Abstract:
The present invention relates to a transmission unit for a vehicle driven by muscle force, comprising an input shaft which can be connected to cranks on opposite sides for driving the vehicle, comprising a first partial transmission. The first partial transmission comprises a countershaft, wherein a plurality of driving gear wheels is mounted on the input shaft, and wherein a corresponding plurality of driven gear wheels of the first partial transmission is mounted on the partial transmission shaft. The driven gear wheels of the first partial transmission are designed as idler gears that can be connected to the countershaft by means of shifting means in a rotationally rigid manner. The countershaft forms an input shaft of a second partial transmission. A plurality of second drive wheels are mounted on the output shaft, wherein the second partial transmission comprises an output shaft, on which a corresponding plurality of second driven gear wheels are mounted.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of international Patent Application No. PCT/EP 2009/009193, filed Dec. 21, 2009, the complete contents of which are hereby incorporated by reference. 
    
    
     BACKGROUND OF THE INVENTION 
     The invention relates to a shifting device for a transmission unit of a vehicle operated by muscle force. 
     The invention also relates to a transmission unit for a vehicle operated muscle force. 
     The invention additionally relates to a transmission housing for a transmission unit of a vehicle driven by muscle force. 
     Such transmission units serve to step up or step down the muscle force and as a result the provision of driving force to the vehicle. 
     Basically, there are three types of gearshifts for vehicles or motorbikes which are driven by muscle force, specifically derailleur gearshifts, hub gearshifts and bicycle gearshifts. 
     The derailleur gearshift has not changed significantly in the last decades. In this context, a chain transmits the driving force from a foot pedal to the rear axle of the bicycle, wherein a sprocket cassette which is mounted on the rear axle is mounted with up to ten sprockets, between which sprockets it is possible to shift to and fro by means of a shifting mechanism for guiding the chain, which mechanism is attached to the frame. Furthermore, most bicycles are additionally equipped with a shifting means on the chainwheel of the bottom bracket. Here, up to three chainwheels are attached to the foot pedal, and it is possible to shift to and fro between said chainwheels by means of a derailer which is attached to the frame. Such derailer gearshifts offer up to thirty gearspeeds, but due to the system a large number of gearspeeds are redundant, and some gearspeeds cannot be used, or can only be used to a restricted degree, as a result of high frictional losses due to a diagonal course of the chain. 
     A disadvantage with the principle of the derailer gearshift is not only the large number of redundant gearspeeds and the frictional losses but also the fact that the components are exposed and are therefore subject directly to environmental influences such as water and dirt and can very easily be damaged by shocks. 
     The second type of commercially available bicycle gearshift is the hub gearshift. In contrast to the derailer gearshift, this is understood to be a gearshift which is installed in the hub housing of the rear axle. A hub gearshift usually does not have any shifting components which are located on the outside and is therefore insensitive to shocks and subject to a lesser degree to the environmental influences than the derailer gearshift. A hub gearshift such as is known, for example, from DE 197 20 794 A1 can currently implement up to 14 gearspeeds. A disadvantage with the principle of the hub gearshift in the rear axle is that the weight of the rotating masses is increased and, in the case of bicycles which have rear wheel suspension, the non-suspended mass is increased considered relative to the overall weight. In addition, the center of gravity of the bicycle is displaced in the direction of the rear axle, which has an unfavorable effect on the riding properties of the bicycle, in particular in the case of mountain bikes with rear wheel suspension. 
     Such a hub gearshift is known, for example, from EP 0 383 350 B1, in which two planetary gear mechanisms are arranged coaxially with respect to a hub which is fixed to the housing, wherein the input shaft can be connected to planetary carriers, and the sun gears of the planetary gear mechanisms can be connected in a rotationally fixed fashion via a rotatable shifting device to the hub which is fixed to the housing, in order to implement different transmission ratios of the transmission as a whole. A disadvantage with this transmission is that the transmission as a whole is complicated and accordingly is, on the one hand, costly to manufacture and, on the other hand, has, due to the large number of components, a large weight with, at the same time, small number of gearspeeds which can be implemented. 
     The third variant of the bicycle gearshift is the bicycle gear mechanism or bottom bracket gear mechanism which is mounted in the region of the bottom bracket. This type of bicycle gearshift is not widespread, or is only found rarely, in commercially available bicycles. In general, such bicycle gear mechanisms have the advantage over conventional derailer gearshifts or hub gearshifts that they do not have any exposed components and accordingly are protected against shocks and environmental influences, and on the other hand they displace the center of gravity of the bicycle into the center, in which case at the same time the sum of the non-suspended mass is reduced. This is of particular advantage in the sport of mounting biking. A technical requirement in such bicycle gear mechanisms is to implement a compact design accompanied at the same time by a large number of shiftable gearspeeds. 
     U.S. Pat. No. 5,924,950 A discloses a bicycle gear mechanism having an input shaft on which a multiplicity of driving gears are mounted, and a countershaft on which a corresponding number of shiftable driven gears is mounted. The shiftable gears of the countershaft are shifted by means of a plurality of axially displaceable shifting pins and freewheels which are arranged in the countershaft, wherein the countershaft is connected via a planetary gear mechanism to a sprocket as an output element of the bicycle gear mechanism. The sprocket is connected via a clutch to the sun gear of the planetary gear mechanism, and the ring gear of the planetary gear mechanism can be braked by means of a Bowden cable. Fourteen gearspeeds can be implemented by means of this bicycle gear mechanism. A disadvantage of this system is the axially large design and the comparatively small number of fourteen implementable gearspeeds. 
     In addition, WO 2008/089932 A1 discloses a transmission unit for bicycles, in which transmission units a large number of gearspeeds can be implemented by means of two countershafts and a further partial transmission by multiplying the individual gearspeeds of the two partial transmissions, and at the same time a compact design can be implemented. A disadvantage with this transmission unit is that, in order to shift the idler gears, a camshaft is displaced axially and the transmission therefore requires a large amount of space in the axial direction. 
     BRIEF SUMMARY OF THE INVENTION 
     The object of the present invention is therefore to provide an improved transmission unit, an improved shifting device and an improved transmission housing for a vehicle operated by muscle force, as a result of which a more compact design and a large number of gearspeeds accompanied by a reduced weight can be implemented. 
     According to a first aspect of the invention there is provided a shifting device for a transmission unit of a vehicle operated by muscle force, having a first shaft on which a plurality of idler gears is mounted, a corresponding number of gearwheels, which are mounted on at least one second shaft, wherein the idler gears can each be connected to the first shaft by means of shifting means, wherein the first shaft is formed as a hollow shaft and has two shifting pins which lie coaxially on the inside, wherein the shifting pins are each connected to drive means which are configured to rotate the shifting pins in order to actuate the shifting means. 
     According to a second aspect of the invention there is provided a transmission unit for a vehicle which is driven by muscle force, having an input shaft which can be connected on opposite sides to cranks for driving the vehicle, having a first partial transmission, wherein the first partial transmission has a countershaft, wherein a plurality of driving gearwheels is mounted on the input shaft, and wherein a corresponding plurality of driven gearwheels of the first partial transmission is mounted on the countershaft, wherein the driven gearwheels of the first partial transmission are formed as idler gears which can be connected rotationally fixed to the countershaft by means of shifting means, wherein the countershaft forms an input shaft of a second partial transmission, on which input shaft a plurality of second driving gears are mounted, wherein the second partial transmission has an output shaft on which a corresponding plurality of second driven gearwheels is mounted, wherein the second driving gears of the second partial transmission are formed as idler gears which can be connected rotationally fixed to the input shaft by means of shifting means, and wherein the output shaft of the second partial transmission is formed as a hollow shaft which is arranged coaxially with respect to the input shaft. 
     According to a third aspect of the invention there is provided a transmission unit for a vehicle driven by muscle force, having a first partial transmission, wherein the first partial transmission has an input shaft and a countershaft, wherein a plurality of driving gears is mounted on the input shaft, and wherein a corresponding plurality of driven gears is mounted on the countershaft, wherein the countershaft of the first partial transmission can be connected to an input shaft of a second partial transmission, on which input shaft a plurality of second driving gears is mounted, wherein the second partial transmission has an output shaft on which a corresponding plurality of driven gearwheels is mounted, and wherein the countershaft of the first partial transmission can be connected to the input shaft of the second partial transmission by means of at least one epicyclic transmission. 
     According to a fourth aspect of the invention there is provided a transmission housing for a transmission unit of a vehicle operated by muscle force, having a housing casing which forms a circumferential face of the transmission housing, and a transmission cage for mounting the transmission unit, which transmission cage has pins for axially connecting at least two bearing plates, wherein at least one of the bearing plates is formed in such a way that in the assembled state of the transmission housing it substantially closes off the housing casing at one axial end. 
     One advantage of the first aspect of the invention is that the shifting pins permit an axially compact design of the transmission unit because the shifting pins are rotated in order to actuate the shifting means in the second shaft. 
     One advantage of the second aspect of the transmission unit according to the invention is that a compact design and at the same time a large number of shiftable gearspeeds can be implemented because the two partial transmissions are connected one behind the other and have a common shaft on which gearwheels of both partial transmissions are mounted. In addition, the compact design is implemented by the output shaft which is arranged coaxially with respect to the through shaft, wherein the through shaft is used simultaneously as an input shaft of the first partial transmission, and the output shaft is used simultaneously as an output shaft of the second partial transmission. 
     An advantage of the third aspect of the transmission unit according to the invention is that by virtue of a simple epicyclic transmission requiring little space, a further partial transmission for the transmission unit can be implemented and as a result the number of gearspeeds which can be implemented is at least doubled. 
     An advantage of the fourth aspect of the transmission unit according to the invention is that at least one of the bearing plates serves simultaneously as a housing cover and closes off the transmission housing axially. As a result, it is possible to dispense with at least one transmission housing cover, as a result of which the weight of the transmission unit can be advantageously reduced and the installation space made smaller. 
