Abstract:
A multiple-speed ratio transmission mechanism for an automotive vehicle includes a torque input shaft, an output shaft, an intermediate shaft, these three shafts rotatably supported on a common axis, and a countershaft mounted parallel to the other shafts. Multiple pinion-gear pairs are supported on the three coaxial shafts and countershaft. A planetary gear unit, including a sun gear, ring gear, and a set of planet pinions rotatably supported on a carrier, has its input connected to the intermediate shaft and the output connected to the output shaft. Synchronizers drivably connect a member of the pinion-gear pairs to one of the shafts. A shift mechanism transmits motion of a shift lever to shift forks that move synchronizer sleeves into engagement with the selected gear.

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     This invention pertains to the arrangement of gears, shafts, and synchronizers of an automotive transmission, especially to such a transmission of the type having a countershaft or layshaft disposed parallel to a power input shaft. 
     2. Description of the Prior Art 
     When the engine of a motor vehicle is idling and the neutral clutch that connects the engine to a manual transmission is disengaged, the speed of the engine varies as a result of cyclic irregularity intrinsic in the operation of an internal combustion engine. Countershaft manual transmissions for automotive use contain gear sets whose members are in continuous meshing engagement, though the gear selector of the transmission is in the neutral position. When the maindrive clutch is engaged and the gear selector is in the neutral position, speed changes of the engine transmitted to the gearset produce noise called &#34;neutral rollover gear rattle.&#34; When the engaged gear teeth change position with respect to each other due to engine speed variations, the teeth of the driven gear contact alternately the driving face and the nondriving face of the driving pinion. 
     To reduce or eliminate this objectionable noise, various devices have been employed to take up clearances between the faces of the gear and mating pinion. These devices include a wheel coaxially arranged with the gear having similar teeth to those of the gear, and means for elastically maintaining the teeth of the gear in position that is angularly offset relative to the gear. A mechanical spring, usually a helical spring, located radially offset from the axis of the gear, applies a force directed tangentially to the axis tending to rotate the angular element with respect to the gear wheel. Other means such as elastomeric material have been used in the prior art to bias an angular member angularly with respect to a gear to maintain contact between the gear teeth in order to avoid gear rattle. 
     U.S. Pat. No. 4,718,295 describes use of a one-way clutch located on the input shaft between a gear wheel and a cylindrical surface of the shaft. The one-way clutch permits a one-way drive connection so that power can be transmitted through the clutch from the input shaft to either of two countershafts when fourth and fifth gear ratios are selected by the vehicle operator. 
     It is an object in the design of a manual transmission to minimize the effort required to make gear ratio changes, i.e., to reduce as much as possible the magnitude of force that must be applied to a shift mechanism to disengage a current gear ratio and to engage a new gear ratio. The magnitude of that force is to a great extent determined by the magnitude of inertia of rotating components that is reflected to the synchronizer units. Minimizing the weight of components of a motor vehicle is desired in order to reduce fuel consumption. Various attempts have been made to reduce noise produced by a manual transmission and to reduce rotating inertia of the gears, shafts, synchronizers and other such components. 
     SUMMARY OF THE INVENTION 
     It is an object of the present invention to provide a multiple-speed, motor vehicle manual transmission having minimum size, weight, and rotating inertia consistent with the requisite torque capacity. Use of a planetary gear unit to produce speed reduction in the lower speed ratios permits the diameters of mating pinions and gears to be smaller than otherwise they would be and helps to reduce the distance between the main shafts and a countershaft. The result is a compact transmission having a high power density, low rotating inertia and low weight. 
     In realizing these advantages the transmission according to the present invention includes an input shaft adapted for driveable connection to a power source, such as an engine, an output shaft arranged coaxially with the input shaft and adapted for connection to a load, and an intermediate shaft arranged coaxially with the input shaft and output shaft. A countershaft is disposed substantially parallel to the other shafts. 
     Multiple pairs of gears and pinions have one member of each pair supported on the countershaft and another member of each pair supported on either the input shaft, output shaft or intermediate shaft. These pairs of gears and pinions produce various speed reductions corresponding to the gear ratio selected by the vehicle operator. The gear of each gear-pinion pair that operates to produce the higher speed ratios is connected to the output shaft; the gear of each gear-pinion pair that operates to produce the lower forward speed ratios and the output member of a mechanism that operates to produce reverse drive are connected to the intermediate shaft. 
