JPS63303245A - Planetary gear type multiple stage transmission device - Google Patents

Abstract

PURPOSE:To prevent the occurrence of a gear-change shock by providing a one-way clutch on each of connecting passages between an input shaft and the sun gear of a double pinion type planetary gear train, between the sun gear and a case, between the carrier of a single pinion type planetary gear train and the case, and between its sun gear and the case. CONSTITUTION:The gear change between the first gear stage and the second gear stage can be attained merely by the release of a second clutch C2 and the engagement of a third clutch C3. The gear change between the second gear stage and the third gear stage can be carried out merely by the engagement or release of a second brake B2. The gear change between the third gear stage and the fourth gear stage can be carried out merely by the engagement or release of a first clutch C1. And, the gear change between the fourth gear stage and the fifth gear stage can be carried out merely by the engagement or release of a brake B3. Thus, since most of the gear changes can be attained merely by means of the engagement or release of only one frictional engaging element, it is prevented that engine speed is increased, that a large load is generated, or that a gear ratio in the opposite direction to the corresponding gear change is attained resulting from the engagement between plural frictional engaging elements or the shift in the timing of the engagement.

Description

[Detailed description of the invention] [Industrial application field]

The present invention relates to a planetary gear type multi-stage transmission, in particular, a planetary gear type multi-stage transmission which is used together with a hydraulic torque converter and the like for an automatic transmission of a vehicle, and is capable of obtaining five forward speeds and one reverse speed. Regarding.

[Conventional technology]

In automatic transmissions for vehicles, increasing the number of gears allows the difference in rotational speed between the pump side and turbine side of the torque converter to be kept small during gear changes.
The efficiency of the torque converter can be increased accordingly. As a result, the fuel efficiency of the entire automatic transmission can be improved. Furthermore, by being able to reduce the relative rotational speeds of the pump side and the turbine side during gear shifting, it is possible to reduce gear shift shock accordingly. Furthermore, if the number of gears is large, a wide range of gear ratios can be achieved without placing an excessive burden on the engine or torque converter. In view of these points, with the development of clutch switching technology, the number of gears has been gradually increased from two to three and from three to four. Under these circumstances, as a five-speed transmission device, the one disclosed in, for example, Japanese Utility Model Application Laid-open No. 117950/1983 is known. This transmission includes a single pinion type planetary gear train 1 and a double pinion type planetary gear train 2, and these two planetary gear trains are connected to three clutches 1 to 3 and three brakes B1 to B3. control, 5 forward speeds,
This achieves one reverse gear. The connection state of each element is
The skeleton diagram is shown in FIG. 9, and in order to achieve each gear, each clutch 1 to 3 and each brake B1 to B3 are selectively engaged as shown in FIG. 10. .

[Problems to be solved by the invention]

However, when the above-described conventional transmission device is used in an actual vehicle, there is a problem in that the shift shock sometimes becomes extremely large. That is, according to the above-mentioned conventional transmission, shifting from the first gear to the second gear, for example, is performed using the clutch C1 and the brake B3.
Clutch C2 and brake B1 must be released at the same time that the brake is brought into the Sendai state. Generally, when changing gears by simultaneously applying two or more frictional engagement elements like this, if the timing is off, a neutral state will momentarily occur, causing the engine to rev up or the transmission to be reversed. The system may lock up, creating an extremely large load, or a gear ratio may be achieved that deviates in the opposite direction to the gear shift in question. In particular, when changing gears in which four frictional engagement elements are simultaneously switched in this way, it is extremely difficult to maintain the timing in an optimal state at all times. Shifting between the second gear and the third gear, shifting between the third gear and the fourth gear,
Furthermore, the shift between the 4th gear and the 5th gear is also done by one r each.
! This is achieved by simultaneously engaging the J frictional engagement element and releasing the other frictional engagement elements, and it is thought that similar problems will occur.