     In the first aspect of the invention it is preferred if the drive means are configured to rotate the corresponding shifting pin synchronously with respect to the first shaft in order to maintain a shifted state and to rotate the shifting pin in relation to the first shaft in order to carry out a gear change. 
     As a result, the shifting pin can actuate the shifting means by a relative movement, and individual idler gears can be connected rotationally fixed to the first shaft. 
     In addition it is advantageous if the two shifting pins can be rotated independently of one another. As a result, the idler gears can be shifted independently of one another, as a result of which a large number of gearspeeds can be implemented. 
     Furthermore, it is advantageous if the drive means each have a rotational speed super-imposition transmission. 
     As a result, the shifting pin can rotate synchronously with the first shaft, and a second rotational movement can easily be superimposed. 
     It is advantageous here if the rotational speed super-imposition transmission is formed as a planetary gear mechanism. 
     As a result, a simple and compact design of the rotational speed super-imposition mechanism is possible. 
     The planetary gear mechanism is preferably formed as a stepped planetary gear mechanism. As a result, a particularly compact design is possible. 
     It is preferred here if the first shaft is connected rotationally fixed to a sun gear of the planetary gear mechanism, and the shifting pin is connected rotationally fixed to a planetary carrier of the planetary gear mechanism. 
     As a result, the rotation of the first shaft can be transmitted with little expenditure to the shifting pin. 
     It is also advantageous if rotation of a ring gear of the planetary gear mechanism can be transmitted as rotation of the shifting pin in relation to the shaft. 
     As a result, a rotation in relation to the shaft can be transmitted to the shifting pin with simple mechanical means. 
     In general it is preferred if the rotation of the first shaft can be transmitted to the sun gear of the planetary gear mechanism by means of a constant gear set. 
     Furthermore, it is preferred if the rotation of the planetary carrier of the planetary gear mechanism can be transmitted to the shifting pin by means of a constant gear set. 
     As a result of this arrangement, the complicated relative rotation of the shifting pin in the first shaft can be implemented with mechanically simple means. The constant gear sets and the planetary gear mechanism are preferably configured in such a way that the shifting pin and the first shaft rotate at the same rotational speed if the ring gear of the planetary gear mechanism is secured or held in relation to the transmission housing. 
     The rotational movement of the ring gear is preferably carried out by means of a tension disk. The tension disk is preferably actuated by means of a Bowden cable. The tension disk translates the tensile movement carried out by the Bowden cable in a rotational movement of the ring gear. 
     The shifting means are preferably formed as shiftable freewheels. 
     As a result, the idler gears can be mounted on the first shaft in such a way that they can be shifted with simple and compact shifting means. 
     It is preferred here if the freewheels have shifting pawls which can be engaged with an internal toothing of the idler gears. 
     As a result, it is possible to implement freewheels which can easily be activated and which can take up a large torque because they transmit force in the tangential direction from the idler gear to the first shaft. 
     In addition it is preferred if the shifting pin has actuation portions by means of which the freewheels can be actuated. 
     As a result, the freewheels can be engaged with the idler gears by means of a structurally simple measure. 
     It is preferred here if the actuation portions are formed as recesses in the shifting pin. 
     As a result, actuation portions of the shifting pawls can pivot out in the inward direction, with the result that the actual shifting pawl pivots radially outward and can be engaged with the internal toothing of the idler gears. As a result, a shaft with a small diameter can be implemented. 
     Alternatively it is preferred if the actuation portions are of proud design. 
     As a result, the shifting pawls can be pressed directly radially outward, and it is possible to dispense with pretensioning devices such as, for example, springs. 
     Furthermore it is preferred if the actuation portions are arranged on the shifting pin in such a way that the freewheels of two successive gear stages can be activated simultaneously. 
     As a result it is possible to implement a power shift transmission because the freewheel of the relatively high gearspeed is engaged with the corresponding idler gear, while the freewheel of the relatively low gearspeed freewheels. 
     Furthermore it is preferred if a plurality of freewheeling pawls is assigned to each idler gear. 
     As a result, a relatively high torque can be transmitted from the idler gear to the shaft, and there is no risk of injury in the case of a fracture in a freewheeling pawl because at least one further freewheeling pawl can transmit the torque for a brief time. 
     Furthermore, it is preferred if the freewheeling pawls of a freewheel are distributed over the circumference of the shaft in such a way that just one of the freewheeling pawls can be engaged simultaneously with the idler gear. 
     As a result, the rotational angle of the shiftable idler gear decreases until the shifting pawl engages in the internal toothing, which improves the traveling comfort. 
     Furthermore it is preferred if the actuation portions are formed in such a way that only correspondingly shaped freewheels can be actuated. 
     As a result, specific actuation portions can only actuate specific freewheels, as a result of which a relatively large number of different freewheels can be actuated by means of one shifting pin. 
     Generally it is preferred if the shifting pin is of axially displaceable design. 
     As a result, the useable rotational range of the shifting pin can be enlarged. 
     Alternatively it is preferred that the drive means have an electric actuator. 
     As a result, it is possible to dispense with gearwheels for driving the shifting pin, as a result of which the weight can be reduced further. 
     In this context it is preferred if a stator of the electric actuator is connected rotationally fixed to the first shaft. 
     As a result, it is possible to dispense with additional rotational means in order to rotate the electric actuator synchronously with the first shaft. 
     In this context it is preferred if the electric actuator is arranged in the first shaft. 
     This makes a compact design of the transmission unit possible. 
     Furthermore it is preferred if the electric actuator is formed as an electric motor and particularly preferably as a stepping motor. 
     As a result, a relative movement of the shifting pin with respect to the second shaft can be implemented with a simple controller. 
     Alternatively it is preferred if the drive means have a hydraulic actuator. 
     As a result, the shifting pin can be actuated without additional electrical energy having to be made available. 
     In this context it is preferred if the hydraulic actuator has a hydraulic master which is connected rotationally fixed to a transmission cage. 
     As a result, hydraulic pressure can be fed to the hydraulic actuator without complex rotational feedthroughs. 
     In addition it is preferred if the hydraulic actuator has a hydraulic slave which is mounted so as to be capable of rotating in relation to the hydraulic master. 
     As a result, the hydraulic pressure can easily be transmitted to rotating components. 
     In this context it is preferred if the hydraulic slave has a first slave component which is connected rotationally fixed to the first shaft. 
     As a result, part of the hydraulic slave can be supported on the first shaft and can rotate the shifting pin with the rotational speed of the first shaft. 
     Furthermore it is preferred if the hydraulic slave has a second slave component which is connected rotationally fixed to the shifting pin. 
     As a result, the second slave component can transmit a rotational movement of the shifting pin in relation to the first shaft. 
     Furthermore, it is preferred if the first and second slave components form at least one slave cylinder, and wherein the second slave component forms at least one slave piston. 
     As a result, hydraulic pressure, which is built up in the slave cylinder, can actuate the slave piston and can bring about a relative rotation of the shifting pin in relation to the second shaft. 
     It is preferred here if the slave piston is mounted so as to be moveable in a circumferential direction. 
     As a result, no further mechanical means are necessary to translate the movement of the slave piston into a rotational movement. 
     Furthermore it is preferred if the slave cylinder is formed as a double-acting cylinder. 
     As a result, by simply reversing the applied hydraulic pressure it is possible to reverse the rotational direction of the slave piston. 
     Furthermore it is preferred if the first slave component and the second slave component form at least two slave pistons. 
     As a result, a greater force can be applied to the shifting pin. 
     It is preferred here if the two slave cylinders form a series circuit. It is possible that at first one piston is actuated and moves the shifting pin into a first rotational position, and then the second slave piston is activated by increasing the hydraulic pressure further, in order to rotate the shifting pin into a second rotational position. In addition, the useable rotational range of the shifting pin can be enlarged. 
     Furthermore it is preferred if the shifting pin has a latching device which secures the shifting pin in different rotational positions in the first shaft. 
     As a result it is possible to implement in a reproducible fashion a precise rotational position of the shifting pin in the first shaft, which position has to be released by an increased application of force. 
     Furthermore, it is preferred if the shifting pin is mounted so as to be axially moveable in relation to the second slave component. 
     As a result, a further shifting function can be implemented by means of axial movement of the shifting pin. 
     In addition it is preferred if the shifting pin has an actuation portion by means of which a clutch of a separate partial transmission can be actuated. 
     As a result, two transmissions which are connected in series can be shifted with one shifting pin, which improves the shifting comfort. 
     In the third aspect of the invention it is preferred if the countershaft of the first partial transmission can be connected rotationally fixed to the input shaft of the second partial transmission by means of a clutch at least in one rotational direction. 
     As a result, the first gear stage of the epicyclic transmission can be implemented with simple mechanical means. 
     Furthermore, it is preferred if the countershaft of the first partial transmission and the input shaft of the second partial transmission are arranged coaxially with respect to one another. 
     As a result, it is possible to transmit force from the countershaft to the input shaft of the second partial transmission with low structural complexity. 
     In this context it is also advantageous if the countershaft of the first partial transmission can be connected rotationally fixed to a planetary gear set of the epicyclic transmission. 
     As a result, a further gear stage can be implemented with structurally simple means. 
     Furthermore it is preferred if a ring gear of the epicyclic transmission can be connected rotationally fixed to the input shaft of the second partial transmission. 
     As a result, an output of the planetary gear mechanism can be implemented with simple means. 
     Furthermore, it is preferred if a sun gear of the epicyclic transmission can be connected rotationally fixed to a transmission housing of the transmission unit by means of a clutch. 
     As a result, a further gearspeed of the epicyclic transmission can be implemented with simple structural means. 
     Furthermore it is preferred if the input shaft of the first partial transmission is formed as a through shaft which can be connected on opposite sides to cranks for driving the vehicle. 
     As a result, it is possible to dispense with further gearwheels which transmit the input torque of the transmission unit to the input shaft of the first partial transmission. 