     A planetary gear unit is also employed to produce a speed reduction in certain selected gear ratios. The planetary gear unit includes a sun gear connected to the intermediate shaft, a ring gear fixed against rotation on the casing, a pinion carrier driveably connected to the output shaft, and a set of planet pinions in continual meshing engagement with the sun gear and ring gear, the pinion set being supported on the pinion carrier. The planetary gear unit can be arranged such that the sun gear is fixed against rotation on the casing, the ring gear is connected to the intermediate shaft, the pinion carrier is driveably connected to the output shaft, and the set of planet pinions is in continual meshing engagement with the sun gear and ring gear, the pinion set being supported rotatably on the pinion carrier. 
     Synchronizers releasably connect and disconnect a member of each pinion-gear pair to the shaft on which each synchronizer is supported. 
    
    
     BRIEF DESCRIPTION OF THE DRAWING 
     FIG. 1 a cross section taken through a vertical plane through the transmission according to this invention. 
     FIG. 2 a diagram representing the location of forward drive and reverse drive positions, among which the gearshift lever is moved to select the desired gear ratios. 
     FIG. 3 shows an alternate arrangement of the planetary gear unit in which the ring gear is driven and the sun gear is held. 
     FIG. 4 shows a mechanism for selecting and engaging the gear ratios of the transmission of FIG. 1. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring first to FIG. 1, the transmission, which is housed in mating casings, includes an input shaft 12 adapted for drivable connection to a power source such as an internal combustion engine. Located adjacent the input shaft is an output shaft 14 supported rotatably partially on a pocket bearing 15 located in a recess formed on the end of the input shaft. A sleeve shaft 16, aligned coaxially with the output shaft and input shaft, extends along a portion of the length of the output shaft. Bearings 11, 13 support these shafts on the casing 10. 
     A countershaft 18, supported at the transmission casing 10 on bearings 20, 22, is disposed parallel to shafts 12, 14, and 16. The countershaft carries a gear 24 associated with a fourth forward speed ratio; pinions 26, 28, 30, which are members of gear-pinion pairs associated with the third, first and second forward speed ratios; and pinion 32, a member of a creeper/low gear-pinion pair. Pinions 18, 30 are splined to countershaft 28. 
     Input shaft 12 carries a fourth speed ratio pinion 34, which is in continuous meshing engagement with gear 24. 
     Output shaft 14 supports gear 36, which is in continuous engagement with pinion 26, and gear 38, which is in continuous engagement with pinion 32. 
     Shaft 16 supports a first speed ratio gear 40, which continuously meshes with gear 28; and second speed ratio gear 42, which continually meshes with pinion 30. Second speed gear 42 includes a portion 44 that extends along the longitudinal axis on the outer surface of shaft 16 and supports a synchronizer 46 having hub, which is splined at 50 to gear 42. 
     Gears 36 and 38 are journalled on the outer surface of output shaft 14; gears 40 and 42 are journalled on the outer surface of shaft 16. 
     A 3-4 synchronizer 52 includes a hub 54 splined at 56 to the outer surface of shaft 14. Synchronizer 52 includes a sleeve 58, which moves rightward and leftward from the neutral position shown in FIG. 1 to positions of engagement with dog teeth 60, formed on the outer end of a conical ring 62 fixed to pinion 34, and dog teeth 64, formed on the surface of conical ring 66 fixed to gear 36. When synchronizer sleeve 58 moves leftward, synchronizer 52 connects input shaft 12 to output shaft 14. When sleeve 58 moves rightward, synchronizer 52 drivably connects gear 36 to output shaft 14. 
     Synchronizer 68 includes a sleeve 70 and a hub 72, which is splined at 74 to shaft 16. Sleeve 70 moves leftward into engagement with dog teeth formed on conical ring 76 fixed to gear 38, thereby drivably connecting gear 38 to shaft 16. The reverse drive sprocket wheel 80 is supported rotatably on an axially extending portion of gear 40. Countershaft 18 is formed with a sprocket wheel 94, which is aligned axially with sprocket wheel 80, these sprocket wheels being continually drivably connected by chainbelt 96. Sleeve 70 moves rightward into engagement with the dog teeth on conical ring 78 of sprocket wheel 80, thereby drivably connecting reverse sprocket wheel 80 to shaft 16. 
     Reverse drive can be produced instead by gears and a pinion. For example, sheave 94 and chain 96 can be replaced by a reverse pinion fixed to the countershaft at the location of sheave 94, and sheave 80 can be replaced by a reverse gear journalled on shaft 16 at the location of sheave 80. A reverse idler, rotatably supported on the casing and continually meshing with the reverse gear and reverse pinion, operates to reverse the direction of rotation of the reverse gear relative to input shaft 12. U.S. Pat. No. 4,377,093 describes a reverse drive gear mechanism of this type. 