[Purpose of the invention]

The present invention was made in view of such conventional problems, and it is possible to achieve most speed changes by only engaging or disengaging one frictional engagement element, and, for example, it is possible to achieve most speed changes by engaging or disengaging only one frictional engagement element. It is an object of the present invention to provide a planetary gear type multi-stage transmission that does not cause new problems such as lack of gear movement. [Means for Solving the Problems] -4 The present invention provides an input shaft, an output shaft disposed coaxially with the input shaft, and a single pinion type planetary gear train disposed coaxially with the input and output shafts. and a planetary gear type multi-stage transmission comprising a double pinion type planetary gear train and a case housing them, wherein the input shaft is connected to the ring gear of the double pinion type planetary gear train and the single ring gear via a first clutch. It is connected to the carrier of the pinion type planetary gear train, and is also connected to the carrier of the double pinion type planetary gear train and the sun gear of the single pinion type planetary gear train via a second clutch, and is connected to the third clutch and the first gear. The output shaft is connected to the sun gear of the double pinion type planetary gear train through a series connection of one-way clutches and a parallel connection of a fourth clutch.
The carrier of the single pinion type planetary gear train is connected to the case through a parallel connection of a first brake and a second one-way clutch; A sun gear of the pinion type planetary gear train is connected to the case via a series connection of a second brake and a third one-way clutch and a parallel connection of the third brake, and the sun gear of the double pinion type planetary gear train is Sun gear is the 4th
The above object is achieved by connecting the brake and the fourth one-way clutch in series and the fifth brake in parallel to the case.

[Operation and effects of the invention]

In the present invention, a connection path between an input shaft and a sun gear of a double pinion type planetary gear train, a connection path between a carrier and a case of a single pinion type planetary gear train, and a connection path between a sun gear and a case of a single pinion type planetary gear train are provided. , and a one-way lug is arranged in the connecting path between the sun gear and the case of the double pinion type planetary gear train. As a result, most gear shifts can be achieved by only engaging or disengaging only one frictional engagement element (only two gear shifts between the first gear and the second gear). As a result, the engine may rev up, excessive load may occur, or a gear ratio in the opposite direction to that of the gear shift may be achieved due to misalignment or timing of multiple frictional engagement elements. can be prevented. In addition, only the shift between the first gear and the second gear is achieved by simultaneously engaging or disengaging two frictional engagement elements. If the drive range is used only when required and normal drive range driving is started from the second gear, the design can avoid this inconvenience. Further, according to the present invention, such advantages can be obtained, and the engine brake can also be applied at any time in each gear stage. In addition, in the present invention, the types of each clutch and brake are not limited. For example, it may be a wet type multi-disc clutch or a band brake.