     Furthermore it is preferred if the output shaft is formed as a hollow shaft which is arranged coaxially with respect to the through shaft. 
     As a result, a compact design is possible and as a result the sprocket can be arranged in a compact design coaxially with respect to the bottom bracket of the bicycle. 
     Furthermore it is preferred if the epicyclic transmission rotates about a rotational axis which is arranged offset in parallel with the input shaft. 
     As a result, an axially compact design of the transmission unit can be implemented. 
     It is generally preferred if the input shaft can be connected rotationally fixed to the output shaft by means of a clutch at least in one rotational direction. As a result, a further gearspeed of the transmission unit can be implemented as a direct gearspeed with simple means. 
     Furthermore it is preferred if the second driving gears of the second partial transmission are formed as idler gears which can be connected rotationally fixed to the input shaft by means of shifting means. 
     As a result, the second partial transmission can be shifted with low structural complexity if the countershaft is not connected at its axial ends to drive means, for example foot pedals. 
     In addition, in the third aspect of the invention it is preferred if the driven gearwheels of the first partial transmission are formed as idler gears which can be connected rotationally fixed to the countershaft by means of shifting means. 
     As a result, the first partial transmission can be shifted with low structural complexity because the axial ends of the countershaft are accessible and are not connected to drive means, for example foot pedals. 
     It is generally preferred to combine one of the transmission units according to the invention with the shifting device according to the invention. This makes it possible to implement a bicycle gear mechanism which is compact overall. 
     Of course, the features which are mentioned above and the features which will be explained below can be used not only in the respectively specified combination but also in other combinations or alone without departing from the scope of the present invention. 
    
    
     
       BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS 
       Exemplary embodiments of the invention are illustrated in the drawing and will be explained in more detail in the following description. In the drawing: 
         FIG. 1  shows a side view of a bicycle frame with a multi-gearspeed transmission; 
         FIG. 2  shows an exploded illustration of a transmission housing with a multi-gearspeed transmission; 
         FIG. 3  shows a shifting diagram of a multi-gearspeed transmission with two partial transmissions and a common countershaft; 
         FIG. 4  shows a shifting diagram of a shifting device with a rotatable shifting pin; 
         FIG. 5  shows a shifting diagram of a shifting device with two rotatable shifting pins; 
         FIG. 6  shows a shifting diagram of a transmission unit with a planetary gear mechanism; 
         FIG. 7  shows a perspective illustration of an embodiment of a transmission unit with two partial transmissions and a common countershaft; 
         FIG. 8  shows a perspective illustration of an idler gear with internal toothing; 
         FIG. 9  shows a perspective illustration of a shifting pawl; 
         FIG. 10  shows a perspective exploded illustration of a shaft for bearing shiftable idler gears with a rotating shifting pin; 
         FIG. 11  shows a perspective illustration of a shaft with shifting pawls and a rotatable shifting pin; 
         FIGS. 12A-F  show basic outlines explaining shifting processes with a rotatable shifting pin; 
         FIG. 13  shows an exploded illustration of a shaft with a rotatable shifting pin and a planetary gear mechanism for rotating the shifting pin; 
         FIG. 14  shows a perspective illustration of the shaft and of the shifting device according to  FIG. 13 ; 
         FIG. 15  shows an exploded illustration of a rotatable shifting pin with a clutch of the planetary gear mechanism; 
         FIG. 16  shows an exploded illustration of the planetary gear mechanism with a shifting fork; 
         FIG. 17  shows a perspective illustration of the shifting means with a rotatable shifting pin and a shiftable planetary gear mechanism; 
         FIG. 18  shows a schematic side view of the shifting device with a rotatable shifting pin and a shiftable planetary gear mechanism; 
         FIG. 19  shows a schematic sectional illustration of a transmission unit with a rotatable shifting pin and a planetary gear mechanism; 
         FIGS. 20A-C  show basic outlines explaining the method of functioning of two hydraulic cylinders which are connected in series; 
         FIG. 21  shows a basic outline explaining a double-acting hydraulic cylinder; 
         FIG. 22  shows an exploded illustration of a hydraulic unit for activating the rotatable shifting pin; 
         FIG. 23  shows a schematic sectional view of a shaft with a shifting pin and a hydraulic unit according to  FIG. 22 ; 
         FIG. 24  shows a perspective illustration of the hydraulic unit for activating the rotatable shifting pin; 
         FIG. 25  shows a sectional view of the shifting pin from  FIG. 23 , showing the section along the line B-B; 
         FIG. 26  shows a sectional view of a shifting pin with a hydraulic unit showing the section along the line C-C from  FIG. 25 ; 
         FIG. 27  shows a sectional view of a latching device of the shifting pin showing a section along the line A-A from  FIG. 23 ; 
         FIG. 28  shows an exploded illustration of a transmission housing for holding a transmission unit; 
         FIG. 29  shows a circuit diagram of a shifting device with two rotatable shifting pins; and 
         FIG. 30  shows an exploded illustration of the shifting device according to  FIG. 29 . 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     In  FIG. 1 , a transmission unit is denoted generally by  10 . 
       FIG. 1  shows a side view of a bicycle frame  12  which has a transmission housing  14  in which the transmission unit  10  is held. The transmission unit  10  is only indicated schematically in this illustration and is formed as a compact unit which is preferably arranged in a transmission cage (not illustrated here). The transmission unit  10  is described here by way of example for the use in a bicycle, but the use in other vehicles which are operated by muscle force is also possible. In addition, it is also conceivable to use the transmission unit  10  for vehicles in which muscle force is used in combination with a drive machine for driving the vehicle. 
     The transmission unit  10  and the transmission housing  14  form, together with foot pedals  16  and  16 ′, a multi-gearspeed transmission  18 . 
       FIG. 2  shows an exploded illustration of the multi-gearspeed transmission  18 . Identical components are provided with identical reference symbols, and in this respect reference is therefore made to the description relating to  FIG. 1 . 
     The multi-gearspeed transmission  18  has a transmission housing  20  which is formed by a housing casing  22  and two housing covers  24 ,  26 , which close off the housing casing  22  at its axial ends. The multi-gearspeed transmission  18  also has a chainwheel  28 , which transmits, by means of a chain (not illustrated), a torque, which is stepped up or stepped down by means of the transmission unit  10 , to a rear wheel (not illustrated) of the bicycle. 
     The foot pedals  16 ,  16 ′ can be connected to an input shaft  30  of the transmission unit  10  and form the torque input for the multi-gearspeed transmission  18 . The chainwheel  28  is connected to an output shaft  32  of the transmission unit  10  and forms the output of the multi-gearspeed transmission  18 . The input shaft  30  and the output shaft  32  are arranged coaxially with respect to one another. 
     A transmission cage  34  is preferably arranged in the transmission housing  20 . The transmission cage  34  serves to hold a plurality of transmission shafts, bearings, shifting means, gearwheels and feed lines as well as other components of the multi-gearspeed transmission  18 . 
     The transmission cage  20  preferably has two bearing plates  36 ,  38  which can be connected to one another by means of a multiplicity of pins  40 . The bearing plates have bearings on which shafts are rotatably mounted. Gearwheels of the transmission unit  10  are mounted on the shafts. 
     Alternatively, the pins  40  and the shafts of the transmission unit  10  can be mounted on the housing covers  24 ,  26 , and it is therefore possible to dispense with separate bearing plates  36 ,  38  in order to safe weight and space. 
       FIG. 3  shows a circuit diagram of the transmission unit  10 . 
     The transmission unit  10  has the input shaft  30  and the output shaft  32 . The input shaft  30  is formed as a through shaft. The output shaft  32  is formed as a hollow shaft. The input shaft  30  and the output shaft  32  are arranged coaxially with respect to one another. The output shaft  32  is connected in a rotationally fixed fashion to the chainwheel  28  which forms an output element of the transmission unit  10 . 
     The transmission unit  10  has a first partial transmission  42  and a second partial transmission  44 . A multiplicity of driving gears  46 ,  47 ,  48 ,  49 ,  50 ,  51  are mounted on the input shaft  30 . The first partial transmission  42  has a countershaft  52 . Driven gears  53 ,  54 ,  55 ,  56 ,  57 ,  58  are mounted on the countershaft  52 . The driven gears  53 ,  54 ,  55 ,  56 ,  57 ,  58  are formed as idler gears. 
     The driven gears  53  to  58  can be connected to the countershaft  52  by means of shifting means (not illustrated). The driven gears  53  to  58  and the driving gears  46  to  51  form gear pairs which have different transmission ratios, and by selective connection of the driven gears  53  to  58  to the countershaft  52  it is therefore possible to implement different gear stages. 
     The second partial transmission  44  has an input shaft  60 . Driving gears  62 ,  63 ,  64  are mounted on the input shaft  60 . The driving gears  62 ,  63 ,  64  are formed as idler gears. The driving gears  62 ,  63 ,  64  can be connected in a rotationally fixed fashion to the input shaft  60  by means of shifting means. The driven gears  66 ,  67 ,  68  are mounted on the output shaft  32 . The driven gears  66 ,  67 ,  68  are in meshing engagement with the driving gears  62 ,  63 ,  64 . 
     By means of the driven gears  66 ,  67 ,  68  and driving gears  62 ,  63 ,  64  which mesh with one another, gear pairs are formed which have different transmission ratios. The driving gears  62 ,  63 ,  64  can be connected in a rotationally fixed fashion to the input shaft  60  by means of shifting means (not illustrated), as a result of which different selectable gear stages of the second partial transmission  44  are formed. 
     The countershaft  52  of the first partial transmission  42  is connected in a rotationally fixed fashion to the input shaft of the second partial transmission  44 . The countershaft  52  is preferably formed in one piece with the input shaft  60 . 
