Multiple speed double planetary gear transmission

Multiple speed double planetary gear transmission

Abstract
A multiple speed gear transmission includes a pair of planetary gear configurations connected in series, the first output connected to the input of the second. Each of said planetary gear arrangement includes in its planets shafts, at least two different engagement surfaces of different diameters, and in one embodiment, are provided three such surfaces gears. A ring gear unit connecting the inlet of each such arrangement to a gear surface, and at least two ring gears, namely sun gears or crowns, are connected with each satellite further for providing different units forward and / or reverse. The planet carrier is connected to the output of each array. A direct drive clutch is provided for connecting the ring gear drive directly to the planet carrier direct drive. In one embodiment the ring gear drive arrangement is the first ring gear and ring gear drive of the second arrangement is a sun gear. In another embodiment, the drive gears of both provisions are annular ring gears.
Description
This invention relates to the provision of multi-speed gear, especially a type adapted to trucks, trucks and earthmoving equipment. Passenger vehicles have a high specific power, and therefore when using a hydraulic torque converter with a passenger car, which is normally sufficient for use in combination with the same two or three gear steps for both the acceleration and climbing, and to get through the direct drive torque converter lock-up and sufficient engine braking downhill runs. However, in the case of vehicles with a lower specific power, even if they have a lower maximum speed, you need to run many more steps to ensure that the engine can be used as close to the maximum power (hp ) as possible over a wide speed range and to maintain the highest possible average speed. Mechanical gears for such applications are typically synchronized transmissions designed to provide a sufficient number of well-spaced gear ratios. These transmissions have a number of highly loaded dog clutches synchronization arrangements along with a link type release cutting torque transmission during periods of change of the friction. When using the combination of hydrodynamic torque converter and planetary gears, satellites have been higher than normal transmissions and provide synchronized reverse gear that have had an additional planetary gear. Dimensions, weight and manufacturing costs of these gears are quite large.It is a purpose of the present invention is to provide a multi-gear transmission for speed multistep trucks, trucks and earth moving equipment which is relatively simple to manufacture, compact in size and includes a large number of gear steps properly appropriate, including reverse and the transmission does not need special couplings except the rate of change of power.The purpose of the present invention is achieved by providing a pair of multiple planetary gear speed, also referred to as "planetary gear arrangements", connected in series, each planetary gear arrangement of the series having only one planet carrier , each of which carrier is mounted on each of their axes at least two surfaces planet gears of different diameters, in which the surfaces of the planetary gears of each shaft involve at least three ring gears. The term "crown gear" as used herein, refers to either a gear or crown. One of the ring gears is a ring gear which forms the driving connection unit with the input member of the planetary gear arrangement in which the planet carrier is operatively coupled to the output member of said planetary gear device. All except the aforementioned ring gears drive ring gear of each planetary gear arrangement can be connected to the housing by means of a servo motor mounted in the housing and operable to engage a friction brake associated with the ring gear .

By using one of the ring gears, ie the drive sprocket, as the input to the planetary gear unit, and causing the planet carrier as the connection to the output shaft or the same, and said two surfaces having different diameter gears in each planetary gear engaged with said ring gear remaining, it may cause the brake to such ring gears of a relatively large diameter, making it possible to use an air bearing effect to reduce drive torque and providing sufficient surface area for the changes between gear steps without the surfaces of brake rotors too close to one another. Furthermore, since all the servo motors for the brake are mounted in the housing, it becomes easier to provide seals that will not leak, so it is also possible to use only small amounts of oil with a high pressure buildup. Furthermore, with this type of planetary gear, it is possible to provide a reverse gear as required.To increase the number of gear steps in each planetary gear system and in general two series planetary gear, and to provide a direct drive to each planet gear in a preferred embodiment of the invention, there are provided clutches of friction between the primary and secondary sides of each planet gear, namely between the drive ring gear and the planet carrier. These friction clutches direct drive servo motors are driven by stationary mounted within the housing and mounted such that the force of the same passes over and thrust bearings acting on a lever system operatively connected to cause the coupling of the friction clutches. This is particularly advantageous since the bushings or ring gears are loaded while the connection is made direct drive or while driving in the direct transmission. This will obviously save wear on the gear surfaces.According to a first embodiment of the invention the primary gear, ie, the unit of the first ring gear of the planetary gear in the direction of power flow, operatively coupled to the larger surfaces of the planetary gear, and drive the ring gear is preferably a crown. Simultaneously, the primary gear or ring gear unit in the second planetary gear is a planetary gear meshing with the surface of smaller-diameter gear. This arrangement provides a greater number of gear steps. Such a transmission is especially suitable for earth-moving equipment, because for such applications, the steps of low-speed gears must be close to each other, while at high speeds the gears are used for transport purposes. The presently described embodiment may have, as a practical form of the invention, up to nine forward speeds, including direct drive and reverse.Another preferred embodiment of the invention for use on trucks, or the like. In this embodiment, the primary gear or the ring gear drive system both planetary gears ring gear are in mesh with the surface of larger-diameter gear of the respective planetary gears. In this arrangement, the steps of low-speed gear are larger