【Example】

Embodiments of the present invention will be described in detail below based on the drawings. FIG. 1 shows an embodiment of a planetary gear type multi-stage transmission according to the present invention. This transmission includes an input shaft IS, an output shaft O8 disposed coaxially with the input shaft IS, a single pinion type planetary gear train 10 and a double pinion type planetary gear train 10 disposed coaxially with the input/output shafts IS and O3. row 20 and a case 90 that accommodates them, and further includes first to fourth clutches 01 to C4 and first to fifth brakes B1 to B5.
, and first to fourth one-way clutches F1 to F4. The input shaft IS can receive an output from an engine (not shown) via a well-known torque converter or an electromagnetic clutch. The output shaft O8 is connected to a front wheel or a rear wheel via a well-known differential device (not shown). The single-binion planetary gear train 10 includes a sun gear 1
1. Planetary pinion 12 meshing with the sun gear 11
, a ring gear 13 meshing with the planetary binion 12
, a carrier 14 supporting the planetary pinion 12
This is a well-known simple planetary gear train with The double pinion type planetary gear train 20 includes a sun gear 21
, a first planetary pinion 2 meshing with the sun gear 21
2. A well-known simple device having a second planetary pinion 23 that meshes with the first planetary pinion 22, a ring gear 24 that meshes with the second planetary pinion 23, and a carrier 25 that supports the first and second planetary pinions 22, 23. It is a planetary gear train. The input shaft IS is connected to the ring gear 24 of the double pinion type planetary gear train 20 and the carrier 14 of the single pinion type planetary gear train 10 via the first clutch C1. Also, this input shaft IS is connected to the second clutch C.
2 to the carrier 25 of the double pinion type planetary gear train and the sun gear of the single pinion type planetary gear train 10. Further, this input shaft Is is connected to the double pinion type planetary gear train via a series connection of the third clutch C3 and the first one-way clutch F1 and a parallel connection of the fourth clutch C4. It is connected to 20 sun gears 21. The output shaft O8 is always connected to the ring gear 13 of the single pinion type planetary gear train 10. Carrier 14 of the single pinion type planetary gear train 10
is connected to the case 90 via a parallel connection of a first brake B1 and a second one-way clutch F2. Sun gear 11 of the single pinion type planetary gear train 10
is connected to the case 90 via a series connection of a second brake B2 and a third one-way clutch F3 and a parallel connection of a third brake B3. The sun gear 21 of the double pinion type planetary gear train 20 is connected to the case 90 via a series connection of a fourth brake B4 and a fourth one-way clutch F4 and a parallel connection of a fifth brake B5. is connected to. Each of the forward gears from the first gear to the fifth gear and the reverse gear are determined by using the first to fourth clutches C1 to C4 and the first to fifth brakes B1 to B5 in a well-known manner as shown in FIG. This is achieved by controlling as shown. In this case, the gear ratio at each gear stage is as explained below. In the following, F1 is the number of teeth of the sun gear 11 of the single and nion type planetary gear train 10/the number of teeth of the ring gear 13, and F2 is the number of teeth of the sun gear 21- of the double pinion type planetary gear train 20/the number of teeth of the ring gear 24. It is the number of teeth. Incidentally, the rightmost end of FIG. 2 shows the actual gear ratios when ρ is set to 0, 42, ρ220, and 50. The explanation will be given below in order starting from the first gear. As is clear from Fig. 2, in the first gear, the second
Clutch C2 and fourth brake B4 are brought to a stopped state. In addition, in connection with this, the fourth one-way clutch F4 enters a stopped state. As a result of these engagements, power transmission occurs as shown in FIG. That is, the power input from the input shaft Is is transmitted to the carrier 25 of the double pinion type planetary gear train 20 via the second clutch C2. Double pinion type planetary gear train 20
, the fourth brake B4 (fourth one-way clutch F
4), the sun gear 21 is stopped. As a result, the power transmitted to the carrier 25 is transferred to the ring gear 2.
4. Carrier 14 of single pinion type planetary gear train 10
, the planetary pinion 12, the sun gear 11, and the carrier 25 of the double pinion type planetary gear train 20, and is taken out from the ring gear 13 of the single pinion type planetary gear train 10, and reaches the output shaft O8. The gear ratio at this time is 1/(1-ρ2-ρ1·F2). The engine brake can be applied by engaging the fifth brake B5. As is clear from Fig. 2, in the second gear, the third
Clutch C3 is brought to a stopped state. This engagement brings the first and second one-way clutches F1 and F2 into a stopped state. Although the fourth brake B4 is also in the stopped state, it does not particularly contribute to power transmission from the input shaft IS to the output shaft O8. As shown in FIG. 4, the power input to the human power shaft IS is transmitted to the sun gear 2 of the double pinion type planetary gear train 20 via the third clutch C3 and the first one-way clutch F1. In the double pinion type planetary gear train 20, the second one-way clutch F2 is in a stopped state, so the ring gear 24 is stopped. As a result, the power transmitted to the sun gear 21 is transmitted from the carrier 25 to the sun gear 11 of the single pinion type planetary gear train 10. In the single pinion type planetary gear train 10, the carrier 14 is engaged by the second one-way clutch F2.
has been stopped. As a result, the power transmitted to the sun gear 11 is transmitted to the planetary pinion 12 and further to the ring gear 13, and reaches the output shaft O3. The gear ratio at this time is (1-F2)/ρ1·F2. The engine brake can be applied by engaging the fourth clutch C4 and the first brake B1. In the third gear stage, the third clutch C3 and the second brake B2 are engaged. This engagement brings the first one-way clutch F1 and the third one-way clutch F3 into a stopped state. Although the fourth brake B4 is also in the stopped state, it does not particularly contribute to power transmission from the input shaft IS to the output shaft O8. By engaging the third clutch C3, the power transmitted to the input shaft IS is transmitted as shown by the arrow in FIG. 5, and is transmitted to the sun gear 21 of the double pinion type planetary gear train 20. In the double pinion type planetary gear train 20, by engagement of the second brake B2 (one-way clutch F3),
Carrier 25 is stopped. Therefore, sun gear 21
The power transmitted to the first planetary binion 22
, the second planetary pinion 23 and the ring gear 24, and then to the carrier 14 of the single pinion type planetary gear train 10. Single pinion type planetary gear train 1
0, the sun gear 11 is stopped by engagement of the second brake B2 (third one-way clutch F3). Therefore, the power transmitted to the carrier 14 is transmitted to the ring gear 13 via the planetary pinion 12, and reaches the output shaft O8. The reduction ratio at this time is 1/ρ2・