     Synchronizer 82 includes a sleeve 84 having internal spline teeth that engage the dog teeth on synchronizer rings 86 and 88; and a hub 90, which is splined at 92 to the outer surface of shaft 16. When sleeve 84 moves leftward, it engages the dog teeth on conical ring 86 fixed to gear 40, thereby drivably connecting gear 40 to shaft 16. When sleeve 84 moves rightward, gear 42 is drivably connected to shaft 16 and gear 40 is released from shaft 16. 
     Planetary gear unit 98 includes ring gear 100, permanently held against rotation on casing portion 102; ring gear 104, permanently splined to shaft 16; planet pinions 106 in continuously meshing engagement with the ring gear and sun gear; and pinion carrier 108, which rotatably supports the set of planet pinions and is splined at 110 to output shaft 14. In the arrangement of FIG. 1, the sun gear is the torque input, the torque reaction is located at the ring gear 100, and the output is taken at the pinion carrier 108 and output shaft 14. 
     Synchronizer 46 includes a sleeve 112 having internal spline teeth that engage the dog teeth 114 on a conical ring, which is drivably connected to pinion carrier 108. When synchronizer sleeve 112 moves rightward, it produces a drivable connection between pinion carrier 108 and second speed ratio gear 42 through spline 50 and hub 48. 
     FIG. 3 shows an alternate arrangement of the planetary gear unit 98&#39; in which shaft 16&#39; is drivably connected to ring gear 100&#39; so that the ring gear and shaft 16&#39; turn as a unit. Sun gear 104&#39; is held against rotation on transmission casing 102&#39;. Carrier 108&#39; is drivably connected to output shaft 14&#39; and rotatably supports a set of planet pinions 106&#39;, which are in continuous meshing engagement with the ring gear and sun gear. In this arrangement, the ring gear 100&#39; is driven, the torque reaction is provided at the sun gear 104&#39;, and the output is taken on the carrier 108&#39; and output shaft 14&#39;. 
     The transmission casing includes a casing portion 120, which supports a shift lever 122 that moves among the gear selector positions shown in FIG. 2. Casing portion 120 supports a ball 124, through which shift lever 122 extends, thereby providing a universal pivoting support for the shift lever. A leaf spring 126 biases the shift lever to a particular predetermined position when the vehicle operator is not manipulating the shift lever. Preferably, the leaf spring biases the shift lever toward the neutral position in the 1-2 shift plane of FIG. 2. A seal 130 protects against passage of contaminants into housing 120, cover 128 closes the upper portion of the shift lever, and snap ring 132 maintains the position of the shift lever relative to the cover. 
     A shift rail 134, supported slidably for movement along its axis within bores formed on the casing 10, carries a pin 136 that extends radially into a system of interconnected slots formed in a gate plate 138, the slots defining the 1-2, 3-4, R-CL, and OD vertical planes and the horizontal neutral plane shown in FIG. 2. A gate plate and pin combination of this type is described and illustrated in U.S. Pat. No. 4,297,910, whose assignee is the owner of the present invention. 
     Pinned to shift rail 134 is a socket 140 having a bore 142, into which a spherical ball 144 located at the end of the shift lever is fitted. Shift rail 134 carries several shift forks 146, 148, and other forks not shown in FIG. 1, but illustrated in FIG. 4. Each shift fork engages a synchronizer sleeve 58, 70, 84, 112, thereby transmitting movement of the shift lever through socket 140, the shift fork associated with the selected ratio, and the appropriate synchronizer sleeve. Rail 134 carries a radially directed finger 150 and an interlock sleeve 152, which is held by bolt 154 against displacement along the axis of rail 134. 
     Each shift fork includes an arm that extends to the vicinity of bolt 154 or bolt 192 and provides a slot aligned with the radial path of the shift fingers 150 or 188 as it rotates about the axis of shift rail 134 due to displacement of the shift selector 122 along the neutral horizontal plane of the diagram of FIG. 2. When a selector finger rotates into engagement with a slot on a shift fork corresponding to the position of the selected gear ratio, the shift fork is subsequently moved axially causing the associated synchronizer sleeve to engage and to connect the selected gear, pinion, sleeve, or carrier with the shaft on which the hub of the corresponding synchronizer is supported. 
     The mechanism that transmits movement of the gear shift lever to the components of the transmission that engage and release the selected gear ratios is shown in greater detail in FIG. 4. In addition to the components of the shift mechanism described with reference to FIG. 1 above, the mechanism includes a lower shift rail 170, which is journalled at the left-hand side on casing 10, and at the right-hand side on a web 172 formed integrally with casing 10. Rail 170 carries a 1-2 shift fork 174, which is in continuous engagement with an annular recess 176 formed on the outer surface of the 1-2 synchronizer sleeve 84. 