than in the first described embodiment and the steps are smaller at the higher speeds, these characteristics are desirable for trucks and lorries. The number of possible combinations of gear ratios across the system are closer together so that only six are used. However, this is normally sufficient. A greater number of gear steps useful course be provided if the planetary gears, instead of having just two toothed surfaces have three gear surfaces thereon, each of a different diameter.In each of the planetary gears, means are provided for connecting the primary gear, ie the ring gear to drive the planetary carrier for direct drive connection. Furthermore, as noted above, each of the remaining sun and ring gears can be connected to the housing by the brakes. With this arrangement, provided according to an advantageous feature of the invention, a provision that the strength of the stationary servo motor operable to engage the drive clutch transmits its power directly from the servo-motor Through one of sun gears which slides axially and one or two thrust bearings axial force, so that the sun gear, acting through said bearing, moves a lever mechanism, which in turn makes engaging the friction clutch direct drive.In many cases, it is necessary to have a high ratio of the reverse gear. This is also possible in accordance with the present invention by providing three gear surfaces of different diameters on each planet shaft, and providing a ring gear engaged with the outer side surface of the smaller gear and the stagnation of this ring gear respect to the housing. It is also possible to provide an additional sun gear engaged with the third engagement surface to provide a smaller transmission ratio for use with the primary ring gear or drive, which adds three gear ratios for the number of relationships transmission in the entire system.Friction brakes may be disc brakes, or in the case of power transfer arrangements, which may include a plurality of disks.The objects and advantages of the invention will become apparent from the following detailed description together with the accompanying drawings in which:. Figure 1 is a longitudinal axial sectional view through a first embodiment of a multi-speed gear according to the present invention, particularly adaptable for earthmoving equipment.. 1a is a diagram relating to the embodiment of FIG. 1, which represents the tensile stress and the speed of the motor for driving the various steps plotted against the vehicle driving speed in earthmoving vehicle is also capable of traveling at high speed between operating sites individual.. Figure 2 is an axial sectional view longitudinal through a second embodiment of a multi-speed gear of the present invention, particularly suitable for a truck or truck.. 2a is a diagram similar to FIG. 1a for a truck engine.. Figure 3 is a schematic representation of another embodiment of the invention which is similar to that shown in Fig. 1 but modified such that the first planetary gear has three axes planet gear surfaces of different diameters each meshing with a sun gear and / or ring gear.. 3a is a diagram similar to FIG. 1a, but corresponding to the embodiment of FIG. Three.Referring now to the drawings, like elements are represented by like numerals throughout the several views.In all embodiments of the transmission multi-speed gear according to the invention illustrated in the drawings, the input shaft is indicated by the reference I, the output shaft of O, an intermediate shaft between the two planetary gear S and the non-rotatable housing, which shows, by H.In the embodiment of the transmission shown in Fig. 1, the two axes I and O are mounted in the housing H by means of ball bearings 10 and 12, respectively, and the input shaft I, which extends over almost the entire length of the transmission is focused by means of a roller bearing 16 in an exit orifice in the short axis O. S hollow shaft encloses the intermediate input shaft for most of its length and its front end is mounted on the input shaft I via a needle bearing 14 and its rear end is mounted by a bearing R Friction 17. The housing H contains two planetary gear P1a and P1b which are separated from one another by a middle portion M housing formed by a partition, and numerous non-rotating servo motors.P1a of the left planetary gear will first be described: A radial flange 18 on the intermediate shaft is splined S to carry a planet carrier 20 having first planet gear wheels 24 of relatively small diameter mounted on their shafts 22 which are distributed over circumference. Only one shaft 22 and planet gear wheel 24 can be seen in the drawing plane. Mounted on each gear wheel 24 is a second gear wheel 26 in positive engagement with the gear wheel 24 through a slotted hole. Gear wheel 26 has a substantially greater diameter, but is axially much shorter than the gear wheel 24, so that leaves a substantial portion of the exposed smaller wheel 24, the wheels 24 and 26 thereby forming a planetary gear surfaces having two gears of different diameters.A disc 28 is slightly conical non-rotatably fixed to the input shaft I by splines. The periphery of this disc 28 engages a locking mode with the internal teeth of a first ring gear 30 which is in the form of a sleeve. Internal teeth which extend over the entire length of the mesh gear 30 at its other end with the large planet gears 26 to the ring gear 30 thus forms a gear ring. A second ring gear 32 in the form of a sleeve S is focused on the intermediate shaft and meshes with a planetary gear as the big planetary gears 26. A third ring gear 34, which is also in the form of a sleeve is mounted on a disc flange 36 that has been pushed onto the front end of the sun gear 32 and non-rotatably fixed therein by splines. This ring gear 34 functions as a gear that meshes with the small planet gears 24. The third ring gear 34 also carries a disk clamp 38 which is mounted thereto by splines. Finally, a fourth ring gear 40 functions as a ring gear which encloses the small planetary gears 24 on its outer side and engages with them.Flanged disk has an external toothing 38 which carries the single friction disc 42 of a first disc brake for fixing the sun gear 34 to the housing H. A disc brake coupling is moved by means of an annular piston 44 which is pushed in the direction of disengagement by a Belleville spring 46. The disc flange 36 also carries the friction disc 48 of a second single disc brake B for fixing the sun gear 32 to the housing H by means of a piston ring 50 which is pushed in the direction of disengagement by Belleville spring 52. Ring pistons 44, 48 are mounted slidably in annular cylindrical recesses in the middle portion M of the casing and the Belleville springs 46, 52 bear against the axial projections on the inside of the housing.