     By virtue of the fact that the first partial transmission  42  is connected to the second partial transmission  44 , the possible gear stages which can be implemented in the first partial transmission  42  are multiplied by the gear stages of the second partial transmission  44 . As a result, eighteen gearspeeds can be implemented by means of the transmission unit  10  which is illustrated in  FIG. 3 . 
     Furthermore, it is conceivable that the input shaft  30  can be connected in a rotationally fixed fashion to the output shaft  32  by means of a clutch (not illustrated). As a result, a further gearspeed could be implemented as a direct gearspeed. 
     In  FIG. 4  a shifting device for rotating a rotatable shifting pin is denoted generally by  70 . The shifting device  70  serves to connect in a rotationally fixed fashion idler gears (not illustrated), mounted on a shaft  72 , to the shaft  72  by means of shifting means (not illustrated). The shifting device  70  has a shifting pin  74  which is mounted so as to be rotatable in a coaxial fashion in the shaft  72  which is formed as a hollow shaft. The shifting pin  74  is formed in such a way that specific shifting means are activated in a specific rotational position in relation to the shaft  72 , with the result that at least one of the idler gears is connected in a rotationally fixed fashion to the shaft  72  at least in one rotational direction. The shifting device  70  which is illustrated in  FIG. 4  serves generally either to maintain the rotational position of the shifting pin  74  in relation to the rotating shaft  72 , in order to maintain the engaged gear stage, or serves to change the rotational position in a targeted fashion in order to change the gear stage. 
     The shaft  72  is connected in a rotationally fixed fashion to a driving gear  76 . The driving gear  76  is connected in a rotationally fixed fashion to a driven gear  78  which is mounted on a secondary shaft  80 . The driving gear  76  and the driven gear  78  form a first transmission gear  82 . 
     The shifting device  70  also has a variable-ratio epicyclic transmission  84  or a summing gear mechanism  84 , which is preferably formed as a planetary gear mechanism  84 . The planetary gear mechanism  84  has a sun gear  86 , planetary gears  88  and a ring gear  90 . The sun gear  86  is connected in a rotationally fixed fashion to the driven gear  78  of the epicyclic transmission  82 . The planetary gears  88  are mounted by means of a planetary carrier  92 . The planetary gears  88  mesh with an internal toothing of the ring gear  90  and with an external toothing of the sun gear  86 . The ring gear is connected in a rotationally fixed fashion to a ring gear shaft  93 . The ring gear shaft  93  is connected to a tension disk  94 . The planetary carrier  92  is rotatably mounted and connected in a rotationally fixed fashion to an output shaft  96 . The secondary shaft  80  and the output shaft  96  are arranged coaxially with respect to one another. The sun gear  86  and the ring gear  90  are arranged coaxially with respect to the secondary shaft  80 . The secondary shaft  80  is arranged offset in parallel with the shaft  72 . The ring gear shaft  93  is arranged coaxially with respect to the secondary shaft  80 . The ring gear shaft  93  can alternatively also be arranged offset in parallel with the secondary shaft  80  and can mesh with an external toothing of the ring gear  90 . 
     The output shaft  96  is connected in a rotationally fixed fashion to the shifting pin  74  via a second transmission gear  98 . The epicyclic transmission  98  has a constant gear set, which is formed by a driving gear  100  and a driven gear  102 . The driving gear  100  is mounted in a rotationally fixed fashion on the output shaft  96 , and the driven gear  102  is connected in a rotationally fixed fashion to the shifting pin  74 . 
     The transmission ratio of the first transmission gear  82 , of the planetary gear mechanism  84  and of the second transmission gear  98  is selected such that an overall transmission ratio of these three partial transmissions which are connected in series of one is obtained if the ring gear is secured or held in relation to the transmission housing. In such a state, the shifting pin  74  rotates at the same rotational speed as the shaft  72  by virtue of the selected transmission ratio. Accordingly, the shifting pin  74  does not carry out any relative rotation with respect to the shaft  72 . A set shifted state is therefore maintained by virtue of the particular embodiment of the shifting pin  74  and of the shifting means. 
     If the ring gear  90  is rotated, this rotation of the ring gear is transmitted as a rotation of the shifting pin  74  in relation to the shaft  72 . Depending on the rotational direction of the ring gear  90 , the shifting pin  74  is rotated at a rotational speed which is faster or slower than the shaft  72 . If the ring gear  90  is secured again in relation to the transmission housing, the shifting pin  74  rotates at the same rotational speed as the shaft  72 . As a result, a rotation of the ring gear  90  through a specific rotational angle can bring about a rotation of the shifting pin  74  through a specific rotational angle in relation to the shaft  72 . 
     The ring gear  90  is connected to the tension disk  94  via the ring gear shaft  93 . The tension disk  94  is preferably connected to a Bowden cable (not illustrated) and transmits a pulling movement of the Bowden cable into a rotational movement of the ring gear shaft  93 . As a result of actuation of the Bowden cable, the shifting pin  74  can rotate in relation to the shaft  72  in order to bring about a specific rotational position of the shifting pin in relation to the shaft  72 . 
     The transmission gears  82 ,  98  can alternatively also be formed as chains, belts or toothed belts. 
     The tension disk  94  is preferably pre-loaded with a spring or a return spring (not illustrated). The spring is formed in such a way that when shifting in the direction of low gearspeeds it is tensioned. When shifting up, the ring gear is actuated by the spring and/or rotated. As a result, shifting up without application of force is possible. In the case of shifting down, the spring is tensioned by the force which is transmitted via the Bowden cable. 
     Alternatively, the tension disk  94  can also be formed without a spring. The shifting movement is then carried out by means of two Bowden cables. In this context, a first of the Bowden cables rotates the ring gear in a first direction, and a second of the Bowden cables rotates the ring gear in the second direction, in order to shift up or shift down. 
       FIG. 5  is a circuit diagram of a shifting device according to the principle of the shifting device  70  in  FIG. 4 . The shifting device in  FIG. 5  is generally denoted by  104 . Identical elements are denoted by identical reference numbers, only the differences being explained here. 
     In principle, the shifting device  104  is identical to the shifting device  70  from  FIG. 4 , wherein the shifting device  104  is configured to rotate two rotatable shifting pins in the shaft  72  independently of one another. 
     Shifting pins  74 ′ and  74 ″ are arranged in the shaft  72 . The shaft  72  is connected via the first transmission gear  82  to a secondary shaft  80 . The secondary shaft  80  is connected to the shifting pin  74 ′ via a planetary gear mechanism  84  and a second transmission gear  98 ′, wherein the functional principle is identical to that of the shifting device  70  from  FIG. 4 . 
     In contrast to the shifting device  70 , the secondary shaft  80  is additionally connected to a planetary gear mechanism  84 ″. The planetary gear mechanism  84 ″ is preferably identical to the planetary gear mechanism  84 ′. The planetary gear mechanism  84 ″ is connected to the shifting pin  74 ″ via a second transmission gear  98 ″. As in the shifting device  70 , the transmission ratios from the shaft  72  to the shifting pin  74 ′ and to the shifting pin  74 ″ are just one, provided that corresponding ring gears  90 ′ and  90 ″ are secured in relation to the transmission housing. 
     The ring gears  90 ′,  90 ″ can each be actuated by means of a tension disk  94 ′,  94 ″ via ring gear shafts  93 ′,  93 ″. The two rotatable shifting pins  74 ′,  74 ″ can be rotated in relation to the shaft  72  by means of the shifting device  104 , and shifting means (not illustrated) can therefore be activated independently of one another. 
     The shifting device  104  can be used, for example, to connect the idler gears of the partial transmissions  42  and  44  in  FIG. 3  in a rotationally fixed fashion independently of one another to the corresponding shafts in order to form two partial transmissions which are connected in series. 
     In  FIG. 6 , a circuit diagram of a transmission unit with three partial transmissions which are connected in series is illustrated and is denoted generally by  110 . 
     The transmission unit  110  is to a certain extent identical to the transmission unit  10  from  FIG. 3 . Identical elements are denoted by identical reference numbers, with only the differences being explained here. 
     The input shaft  30  forms the input shaft of a first partial transmission  112 . The first partial transmission  112  is essentially identical to the first partial transmission  42  from  FIG. 3 , with the first partial transmission  112  only having three different gear sets. 
     The countershaft  52  of the first partial transmission  112  is connected to an epicyclic transmission or a planetary gear mechanism  114 . The countershaft  52  is connected in a rotationally fixed fashion to an input shaft  116  of the planetary gear mechanism  114 . The countershaft  52  is more preferably formed in one piece with the input shaft  116 . An output shaft  118  of the planetary gear mechanism  114  is connected in a rotationally fixed fashion to the input shaft  60  of the second partial transmission  44 . The output shaft  118  is preferably formed in one piece with the input shaft  60 . The planetary gear mechanism  114  has a first clutch  120  by means of which the input shaft  116  can be connected in a rotationally fixed fashion to the output shaft  118 . The clutch  120  is preferably formed as a freewheel. 
     The planetary gear mechanism has a sun gear  122 . The sun gear  122  can be connected in a rotationally fixed fashion to the transmission housing  40  by means of a second clutch  124 . The planetary gear mechanism  114  also has planetary gears  126  which are mounted so as to be rotatable by means of a planetary carrier  128 . The planetary carrier  128  can be connected in a rotationally fixed fashion to the input shaft  116 . In addition, the planetary gear mechanism  114  has a ring gear  130  which can be connected in a rotationally fixed fashion to the output shaft  118 . 
     Three different transmission ratios can be set between the input shaft  116  and the output shaft  118  of the planetary gear mechanism  114 , and three gear stages can therefore be implemented. The first gear stage is formed by closing the first clutch  120 , and opening the second clutch  124 . As a result, the input shaft  116  is connected in a rotationally fixed fashion to the output shaft  118 . The first transmission ratio is consequently equal to 1. 
     The second gear stage is formed by opening the first clutch  120  and closing the second clutch  124 . As a result, the sun gear  122  is held tight and the rotating planetary carrier  128  drives the ring gear  130  which is connected to the output shaft  118 . The second transmission ratio is consequently a step-up transmission ratio. 