(1+ρ1). The engine brake can be applied by engaging the fourth clutch C4 and the third brake B3. In the fourth gear, the first clutch C1 and the third clutch C3 are in a stopped state. This engagement brings the first one-way clutch F1 into a stopped state. It should be noted that the second brake B2 and the fourth brake B4 are also in the stopped state, but this engagement is caused by the rotation from the input shaft IS to the output shaft O3.
It does not particularly contribute to power transmission to. Since the clutch C1 is engaged, the power transmitted to the input shaft Is is transferred to the ring gear 24 of the double pinion type planetary gear train 20 (and the single pinion type planetary gear train 10), as shown in FIG. is transmitted to the carrier 14). On the other hand, the power transmitted to the input shaft IS is transmitted to the sun gear 21 of the double pinion type planetary gear train 20 via the third clutch CB (and the first one-way clutch Fl).
It is also transmitted to The rotational speed of this sun gear 21 is the same as the rotational speed of the ring gear 24 described above. Therefore, the double pinion type planetary gear train 20 rotates together with the carrier 25. This rotation causes the sun gear 11 (and carrier 14) of the single pinion type planetary gear train to rotate at the same time. Therefore, the single pinion type planetary gear train 10 is also rotated as a whole. Along with this integral rotation, the power taken out from the ring gear 13 reaches the output shaft OS. Since the rotational speed of the output shaft O8 is the same as the rotational speed of the input shaft IS, the gear ratio is 1. The engine brake can be applied by engaging the fourth clutch C4. In the fifth gear, the first clutch C1 and the third brake B3 are in a stopped state. Note that the third clutch C3,
Although the second brake B2 and the fourth brake B4 are also in a stopped state, their engagement does not particularly contribute to power transmission from the input shaft IS to the output shaft O8. The power transmitted to the input shaft IS is shown in Figure 7 as = 1
6 = As shown, the first clutch C1 is engaged, thereby transmitting the signal to the carrier 14 of the single pinion type planetary gear train 10. Single pinion type planetary gear train 1
0, the sun gear 11 is stopped because the third brake B3 is engaged. Therefore, the power transmitted to the carrier 14 functions to cause the planetary pinion 12 to revolve around the sun gear 11, and is transmitted to the ring gear 13 at an increased speed, and reaches the output shaft O8. As a result, the gear ratio becomes 1/(1+ρ1). In addition, in the fifth speed stage, since a one-way clutch is not interposed in the power transmission system, engine braking can be ensured without particularly engaging a new frictional engagement element. In the reverse gear, the second clutch C2 and the first brake B1 are engaged. As shown in FIG. 8, as a result, the power transmitted to the input shaft IS is transferred to the carrier 25 of the double pinion type planetary gear train 20 via the clutch C2.
, and is further transmitted to the sun gear 11 of the single pinion type planetary gear train 10. Single pinion type planetary gear train 1
At point o, the carrier 14 is stopped by the engagement of the first brake B1. Therefore, the power transmitted to the sun gear 11 is transmitted to the ring gear 13 via the planetary pinion 12 in a reversely decelerated state, and reaches the output shaft O8. The gear ratio at this time is -1/ρ1. In particular, engine braking can be ensured without engaging other frictional engagement elements. As is clear from the above description, in this embodiment, the shift between the first gear and the second gear can be achieved only by disengaging the second clutch C2 and engaging the third clutch C3. Shifting, which was conventionally achieved by engaging or disengaging four friction engagement elements, can be greatly simplified, and shift shock can be reduced accordingly. For example, in the case of shifting from the first gear to the second gear, the second
When the clutch C2 is released, the second one-way clutch F2 corresponding to the first brake B1 is automatically engaged, and when the third clutch C3 is engaged, the fourth one-way clutch F4 corresponding to the fourth brake B4 is automatically engaged. This is to open up. This is reversed from the second gear to the first gear. In this case, the fourth brake B4
does not particularly contribute to power transmission from the second gear to the fifth gear, but by keeping it engaged in these gears, the fourth
This eliminates the need to newly engage the brake B4. This fourth brake B4 is present to contribute to securing the capacity in the first gear and to be able to be released in the reverse gear. Shifting between the second speed and the third speed can be performed only by engaging or releasing the second brake B2. When the second brake B2 is engaged or released, the second one-way clutch F2 corresponding to the first brake B1, which was in the engaged state in the second gear, is
This is because they are automatically opened or engaged. Shifting between the third gear and the fourth gear can be achieved only by engaging or disengaging the first clutch C1. This is because when the first clutch C1 is engaged or disengaged, the third one-way clutch F3 corresponding to the second brake B2 is automatically disengaged or engaged. Although the brake B2 does not particularly contribute to power transmission in the fourth and fifth gears, it remains engaged in these gears as well. As a result, there is no need to newly engage the second brake B2 when downshifting to the third gear. Shifting between the fourth and fifth gears can be achieved only by engaging or releasing brake B3. this is,
This is because the first one-way clutch F1 is automatically released or engaged when the third brake B3 is engaged or released. Note that this third brake B3 was used to apply engine braking in the third gear, but here it is used as a brake to actively achieve the fourth and fifth gears. It is characterized by its functioning. This means that if you try to change gears with only the second brake B2 in series with the third one-way clutch, the gear changes between the second and subsequent gears will be performed by the engagement of one rJ friction engagement element. This is a lick that cannot be achieved only by opening or closing. Although the third gear C3 does not particularly contribute to power transmission in the 55th Lr gear, by keeping it engaged in the 5th gear, the shift from the 5th gear to the 4th gear is possible. Brake B
This can be done only by opening No. 3. According to the transmission according to this embodiment, the shift that was conventionally achieved by simultaneously engaging or disengaging four or two P friction engagement elements can be achieved by using only two or one rM friction engagement elements, respectively. This can now be achieved by simply engaging or disengaging, and engine braking can be applied when necessary in each gear.