Another piston ring 54 is disposed on a relatively small diameter in a third annular cylindrical recess in the housing middle portion M. This ring portion 54 acts through a first thrust bearing 56 on the sleeve-shaped sun gear 32 which is not only rotatable but also axially displaceable on the intermediate shaft S. The other end of the sun gear 32 acts via a second thrust bearing 58 in a number of spacer pins 60 which are supported to be axially movable in holes formed in the radial flange 18 of the intermediate shaft S. In the example illustrated, the thrust bearings 56, 58 are needle bearings with needles relatively short but other suitable types of bearings can be used equally well, in particular, the bearing 56 may be a sliding bearing and the bearing 58 rows deep balls.The pins 60 bear axially against a thrust ring 62 which is displaceable on the intermediate shaft S. The thrust ring 62 engages with the inner ends of several radial levers 64 which extend through the slots with beaded edges formed in a neck 66 in the planet carrier 20. The shorter lever arms located radially outside these areas are engaged in a form-locking manner with another thrust ring 68. This thrust ring 68 clamps a friction disc 70 fixed on the ring gear 30 and thus, together with the disc 70 forms a clutch disc crown C between the planet carrier 30 and 20 to directly connect the axle I input the S output shaft for direct drive. Will see that this disc clutch is engaged by the pressure acting on the annular piston 54. The clutch is disengaged by a Belleville spring 72 which is screwed to the radial lever 64 and tends to pivot the lever in a position perpendicular to the transmission axis and the thrust bearings kept constantly under load.Finally, the non-rotating housing H has another annular cylindrical recess containing a piston 74 by means of which a friction disc 76 located at the outer teeth of the ring gear 40 can be fixed to the non-rotating casing to form another disk Brake D.The planetary gear P1b planetary gear differs mainly P1a to the input member, that is, the driving ring gear connected to the input shaft, ie in this case the intermediate shaft S, there is a gear ring, but a gear 78 meshing with the wheel of smaller diameter planetary gears. These planetary gears 80 are again formed by relatively large gear wheels are mounted on shafts 82 of a planet carrier 84 which in turn is bolted to a flange on the output shaft O. Gear wheels 86 of greater diameter are pushed onto the gear wheels 80 and formed above the larger wheels of the planetary gears. Each of the brakes A, B and D include reaction members 44a, 50a and 74a, respectively, fixed against rotation relative to the housing.In order to obtain two different gear ratios of planetary gear P1b, a ring gear 88 is provided to mesh with the large planet wheels 86 and provides a gear ring 90 to engage with the small planet gears 80. The ring gears 88 and 90 carry friction discs 92 and 94, respectively, by means of which can be fixed to the transmission case H by ring pistons 96 and 98, respectively, to form a brake disk E and F. Other Sun Gear 102 is formed by a toothed sleeve mounted on the shaft S and carrying intermediate, non-rotatably attached thereto, a disc flange 104 which carries the single friction disk 106 of a friction clutch G for connecting the sun gear 102 to the planet carrier 84 by means of a thrust ring 110 to form a direct driving connection. To operate the clutch G, a piston ring 112 axially movable within an annular cylinder is provided at the end of the planetary gear drive P1b. The piston ring 112 is capable of axially displacing the thrust ring 116 by a thrust bearing 114 designed as a needle bearing. The thrust ring 116 is coupled to the inner ends of radial levers 118 which are formed inwardly directed spokes extending from a Belleville spring 120 that bears at its periphery against a stop 122 in the planet carrier 84 and, radially inward of this position, engages an annular projection of the thrust ring 110. It is obvious that by this construction, the pressure exerted by the annular piston 112 to push the pressure ring 110 against the friction disc clutch 106 is greatly enhanced. Since moreover, unlike the arrangement in the planetary gear P1a, the servomotor formed by the piston ring 112 is located at the drive end of the planetary gear P1b and the sun gear 102 is required to be attached to the planet carrier 84 for direct drive are also available there, it becomes unnecessary to use a planetary gear as a power transmission element axially movable for operation of the direct drive clutch and the planetary gear P1a.Table I below shows the combinations of certain pairs of brakes and / or dog clutches at the same time for producing nine forward gears indicated by Roman numerals and the reverse gear R. The second and fourth columns give the gear ratios of the first and second planetary gears, respectively, while the last column shows the overall gear ratio obtained.