     The third gear stage is formed in that the second clutch  124  is closed and the sun gear is therefore held tight. In addition, the input shaft  116  is connected to the ring gear  130  by means of a further clutch. In addition, the planetary carrier  128  is connected to the output shaft  118 , and the planetary carrier  128  therefore forms the output of the planetary gear mechanism  114 . 
     As a result, in each case three shiftable gear stages are formed by the partial transmissions  112  and  44  and by the planetary gear mechanism  114 , and eighteen gear stages can be implemented by virtue of the fact that the three partial transmissions  112 ,  114 ,  44  are connected in series. 
     The partial transmissions  112 ,  44  are preferably shifted by means of the shifting device  104  in  FIG. 5 , wherein at least one of the shifting pins  74 ′,  74 ″ has shifting means which actuates at least one of the clutches  120 ,  124 . 
     Of course, in the case of the transmission unit  110  in  FIG. 6 , the input shaft  30  can also be connected to the output shaft  32  by a clutch in order to form a gearspeed, which is the nineteenth in this case. 
     In order to increase the number of gearspeeds it is also conceivable to embody the planetary gear mechanism  114  as a multi-stage planetary gear mechanism. 
       FIG. 7  shows a perspective illustration of the transmission unit  10 . The transmission unit corresponds to the circuit diagram according to  FIG. 3 , with identical elements being denoted by identical reference numbers and only the differences being explained here. 
     The countershaft  52  of the first partial transmission  42  is formed in one piece with the input shaft  60  of the second partial transmission  44 . The driven gears  53  to  58  and the driving gears  62  to  64  are formed as idler gears and can be shifted by means of the shifting pins  74 ′ and  74 ″. In addition, the driving gear  76  is mounted on the shaft  52  or  60  in order to drive the shifting pins  74 ′ and  74 ″ via the shifting device  104  (not illustrated). 
     In  FIG. 8  a shiftable idler gear with internal toothing is illustrated and is denoted generally by  132 . 
     The idler gear  132  has an external toothing  134  and an internal toothing  136 . The external toothing  134  is formed on the outer circumferential face. The internal toothing is formed on an inner circumferential face of the idler gear  132 . The internal toothing  136  has sliding portions  138  and engagement portions  140 . The sliding portions  138  are faces which are arranged in the circumferential direction of the idler gear  132 . The engagement portions  140  are formed between the sliding portions  138 , at an angle with respect to the sliding portions  138 . 
     The external toothing  134  serves to mesh with other gearwheels. The internal toothing  136  serves to mount the idler gear  132  on a shaft and to connect it in a rotationally fixed fashion to the shaft by means of shifting means. In this context, the sliding portions  138  serve to mount the idler gear  132  on the shaft and to slide on the shaft. The engagement portions  140  serve to ensure that shifting means (which are not illustrated and which will be explained in more detail below) can be placed in engagement with the idler gear  132  and to connect the idler gear  132  in a rotationally fixed fashion to the shaft. 
     In  FIG. 9  a freewheel body for connecting the idler gear  132  in a rotationally fixed fashion to a corresponding shaft is illustrated and denoted generally by  142 . The freewheel body  142  has an actuation portion  144  which is formed on an underside of the freewheel body  142 . The freewheel body  142  has a bearing portion  146  on each of its two lateral sections. The freewheel body  142  has an engagement portion  148 . The engagement portion  148  is formed at an end of the freewheel body  142  lying opposite the actuation portion  144 . The bearing sections  146  are formed on opposite sides of the freewheel body  142 , specifically between the actuation portion  144  and the engagement portion  148 . 
     The bearing portions  146  serve to mount the freewheel body  142  on a shaft in such a way that it can rotate or pivot about a rotational axis  150 . In this context, the freewheel body  142  is attached to or mounted on the shaft in such a way that the actuation portion  144  points toward the inside of the shaft. In addition, the freewheel body  142  is prestressed by means of a spring element in such a way that in the unloaded state the actuation portion  144  is pivoted in the inward direction and the engagement section  148  is pivoted radially outward. The actuation portion  144  serves to be pressed radially outward by means of the shifting pin  74  in order to pivot the engagement portion  148  radially inward about the rotational axis  150 . 
     If the engagement portion  148  is pivoted radially outward and protrudes with respect to the shaft, said engagement portion  148  can be placed in engagement with the engagement portion  140  of the internal toothing  136  of the idler gear  132  in one rotational direction of the idler gear  132 , and the idler gear can therefore be connected in a rotationally fixed fashion to the shaft in the rotational direction. 
     The freewheel body  142  also has a sliding portion  152 . The sliding portion  152  serves to pivot the freewheel body  142  radially inward if the idler gear is rotated in relation to the shaft in a direction which is opposed to the rotational direction, thereby serving as a freewheel. 
     The actuation portion  144  can have one or more grooves running perpendicular with respect to the rotational axis  150  or in the rotational direction of the shaft, in order to permit selective actuation. This is explained in more detail below. 
       FIG. 10  shows a shaft for mounting shiftable idler gears  132  and a shifting pin for shifting the freewheel bodies  142 , in an exploded illustration. The shaft is generally denoted by  154  and the shifting pin by  156 . The shaft  154  is designed as a hollow shaft to hold the shifting pin  156 . The shaft  154  has bearing portions  158 . Through holes  160  are formed in the region of the bearing portions  158 . The shaft  154  has a first group  161  of bearing portions  158  which are formed axially one next to the other. In addition, the shaft  154  has a second group  163  of bearing portions  158  which are formed axially one next to the other. The first group  161  of the bearing portions  158  is arranged offset with respect to the second group  163  of bearing portions  158  in the circumferential direction. In each case two of the bearing portions  158  are arranged on opposite sides of the shaft  154 . 
     The bearing portions  158  are formed in such a way that they can each hold one of the freewheel bodies  142 . The through holes  160  serve to allow the actuation portion  144  to pivot through the through holes  160  and to be actuated by the shifting pin  156 . The bearing portions  158  are formed in the shaft  154  in such a way that in a pivoted-in state the freewheel bodies  142  do not protrude with respect to the circumferential face of the shaft  154 . In this pivoted-in state, the circumferential face of the shaft  154  and the sliding section  152  of the freewheel bodies  142  essentially form a plane. 
     A pin hole  162 , through which a guide pin can be led, is also formed in the shaft  154 . 
     The shifting pin  156  has actuation portions  164  which are formed over the circumference of the shifting pin  156 . The actuation portions  164  are formed as recesses. The shifting pin  156  further has a circumferential groove  166 . The groove  166  has two circumferential sections which are axially offset and are connected to one another by an oblique section  167 . The actuation portions  164  are arranged axially offset and distributed over the circumference. The actuation portions  164  are arranged in part next to one another in the axial direction. The actuation portions  164  are arranged on opposite sides of the shifting pin  156 , specifically in a way corresponding to the bearing portions  158  in the shaft  154 . 
     The shifting pin  156  is formed in such a way that, depending on the rotational position of the shifting pin in the shaft  154 , the actuation portions  164  are positioned on one of the through holes  160 . As a result, the actuation portion  144  of the freewheel bodies  142  can pivot into the actuation portion  164  and therefore move the engagement section  148  into engagement with the internal toothing  136 . The shifting pin  156  serves to actuate the freewheel bodies  142  of the first group  161  of bearing portions  158 . In the inserted state of the shifting pin  156 , the groove  166  is arranged in the region of the pin hole  162 , with the result that the groove  166  can hold a pin (not illustrated) which is led through the pin hole  162 . As a result, the shifting pin  156  is moved into different axial positions depending on the rotational position in the shaft  154 . 
     This axial displacement of the shifting pin  156  serves to enlarge the useable rotational range of the shifting pin. The axial displacement has the effect that due to the axial position of some of the actuation portions  164  in relation to the actuation portions  144 , said actuation portions  164  cannot activate the freewheel bodies  142 . This means, conversely, that specific actuation portions  144  can be activated by specific freewheel bodies  142  in the second axial position. Consequently, as a result of the axial displacement of the shifting pin  156 , some of the actuation portions  164  are not arranged under the actuation portions  144  and consequently cannot activate the freewheel bodies  142 . As a result, when there are shifting pawls lying opposite one another an additional useable rotational range of 180° is produced for further shiftable gearwheels. It is also conceivable to enlarge the useable rotational range of the shifting pin  156  further by even further axial displacement. 
     A further possible way of extending the useable rotational range of the shifting pin is to have a different configuration of the actuation portions  144 . By virtue of an asymmetrical configuration of the actuation portions  144  and corresponding actuation portions  164 , only specific freewheel bodies  142  are actuated or only specific actuation portions  144  are pivoted into specific actuation sections  164  of the shifting pin  156 . As a result, the useable rotational range of the shifting pin  156  can be extended from 180° to 360° even when there are freewheel bodies lying opposite one another. 
     For example, the actuation portions  144  of the freewheel pawls  142  can have one or more grooves running in the rotational direction of the shaft  154 , with the result that actuation portions  144  which are configured in such a way can only pivot into correspondingly configured actuation portions  164 . In this context, selective actuation can be made possible by means of the number and the position of such grooves. 
       FIG. 11  illustrates the shaft  154  with the inserted shifting pin  156  and the freewheel bodies  142 . Identical elements are provided with identical reference numbers, with only the differences being presented here. 
     The shifting pin  156  is positioned in the shaft  154  in such a way that two of the freewheel bodies  142  are pivoted out, with just one being visible. 
     In addition, a second shifting pin (which is not illustrated or cannot be seen) which actuates the second group  163  of freewheel bodies is inserted into the shaft  154 . This shifting pin is arranged in the shaft  154  in such a way that two freewheel bodies  142  of the second group  163  are pivoted out, with the result that the engagement section  148  can be placed in engagement with the engagement section  140  of the internal toothing  136  of the idler gear  132 . 