     Bracket 178, fixed to the an inner surface of casing 10, provides a pivotal support 180 for a shift lever 182 that extends between lower rail 170 and control rail 134. At the lower end of the lever, a slotted hole 184 receives a pin 186, which extends radially from the axis of rail 170 and fits within hole 184 with a close fit along the axis of the lower rail and with a slotted fit in the direction perpendicular to the axis of the lower rail. The upper end of lever 182 is engageable by a finger 188, which extends radially from the axis of control rail 134 and rotates into engagement with the upper end of lever 182 when control rail 134 is rotated about its axis. Finger 188 moves along the axis of rail 134 during movement of shift lever 122, thereby pivoting lever 182 and displacing lower rail 170, the 1-2 shift fork 174, and synchronizer sleeve 84. This action moves the 1-2 synchronizer 82 into and out of drivable engagement with first speed gear 40 and second speed gear 42. 
     Control rail 134 supports a 3-4 shift fork 190, which is in continuous engagement with an annular slot on the outer surface of the 3-4 synchronizer sleeve 58. As control rail 134 rotates about its axis, shift finger 188 moves into engagement with fork 190 and then moves synchronizer sleeve 58 along the axis of rail 134 in response to movement of the shift selector 122. This action engages and releases a connection between output shaft 14 and third speed gear 36 or fourth speed pinion 34. Second interlock bolt 192, similar to bolt 154, holds an interlock sleeve 194, which surrounds rail 134, against axial displacement. 
     The shift mechanism interlock described in U.S. Pat. No. 4,377,093 is substantially identical to the interlock described here with reference to the shift control rail 134. 
     In operation, the lowest gear ratio, creeper/low, is produced by the transmission when synchronizer sleeve 70 is moved leftward, thereby drivably connecting creeper/low gear 38 to shaft 16. The torque delivery path includes input shaft 12, pinion 34, gear 24, countershaft 18, pinion 32, gear 38, synchronizer 68, shaft 16, planetary gear unit 98, carrier 108, and output shaft 14. The planetary gear unit produces a speed reduction between shaft 10 and output shaft 14 in addition to the speed reduction produced by pinion-gear 34, 24 and pinion gear 32, 38 between input shaft 12 and shaft 16. 
     The first forward gear ratio is produced by moving sleeve 84 of synchronizer 82 leftward into engagement with the dog teeth of first speed gear 40 and moving the other synchronizers to the neutral position. The torque delivery path between the input shaft and output shaft, which produces a slight speed increase through the main drive gearset that includes pinion 34 and gear 24, includes also countershaft 18, pinion 28, gear 40, synchronizer 82, shaft 16, planetary gear unit 98, and carrier 108, which drives output shaft 14. 
     The second forward speed ratio is produced when synchronizer sleeve 84 is moved rightward to connect the second speed gear 42 and shaft 16 while maintaining the other synchronizers in their neutral positions. The torque delivery path includes the main drive, pinion 34 and gear 4, countershaft 18, second ratio pinion 30, second ratio gear 42, synchronizer 82, shaft 16, planetary gear unit 98, and carrier 108, which drives the output shaft 14. 
     The third forward gear ratio results when synchronizer sleeve 58 is moved rightward to drivably connect third ratio gear 36 to output shaft 14. The torque delivery path includes input shaft 12, pinion 34, gear 24, countershaft 18, pinion 26, gear 36, synchronizer 52, and output shaft 14. 
     The transmission produces the fourth speed ratio, a direct drive connection between the input shaft and output shaft when synchronizer sleeve 58 is moved leftward. This action produces a drive connection between the clutch teeth 60 on synchronizer ring 62, which is formed integrally with input shaft 12, and output shaft 14, through synchronizer hub 54 and spline connection 56. 
     The fifth forward speed ratio, an overdrive ratio, is produced by moving synchronizer sleeve 112 rightward into engagement with planetary pinion carrier 114, thereby drivably connecting gear 42 and the output shaft 14 through synchronizer 46 and carrier 108. The torque delivery path includes input shaft 12, main drive pinion 34, gear 24, countershaft 18, second ratio pinion 30, second ratio gear 42, synchronizer 46, carrier 108, and output shaft 14. 
     The reverse drive ratio is produced by moving synchronizer sleeve 70 rightward into engagement with the dog teeth on cone ring 78 formed integrally with reverse output sheave 80. When synchronizer sleeve 70 engages the reverse sheave 80, it completes a drive connection between sheave 80 and shaft 16, which drives sun gear 104. The output is taken at pinion carrier 108 and output shaft 14. The torque delivery path for reverse drive includes input shaft 12, main drive pinion 34, gear 24, countershaft 18, reverse input sheave 94, chain 96, reverse output sheave 80, synchronizer 68, shaft 16, pinion carrier 108, and output shaft 14.