. 1a is a diagram representing the variation in the rotation speed n1 and tensile force P transmission input regarding the vehicle speed into the nine forward I-IX when the when the multi-speed gear according to FIG. 1 is used in a truck engine. The vertical lines between the input velocity curves give the points at which it is appropriate to change speed between the individual forward. One can see that the revolution speed curves are close together in the lower gears. This means that when the vehicle is traveling at relatively low speeds and the pulling force is correspondingly high it is possible to operate within a relatively narrow range of input speeds of rotation, in which the engine produces its maximum torque. The multi-speed planetary gear as shown in Fig. 1 is therefore able to adapt perfectly to the different demands of tractive power and travel speed are in an excavator.Transmission according to Fig. 2, which is intended primarily for motor trucks, comprising two planetary gears P2a and P2b that are basically similar in construction and in which, as in the first planetary gear P1a of the embodiment according to FIG. 1 the inlet members are formed by external ring gear meshing with the larger diameter wheels planetary gears. In this case, however, the additional ring gear meshing with the planet wheels of smaller diameter and can be braked to produce reverse drive are provided only in the first planetary gear P2a.Moreover, the arrangement of FIG. 2 contains certain structural changes compared with Fig. 1 but they are irrelevant to the invention and the two planetary gear P1a and P1b in FIG. 2 therefore need not be described in detail. It just has to be noted that in view of the fact that the arrangement of FIG. 2 is to be used in a heavy truck engine, brakes and friction clutches are equipped with multiple disks. Also, since in this case the ring gear of the planetary gear drive P2b is a ring gear, this requires that the intermediate shaft being slightly shorter S and supported by the output shaft O closer to the front than in the embodiment of Fig. 1, ie, which leads 16a.Use of a designation for brakes and clutches to G 'similar to that used in Fig. 1 the gear combinations that can be obtained in the practice of the construction according to Fig. 2 are shown below:
The other possible combinations, G 'B', C '-e', and C ', G "are not used in practice since B' G 'and C'-E; are too close A'-E' and B '-F', respectively, whereas C ', G', in fact, produce the same overall reduction ratio as A'-F '. By contrast, the effect achieved with the choice of combinations is given to the shaft intermediate is then subjected to minimum loads at different stages I to V reduction from the second planetary gear P2b always participates in the reduction and direct clutch G is connected only to the gear transmission speed direct VI.. 2a is a diagram showing again varying the rotation speed n1 and the tensile force P in relation to the speed of movement in the six forward gear speeds I-VI when using a transmission in accordance with the transmission input figure. 2 in a truck engine. The diagram clearly shows the much thicker classification engine speed and the tractive force in the region of the lower traveling speeds in comparison to the results obtained with the transmission according to Fig. 1.. Figure 3 shows schematically a modified embodiment of the transmission of FIG. 1. The difference is that in this case, the planetary gears of the first planetary gear (left in Figure 3.) Have a third surface of the smaller diameter gear meshing with the second ring gear designed to be braked by D brake for reverse gear and a sun gear designed to be slowed further by an additional brake a0. The purpose of this provision is to obtain a further reduction ratio for the reverse and for extremely high reduction ratios additional forward drive, as represented by the curves of revolution n.sub.10 and n. sub.100 in Fig. 3A. To obtain a reverse gear ratio 5:1, while the proportions of the additional reduction unit is 8.4:1 forward and 7:01.The multi-speed planetary gear according to the present invention may be used alone or connected in series with a hydrodynamic torque converter having stage one-and-one-half or two stages. The sun gear of the present invention is particularly suitable for use with a torque converter having a pump member releasable type as shown in U.S. Pat. No. 3,893,551 or with a torque converter of the type having a releasable turbine member as shown for example in U.S. Pat. No. 3839864.Although the invention has been described in considerable detail with respect to preferred embodiments thereof, it will be apparent that the invention is susceptible of numerous modifications and variations apparent to those skilled in the art without departing from the spirit and scope of the invention as defined in the claims.