     Through selected rotational positions of the two shifting pins  156 , two idler gears  132  can be connected in a rotationally fixed fashion to the shaft  154 , with the result that one of the eighteen possible gear stages is shifted. 
       FIGS. 12A to 12F  show radial sectional views through adjacent idler gears  132 , during three phases of a gear change. 
       FIG. 12A  shows a first of the idler gears  132  whose internal toothing  136  is in engagement with the two assigned freewheel bodies  142 . The shifting pin  156  is in a rotational position in relation to the shaft  154 , with the result that the actuation portions  164  of the shifting pin  156  are arranged in the region of the actuation portions  144  of the freewheel bodies  142 , and the freewheel body  142  can therefore pivot outward. 
     The second of the idler gears  132 , which is assigned to a next highest gear stage, specifically of the second gearspeed, is shown in  FIG. 12B . The freewheel bodies  142  are pivoted in radially in the inward direction, and are consequently not in engagement with the internal toothing  136  of the idler gear  132 . In the rotational position of the shifting pin  156 , the actuation portions  164 , which are assigned to the second gearspeed, are not arranged under the actuation portions  144  of the freewheel bodies  142 , with the result that the actuation portions  144  are pressed outward. 
     If the shifting pin  156  is rotated, as indicated by an arrow  168 , the actuation portion  164  remains underneath the freewheel body  142 , which is assigned to the first of the idler gears  132  and therefore to the first gearspeed, as is illustrated in  FIG. 12C , and the freewheel bodies  142  of the first gearspeed therefore remain pivoted out toward the outside. 
       FIG. 12D  illustrates the second of the idler gears  132  in this rotational position of the shifting pin  156  which is assigned to the second gearspeed. In this rotational position of the shifting pin  156 , the actuation portion  164 , which is assigned to the second gearspeed, is arranged radially underneath the actuation portion  144  of the second gearspeed, with the result that the actuation portion  144  can pivot in the radially inward direction and the engagement section  148  can therefore pivot out in the radially outward direction. As a result, the engagement section  148  can be placed in engagement with the internal toothing  136  of the idler gear  132 . The freewheel bodies  142  are each assigned a spring which prestresses the corresponding freewheel body  142  in such a way that the actuation portion  144  is pressed against the shifting pin  156 . As a result, the engagement section  148  pivots out if one of the actuation portion  164  is rotated under the shifting pawl  142 . 
     Since the higher gearspeed has a relatively low transmission ratio, the freewheel pawls  142  of the higher gearspeed engage in the internal toothing  136  and drive the shaft  154  with a rotational speed which is higher than the rotational speed of the idler gear  132  of the relatively low gearspeed. In this so-called intermediate state, the idler gear  132  of the relatively low gearspeed is therefore rotated in relation to the shaft  154  in the opposite direction. As a result, the sliding portion  138  of the idler gear  132  presses against the sliding portion  152  of the freewheel body  142 , with the result that the freewheel body  142  is pivoted out in the inward direction and the first of the idler gears  132  slides on the shaft  154 . The idler gear  132  of the relatively low gearspeed, that is to say of the first gearspeed, freewheels in the intermediate state. 
       FIGS. 12E and 12F  illustrate the state in which the second gearspeed is completely engaged. For this purpose, the shifting pin  156  has been rotated onward in the direction of the arrow  168 , with the result that the freewheel bodies  142  of the first gearspeed are pivoted in by the shifting pin  156 , as is shown in  FIG. 12E .  FIG. 12F  shows that the freewheel bodies  142  of the second gearspeed continue to be in engagement with the internal toothing  136  because the actuation portions  164  of the second gearspeed are arranged radially underneath the actuation portions  144  of the freewheel bodies  142 . 
     Switching under load is possible by means of the intermediate state in which the freewheel bodies  142  of two subsequent gearspeeds are pivoted out radially. In addition an idling state is avoided. 
     When shifting into a low gearspeed occurs, the sliding section  138  of the internal toothing  136  of the relatively low gearspeed firstly slides over the freewheel bodies  142  in the intermediate state. The relatively high gearspeed initially remains engaged. The freewheel bodies  142  are then pivoted in or disengaged only when the load which is transmitted to the shaft  154  via the idler gear  132  is removed. In addition, the shifting pin  156  must then be rotated onward with the result that the actuation portion  144  is pressed outward. The relatively low gearspeed is then engaged immediately because this gearspeed was already in the intermediate state or in the freewheeling state. This avoids an idling state. 
       FIG. 12  illustrates a shifting pin  156  with actuation portions  164  lying precisely opposite one another. Alternatively it is also conceivable for the actuation portions  164  to be arranged in relation to one another in such a way that only one of the shifting pawls is placed in engagement with the internal toothing  136 . This is implemented by virtue of the fact that the shifting pawls  142  on the shaft  154  are not arranged precisely opposite one another. As a result, the rotational angle of the idler gear  132  can be reduced in size until the actuation portion  148  latches into the internal toothing  136 . 
       FIG. 13  shows the shaft  154  and the shifting device  104  in a perspective exploded illustration. The illustration in  FIG. 13  corresponds to the circuit diagram in  FIG. 5 . Identical elements are denoted by identical reference numbers, with only the difference being explained here. 
     The tension disk  94 ′ is connected to the ring gear  90 ′ via the ring gear shaft  93 ′. The ring gear shaft  93 ′ is formed as a hollow shaft in order to accommodate the output shaft  96 ′. The ring gear  90 ′ is rotated by the tension disk  94 ′ in order to rotate the shifting pin  156  in relation to the shaft  154 . 
     The ring gear  90 ″ has, in addition to the internal toothing  136 , an external toothing  170 . The external toothing  170  serves to connect the ring gear  90 ″ to the tension disk  94 ″ (not illustrated here) via the ring gear shaft  93 ″ (not illustrated). 
       FIG. 14  shows a perspective illustration of the shaft  154  and of the shifting device  104  from  FIG. 13  in the assembled state. Identical elements are provided with identical reference numbers, with only the differences being presented here. 
     The external toothing  170  of the ring gear  90 ″ is connected in a rotationally fixed fashion to the ring gear shaft  93 ″ which is connected to the tension disk  94 ″. The ring gear shaft  93 ″ is arranged offset in parallel with the ring gear shaft  93 ′. The ring gear shaft  93 ″ is connected in a rotationally fixed fashion to a gearwheel  95  which meshes with the external toothing  170 . The driven gears  102 ′,  102 ″ are each connected via a deflection gearwheel  172 ′,  172 ″ to the driven gears  102 ′,  102 ″. The deflection gearwheels  172 ′,  172 ″ serve to reverse the rotational direction of the shifting pin  157 . 
     In order to permit a plurality of shiftable partial transmissions, for example partial transmissions  42  and  44 , to be controlled with just one shifting cable or the like, it is possible to control a plurality of partial transmissions in a combined fashion. For this purpose, for example the shifting pin  74 ′ can be formed in such a way that after the shifting pin  74 ′ rotates onward beyond the last or highest gearspeed of this partial transmission, the first gearspeed follows again. In addition, in order to solve this problem it would be necessary to make available a mechanism which, when the shifting pin  74 ′ rotates onward beyond the highest gearspeed, rotates the shifting pin  74 ″ by one shift position into the next highest gearspeed. This can be implemented by virtue of the fact that the ring gear shafts  93 ′,  93 ″ of the planetary gear mechanisms  84 ′,  84 ″ are shifted together. For example, the two ring gears  90 ′,  90 ″ can therefore be connected together. For example, the transmission unit  10  can be shifted from the sixth gearspeed into the seventh gearspeed by virtue of the fact that the shifting pin  74 ′, which is assigned to the partial transmission  42 , is rotated onward after one rotation through 360°, in order to shift the partial transmission  44  from the sixth gearspeed into the first gearspeed again. The planetary gear mechanism  84 ″, which is assigned to the partial transmission  44 , is configured in such a way that when further rotation occurs after the sixth gearspeed in partial transmission  42 , the second gearspeed is engaged after the first gearspeed in partial transmission  44 . As a result of the fact that the first gearspeed follows the sixth gearspeed in the partial transmission  42  and at the same time the second gearspeed follows the first gearspeed in the partial transmission  44 , the transmission unit can therefore be shifted from the sixth gearspeed into the seventh gearspeed. 
     The tension disk  94  is preferably connected to a shift lever via a cable pull. The tension disk is preferably prestressed with a spring with respect to the ring gear  90 . The tension disk  94  and the ring gear  90  preferably have stops in order to prestress the tension disk  94  and the ring gear  90  one against the other with a defined spring force. The spring is relaxed as a result of the rotation of the ring gear  90 , as a result of which the shifting pin  74  rotates and a gear change is carried out. If a gear change is carried out into a low gearspeed, the shift lever is firstly actuated, as a result of which the tension disk  94  is prestressed with respect to the ring gear  90  without the load of the transmission being reduced. Since under load the engagement sections  148  engage in the internal toothing  136  and latch in this position as a result of the transmitting torque, the shifting pin  156  cannot be rotated. As soon as the loading of the transmission drops, that is to say the rotational force is reduced, the shifting pin  74  can disengage the shifting pawl  142  and the internal toothing  136  owing to the spring prestress of the ring gear  90 . In this context, the oscillating pedaling force profile which is typical during cycling can be used since the pedaling force which is applied to the foot pedals  16 ,  16 ′ is greatly reduced in the vertical position of the foot pedals  16 ,  16 ′. 
     In such a position of the foot pedals  16 ,  16 ′, a prestressed or preselected low gearspeed can be completely engaged. 
