes, the multi-stage transmission according to claim 3, wherein an oil pressure chamber serving as a servo mechanism for one brake is provided in the partition wall. The planetary carrier of one of the planetary gear trains of the front speed change mechanism having at least two planetary gear trains is fastened to the input shaft and other constituent elements by the first and second clutches (C1, C2). In the configuration of the front speed change mechanism that becomes possible,
A side member that axially supports the planetary gear of the planet carrier is connected to a double clutch connecting member (X) of the first and second clutches (C1 and C2), and the side member is provided with the first and second clutches. 2. The multi-stage transmission according to claim 1, wherein a hydraulic chamber serving as a servo mechanism for at least one of (C1, C2) is provided. The front transmission mechanism comprises a main front transmission mechanism and a sub front transmission mechanism,
In the main front speed change mechanism, the input shaft is connected to the first clutch (C1) by a component E which is a planet carrier of one simple planetary gear in which three components D, E and F are arranged in the axial direction. ), The component E and the component D or F can be connected by the second clutch (C2), the component F and the first component of the main transmission mechanism are connected, and the component E is connected. It is a first main front shift mechanism that can be braked by a second brake (B2), or a second main clutch that can brake the constituent element D of the first main front shift mechanism by a fourth brake (B4). It is a front speed change mechanism,
In the sub pre-transmission mechanism, the input shaft is connected to a constituent element C of one planetary gear train in which three constituent elements A, B and C are arranged side by side in the axial direction and the constituent element A is a first constituent element. It is a first auxiliary front speed change mechanism that can be braked by a brake (B1), or is composed of two planetary gear trains in which four constituent elements A, B, C, and G are arranged in axial order. A second auxiliary front transmission mechanism in which the input shaft is connected to the element A so that the component G can be braked by the first brake (B1) and the component C can be braked by the fourth brake (B4).
The component D of the first main front transmission mechanism and the component B of the first auxiliary front transmission mechanism are connected to each other, and the first and second clutches (C1 and C2) and the first and second brakes are connected. By connecting any two of (B1, B2), the component F of the first main front transmission mechanism is configured to rotate the input shaft, reduce the rotation of the input shaft, and increase the speed of the input shaft. Five kinds of rotations, namely, rotation, 0 rotation, and reverse rotation of the input shaft are selectively obtained, and any two of the first, second, and fourth constituent elements of the main transmission mechanism are provided. By selectively restricting the rotation of the constituent elements, the third constituent element is configured to obtain a shift speed of 9 forward speeds and 1 reverse speed,
Alternatively, the component D of the second main front transmission mechanism and the component B of the first sub front transmission mechanism are connected to each other, and the first and second clutches (C1, C2), and the first and second clutches are connected. By engaging any two of the second and fourth brakes (B1, B2, B4), the constituent element F of the second main front transmission mechanism can rotate the input shaft and reduce the rotation of the input shaft. , Six types of rotations of the input shaft, namely, two types of speed-up rotations of the input shaft, zero revolutions, and reverse rotations of the input shaft are selectively obtained. By selectively restricting the rotation of any two of the four constituent elements, the third constituent element is configured to obtain the forward 11th speed and reverse 1st speed.
Alternatively, the component D of the first main front transmission mechanism and the component B of the second auxiliary front transmission mechanism are connected to each other, and the first and second clutches (C1, C2), and the first and second clutches are connected. By engaging any two of the second and fourth brakes (B1, B2, B4), the component F of the first main pre-position transmission mechanism causes the rotation of the input shaft and the deceleration rotation 2 of the input shaft. 8 types of rotations, that is, two types of speed-up rotation of the input shaft, zero rotation, and two types of reverse rotation of the input shaft are selectively obtained. The fifth component is configured to obtain a 14th forward gear and a 1st reverse gear by selectively restricting the rotation of any two of the second and fourth components. The described multi-stage transmission. The front speed change mechanism includes a third planetary gear train (30) having components of a third sun gear (S3), a third planetary carrier (P3), and a third ring gear (R3), which are simple planetary gears. The third sun gear (S3) of the fourth planetary gear train (40) having the components of the fourth sun gear (S4), the fourth planet carrier (P4), and the fourth ring gear (R4) is connected to the input shaft. Then, the third planet carrier (P3) and the fourth ring gear (R4) can be connected by the second clutch (C2), the third ring gear (R3) and the fourth planet carrier (P4) are connected, and A planetary carrier (P3) and the input shaft can be connected by a first clutch (C1), a fourth ring gear (R4) is connected to a first component of the main transmission mechanism, and the first and second clutches are connected. (C1, C2) and the third planetary gear train (30) and the fourth A transmission mechanism for selectively changing the connection of the star gear train (40),
The third planet carrier (P3) can be braked by the first brake (B1), the fourth sun gear (S4) can be braked by the second brake (B2), and the first and second clutches (C1, C2), And a fourth ring gear (R4) connected to the first component of the main transmission mechanism by engaging any two of the first and second brakes (B1, B2) to rotate the input shaft. , A decelerated rotation of the input shaft, an accelerated rotation of the input shaft, 0 rotation, and a reverse rotation of the input shaft are selectively obtained. By selectively restricting the rotation of any two of the first, second, and fourth constituent elements, the third constituent element is configured to obtain a shift stage of 9 forward speeds and 1 reverse speed,
Alternatively, the third planet carrier (P3) can be braked by the first brake (B1), the fourth sun gear (S4) can be braked by the second brake (B2), and the third ring gear (R3) connected to the third The fourth planetary carrier (P4) can be braked by the fourth brake (B4), and the first and second clutches (C1, C2) and the first, second, and fourth brakes (B1, B2, B4) A fourth ring gear (R4) connected to the first constituent element of the main transmission mechanism by fastening any two of them is used to rotate the input shaft, two types of decelerated rotation of the input shaft, and the input shaft. 6 types of rotations, i.e., increased rotation, 0 rotation, and reverse rotation of the input shaft, are selectively obtained, and any one of the first, second, and fourth constituent elements of the main transmission mechanism is provided. By selectively restricting the rotation of the two components, the third component There multistage transmission according to claim 5, wherein when taken to obtain the gear position of the forward 11 speeds and one reverse speed.