[Brief explanation of drawings]

FIG. 1 is a skeleton diagram showing an embodiment of a planetary gear type multi-stage transmission according to the present invention, and FIG. 2 is a diagram showing the engagement state and gear ratio of each clutch and brake in the embodiment described above. Figures 3 to 8 show the 1st to 5th gears, respectively.
and a diagram showing the power transmission state in the reverse gear; FIG. 9 is a skeleton diagram showing an example of a conventional planetary gear type multi-stage transmission;
FIG. 10 is a diagram showing the engagement states of each clutch and brake in the conventional example. IS...Input shaft, O8...Output shaft, 10...Single pinion type planetary gear train, 20...
Double pinion type planetary gear train, 01~C4...1st~
4th brake, B1 to B5... 1st to 5th brake,
F1 to F4...first to fourth one-way clutches.

Claims (1)

[Claims] (1) An input shaft, an output shaft arranged coaxially with the input shaft, a single pinion type planetary gear train and a double pinion type planetary gear train arranged coaxially with the input/output shaft, and a case housing them. In the planetary gear type multi-stage transmission, the input shaft is connected to the ring gear of the double pinion type planetary gear train and the carrier of the single pinion type planetary gear train via a first clutch, and The fourth clutch is connected to the carrier of the double pinion type planetary gear train and the sun gear of the single pinion type planetary gear train through two clutches, and the third clutch and the first one-way clutch are connected in series. The output shaft is connected to the sun gear of the double pinion planetary gear train through a parallel connection, the output shaft is constantly connected to the ring gear of the single pinion planetary gear train, and the carrier of the single pinion planetary gear train is connected to the sun gear of the double pinion planetary gear train through a parallel connection. , 1st
The sun gear of the single pinion type planetary gear train is connected to the case through a parallel connection of a brake and a second one-way clutch, and the sun gear of the single pinion type planetary gear train connects the second brake and a third one-way clutch in series, and the third brake. The sun gear of the double pinion type planetary gear train connects a fourth brake and a fourth one-way clutch in series and a fifth brake in parallel. A planetary gear type multi-stage transmission, characterized in that the planetary gear type multi-stage transmission is connected to the case via the above-mentioned case.