     In general it is advantageous to configure the employed gearwheels in accordance with the torque to be transmitted or in accordance with their transmission ratio. In this context, gearwheels which have to transmit large tangential forces or large torques should be correspondingly wider, that is to say should be made thicker in the axial direction. In contrast, it is appropriate to embody gearwheels which have a small transmission ratio with a relatively short width since said gearwheels have to transmit relatively small tangential forces or relatively small torques. As a result, the installation space in the transmission housing can be optimized. In addition it is preferred for the shifting means which are formed by the freewheel pawls and the shifting pin to be configured in accordance with the tangential forces and the expected torques. In this context it is also conceivable to adapt the number of freewheel pawls  142  to the torques to be transmitted. 
     If the transmission unit  10  is additionally formed with the planetary gear mechanism  114 , as is illustrated in  FIG. 6 , the actuation of the partial transmission  112  has to be combined with the actuation of the planetary gear mechanism  114 . The shifting pin  74  then controls the clutch  120  of the planetary gear mechanism  114 . The ring gear  90  of the planetary gear mechanism  84  from  FIG. 4  or  FIG. 5  then additionally controls a shifting fork which actuates the clutch  124  of the planetary gear mechanism  114 . The method of functioning of this shifting control is explained in more detail below. 
       FIG. 15  shows an exploded illustration of a shifting pin for actuating a clutch of the planetary gear mechanism  114 . This embodiment of the shifting pin is denoted generally by  174 . The shifting pin  174  has the actuation portions  164 . The shifting pin  174  can be connected in a rotationally fixed fashion at an axial end to the driven gear  102 . At an axial end of the shifting pin which lies opposite, a groove  176  is formed in the shifting pin  174 . The groove  176  has two portions running in the circumferential direction, which portions are offset axially with respect to one another. The two circumferentially running portions are connected by an oblique section  178 . 
       FIG. 15  illustrates a spring  180  and an input element  182  of the clutch  120 . The input element  182  is assigned a pin  184  which can be introduced into a drill hole  186  in the input element  182 . 
     In the inserted state, the spring  180  is plugged onto the shifting pin  174 , and the pin  184  is introduced into the drill hole  186 , with the result that the pin  184  engages in the groove  176 . The input element  182  is axially pre-loaded with respect to the shifting pin  174  by the spring  180 . If the shifting pin  174  is rotated in relation to the input element  182  with the result that the pin  184  slides along the oblique section  178  of the groove  176 , the input element  182  is moved in the axial direction by a spring force of the spring  180  and is placed in engagement with an output element (not illustrated) of the clutch  120 . The clutch  120  of the planetary gear mechanism  114  can therefore be actuated by rotation of the shifting pin  174 . 
       FIG. 16  is an exploded illustration of the planetary gear mechanism  84  with a shifting fork for actuating the clutch  124 . Identical elements are denoted by identical reference numbers, with only the differences being illustrated here. 
     The ring gear shaft  93  has a groove  188  which has two portions running in the circumferential direction. The portions which run in the circumferential direction are offset axially with respect to one another and connect by an oblique section  190 . In addition, the shifting device from  FIG. 16  has a shifting fork  192  which has a sleeve section  194  and a fork section  196 . The sleeve section  194  has a drill hole  198  through which a pin  200  can be inserted. In addition, the shifting device has a spring  202  which is arranged between the driven gear  78  and the shifting fork  192 . In the assembled state, the sleeve section  194  is mounted in the region of the groove  188 , with the result that the pin  200  which is guided through the drill hole  198  engages in the groove  188 . The pin  202  is supported on a retaining ring  203  and pre-loads the shifting fork  192  axially. As a result, the pin  200  bears, in the groove  188 , against an edge on which the oblique section  190  is formed. 
     If a gearspeed is shifted by means of this shifting device, the ring gear shaft  93  is rotated through a specific rotational angle, as described above. If the rotation of the ring gear shaft  93  is formed in such a way that the pin  200  slides over the oblique section, the shifting fork is displaced axially depending on the rotational direction of the ring gear shaft  93 . As a result of this axial displacement, the clutch  124  is actuated, as is explained in more detail below. 
       FIG. 17  illustrates the shifting device according to  FIG. 16  in the assembled state with the shaft  52  and parts of the planetary gear mechanism  114 . Identical elements are denoted by identical reference numbers, with only the differences being explained here. 
     The ring gear shaft  93  has a gearwheel section  204 . The gearwheel section  204  is connected in a rotationally fixed fashion to a driving gearwheel  205  of the tension disk  94 . A defined rotation can be transmitted to the ring gear shaft  93  through the tension disk  94  and a gearwheel pair formed by the gearwheel section  204  and the driving gearwheel  205 . If the pin  200  slides over the oblique section  190  during this rotation, the shifting fork  192  (not illustrated in  FIG. 17 ) is moved in the axial direction. As a result, the clutch  124  can be actuated. 
       FIG. 18  is a schematic illustration of a side view of the transmission unit  110  with a shifting device. Identical elements are denoted by identical reference numbers, with only the differences being presented here. 
     As described above, the shifting fork  192  can be displaced axially by rotating the ring gear shaft  93 . The fork section  196  is connected to the input element  182  of the clutch  124 . If the fork section  196  is displaced in the axial direction, specifically in the direction of an arrow  205 , the input element  182  is placed in engagement with an output element  206  of the clutch  124 . The clutch  124  can therefore be actuated by actuating the tension disk  94 . As a result of the fact that the rotation of the shifting pin  174  is connected directly to the rotation of the ring gear shaft  93 , it is possible for the clutch  124  to be actuated when the assigned partial transmission is shifted onward from the highest gearspeed into the first gearspeed. 
     In  FIG. 19 , a schematic sectional view of the transmission unit  110  is illustrated as a section through the input shaft  30  and the countershaft  52 . 
       FIGS. 20   a - c  are schematic illustrations of a hydraulic system of hydraulic cylinders which are connected in series. The hydraulic system has a first hydraulic cylinder  208  and a second hydraulic cylinder  210 . A hydraulic piston  212 ,  214  is arranged in an axially moveable fashion in each of the hydraulic cylinders  208 ,  210 . The hydraulic cylinders  208 ,  210  each have a main opening  216 ,  218  and each have two secondary openings  220 ,  222 ,  224 ,  226 . The secondary opening  220  is connected to the secondary opening  226  via a duct  228 . The secondary opening  222  of the first hydraulic cylinder  208  is connected to the secondary opening  224  in the hydraulic cylinder  210  via a duct  230 . The secondary openings  220 ,  222 ,  224 ,  226  are each arranged opposite the main openings  216 ,  218 . 
     If hydraulic pressure is applied to the first hydraulic cylinder  208  through the main opening  216 , the hydraulic piston  212  moves in the direction of the secondary openings  220 ,  222 . As a result, hydraulic fluid is conducted through the secondary opening  222  and the duct  230  through the secondary opening  224  into the hydraulic cylinder  210 . Since the secondary opening  224  is arranged underneath the hydraulic piston  214 , the hydraulic fluid is forced into the hydraulic cylinder  210  without a force being applied to the hydraulic piston  214 . The hydraulic fluid leaves the hydraulic cylinder  210  through the main opening  218 . 
     In  FIG. 20   b , the hydraulic piston  212  has arrived at one end of the hydraulic cylinder  208 . In this position, the secondary opening  222  is closed off and the secondary opening  220  is opened, with the result that hydraulic fluid can be forced from the hydraulic cylinder  208  and through the duct  228 . The hydraulic pressure acts, in this position, on the hydraulic piston  214  through the secondary opening  226 . This application of pressure moves the hydraulic piston  214  in the direction of the main opening  218 . This is illustrated in  FIG. 20   c.    
     If hydraulic pressure is then applied to the second hydraulic cylinder  210  through the main opening  218 , the hydraulic piston  214  firstly moves in the direction of the secondary opening  226 . The hydraulic fluid is conducted through the duct  228  and into the hydraulic cylinder  208 , and is directed out of the hydraulic cylinder  208  though the main opening  216 . If the hydraulic piston  214  has arrived at the end of the hydraulic cylinder  210 , hydraulic pressure is applied to the hydraulic piston  212  through the duct  230  and the hydraulic piston  212  is moved in the direction of the main opening  216 . 
     Two hydraulic pistons can be moved one after the other through this series connection of two hydraulic cylinders. 
       FIG. 21  illustrates the principle of a double-acting cylinder.  FIG. 21  shows a hydraulic cylinder  232  which has an opening  234  and an opening  236 . The openings  234 ,  236  are arranged on opposite sides of the hydraulic cylinder  232 . An axially moveable hydraulic piston  238  is located between the openings  234 ,  236 . If hydraulic pressure is applied to the hydraulic cylinder  232  through the opening  234 , the hydraulic piston  238  moves in the direction of the opening  236 . Hydraulic oil is discharged from the hydraulic cylinder  232  through the opening  236 . In order to move the hydraulic piston  238  in an opposite direction, specifically in the direction of the opening  234 , hydraulic pressure is applied to the hydraulic cylinder  232  through the opening  236 . As a result, the hydraulic piston  238  moves in the direction of the opening  234 , through which hydraulic oil is discharged from the hydraulic cylinder  232 . 
       FIG. 22  is an exploded illustration of a shifting pin with a hydraulic drive system. The shifting pin is generally denoted by  240 . The hydraulic system is generally denoted by  242 . 
     The shifting pin  240  has the groove  166  into which the pin  184  can engage. A radial drill hole  244 , which is provided for accommodating a spring  246  and a ball  248 , is formed in the shifting pin  240 . The drill hole  244  forms a latching device together with the spring  246  and the ball  248 . 
     The actuation sections  164  are formed in the shifting pin  240 . The hydraulic drive system  242  has a hydraulic master  250 , a hydraulic slave  252  and an actuator  254  or a vane positioner  254 . The hydraulic master  250  has two hydraulic connections  256 ,  257 . The hydraulic connections are provided for being connected to hydraulic hoses and for supplying the hydraulic drive system  242  with hydraulic pressure. The hydraulic slave has a separating disk  258 . Two rotationally symmetrical connecting elements  260 ,  262  are formed on a side of the separating disk  258  facing the hydraulic master  250 . The connecting elements  260 ,  262  each have a groove  264 ,  266  which is formed in the circumferential direction. On the side of the separating disk  258  lying opposite the connecting elements  260 ,  262 , a cylindrical section  268  is formed with two slave vanes  270 ,  272  which protrude radially. The actuator has a cylindrical section  274  on which two actuator vanes  276 ,  278 , which protrude in the axial direction, are formed. In addition, the actuator  254  has a connecting section  280 . The connecting section  280  has a hexagonal profile. 
       FIG. 23  illustrates the shifting pin  240 , the shaft  154  and the hydraulic drive system  242  in an axial sectional view. In this illustration, the shifting pin  240  is mounted in the shaft  154 . 
     The hydraulic connections  256 ,  257  are connected to one hydraulic duct  281 ,  282  each. The hydraulic ducts  281 ,  282  are connected to the grooves  264 ,  266 . The grooves  264 ,  266  are connected to axial ducts  284 ,  286  which are formed in the axial direction in the hydraulic slave. The axial ducts  284 ,  286  are connected to radial ducts  288 ,  290  which are formed in the cylindrical section  268 . The radial ducts  288 ,  290  are positioned in the circumferential direction in the cylindrical section  268  in such a way that they are partially formed in the slave vanes  270 ,  272 . A bearing pin  292  of the hydraulic slave is rotatably mounted in the actuator  254 . The separating disk  258  is connected in a rotationally fixed fashion to the shaft  154 . 
     The connecting section  280  is mounted in a rotationally fixed fashion in a receptacle section of the shifting pin  240 . 
     The hydraulic master  250  is secured to the transmission housing (not illustrated) and/or connected thereto. The hydraulic slave  252  is mounted so as to be rotatable in relation to the hydraulic master  250 . The actuator  254  is mounted so as to be rotatable in relation to the hydraulic slave  252 . By virtue of the fact that the hydraulic ducts  281 ,  282  are connected to the circumferential grooves  262 ,  264 , hydraulic pressure can always be applied to the hydraulic slave  252  irrespective of the rotational position in relation to the hydraulic master  250 . The hydraulic pressure is fed to openings in the region of the slave vanes through the axial ducts  284 ,  286  and the radial ducts  288 ,  290 . 
       FIG. 24  is a perspective illustration in an assembly drawing of the hydraulic drive system  242 . Identical elements are denoted by identical reference numbers, with only the special features being explained here. 
     In the assembled state of the hydraulic drive system  242  which is illustrated in  FIG. 24 , a hydraulic chamber  296  is formed between the cylindrical section  274 , the separating disk  258 , the slave vane  270  and the actuator vane  278 . On the opposite side of the slave vane  270 , a further hydraulic chamber  298  is formed. Likewise, two further hydraulic chambers  296 ′,  298 ′ are formed on opposite sides of the actuator vanes  278 ,  276 . The radial ducts  288 ,  290  are formed as cylindrical grooves on two sides of the slave vanes  270 ,  272 . Formed adjacent to the slave vanes  270 ,  272  are openings  300 ,  302 , into which openings  300 ,  302  the radial ducts  288 ,  290  lead. 
     If a hydraulic pressure is built up by the hydraulic connection  257 , the hydraulic fluid passes through the opening  300  into the hydraulic chamber  296 . The hydraulic pressure applies a force to the actuator vane  278  and moves it in the circumferential direction. As a result, the actuator  254  is rotated and therefore the shifting pin which is connected to the actuator  254  also rotates. Since the separating disk  258  is connected in a rotationally fixed fashion to the shaft  154 , the shifting pin  240  is therefore rotated in relation to the shaft  154 . 
     If a hydraulic pressure is fed through the hydraulic connection  256 , hydraulic fluid passes through the axial duct  284  and the radial duct  288  to the opening  302  and into the hydraulic chamber  298 . The hydraulic pressure in the hydraulic chamber  298  moves the actuator vane  276  and therefore rotates the shifting pin  240  in an opposite direction. 
     The functional principle of the hydraulic drive system  242  will be explained in more detail below. 
       FIG. 25  illustrates a section along the line B-B from  FIG. 23 . Identical elements are denoted by identical reference numbers, with only the differences or special features being described here. 
     The hydraulic chambers  296 ,  296 ′  298 ,  298 ′ are formed between the shaft  154  and the cylindrical section  168 . The hydraulic chambers  296 ,  298 ′ are connected to the hydraulic connection  257  via the radial ducts  290  and the axial duct  286 . If hydraulic pressure is applied to the hydraulic connection  257 , a hydraulic pressure is built up in the hydraulic chambers  296 ,  298 ′ and the actuator vanes  276 ,  278  are rotated in the clockwise direction. The actuator vanes  276 ,  278  are correspondingly rotated in the counter-clockwise direction if a hydraulic pressure is applied to the hydraulic connection  256 . 
     The hydraulic chambers  298 ,  298 ′,  296 ,  296 ′ which are illustrated in  FIG. 25  consequently operate according to the principle of a double-acting cylinder. 
     It is also conceivable for the actuator vanes  276 ,  278  to be moveable independently of one another. The hydraulic chambers  298 ,  298 ′ are connected in series with the hydraulic chambers  296 ,  296 ′ in this alternative embodiment, with the result that a hydraulic system is implemented such as is explained schematically in  FIGS. 20   a  to  20   c . As a result, the shifting pin  240  could be rotated through a rotational angle which is twice as large. The ducts  228 ,  230  are arranged here in the hydraulic slave  252  in such a way that the ducts  228 ,  230  are opened precisely when one of the actuator vanes  276 ,  278  has reached a stop. This ensures that the hydraulic chambers  296 ,  296 ′,  298 ,  298 ′ are filled and respectively emptied sequentially. 
       FIG. 26  illustrates a section through the hydraulic drive system  242  along the line C-C. 
       FIG. 27  illustrates a section along the line A-A from  FIG. 23 .  FIG. 27  shows the latching device which is formed by the ball  248 , the spring  246  and drill holes  304 . The drill holes are formed at different circumferential positions in the shaft  154 . The spring  246  applies a force to the ball  248 . The ball is pressed by this force into the drill hole  304  or partially into the drill hole  304 , and thereby forms a latching connection. This latching connection causes the shifting pin  240  to latch in at predefined rotational positions in relation to the shaft  154 . A predetermined torque must advantageously be applied to the shifting pin  240  in order to release the latching device and rotate the shifting pin  240  in relation to the shaft  154 . The relative rotational position is thereby defined and fixed. 
     In one alternative embodiment, the spring  246  and the ball  248  are arranged in a radial drill hole which is formed in the shaft  154 . In this context, drill holes, with which the ball  240  forms a latching connection, are formed at different circumferential positions in the shifting pin  240 . 
     Alternatively, the freewheel bodies can also be activated magnetically. For this purpose, the actuation sections  164  can be provided with permanent magnets. Alternatively, the shifting pin  156  can be activated by means of electromagnetic actuators. 
       FIG. 28  illustrates the transmission housing  20  in an exploded illustration. The transmission cage  34  is provided for accommodating and mounting the transmission unit  10 ,  110 . The transmission cage  34  is formed by the pins  40  which are connected to the bearing plate  36  and to the housing cover  26 . The housing cover  26  advantageously forms both a termination of the housing casing  22  and the bearing plate  36  for the transmission cage  34 . The bearing plate  36  can preferably also be formed in one piece with the housing cover  24 , with the result that a further saving in weight is achieved. 
     In an alternative embodiment, the transmission units  10 ,  110  can also be shifted with an axially displaceable shifting pin. The shifting pawls  142  are, in this alternative embodiment, of similar or identical design to the shifting device  70 ,  104  with rotatable shifting pin  156 . The axially displaceable shifting pin has recesses with oblique sections, wherein the recesses are arranged under the actuation sections  144 , with the result that the prestressed shifting pawls  142  pivot out. The engagement of the shifting pawls  142  in the internal toothing  136  takes place as in the rotatable shifting pin  156 . The oblique sections of the recesses serve to allow the actuation sections  144  of the shifting pawls  142  to slide more easily out of the recess and therefore permit the engagement section  148  to be pivoted radially inward. As in the case of the rotatable shifting pin  156 , the recesses are arranged on the shifting pin in such a way that two gearspeeds are engaged simultaneously, and the so-called intermediate state is therefore set when shifting from one gearspeed into the other gearspeed occurs. As a result, shifting under load is also possible in this embodiment. The axially displaceable shifting pin can be activated by a shifting cable. The stationary or non-rotating shifting cable is decoupled from the rotating shifting pin by means of a sliding bearing or roller bearing. Alternatively, the shifting cable can be connected to a rotating disk which is connected to the shifting pin via a groove guide. In this context, a pin engages in the obliquely running groove which is formed in the rotating disk. The shifting cable is connected to the rotating disk. The disk is rotated by the shifting cable and the rotational movement of the disk is converted into an axial movement of the shifting pin through the pin which is guided in the groove. Alternatively, the pin can be secured to the disk and the groove can be formed in the shifting pin. 
       FIG. 29  shows a circuit diagram of a shifting device with two rotatable shifting pins. The shifting device which is illustrated in  FIG. 29  is an alternative embodiment to the shifting device  104  illustrated in  FIG. 4  and is generally denoted by  310 . Identical elements are denoted by identical reference numbers, with only the differences being explained here. 
     The driving gear  76  is connected in a rotationally fixed fashion to a driven gear  312 . The driven gear  312  is connected to two planetary gear mechanisms  311 ′,  311 ″. The driving gear  76  forms, together with the driven gear  312 , an epicyclic transmission  313 . The driven gear  312  is connected in a rotationally fixed fashion onto the planetary carriers  92 ′,  92 ″ of the planetary gear mechanisms  311 ′ and  311 ″. Planetary gear sets are mounted on the planetary carriers  92 ′,  92 ″. The planetary gear sets are each formed by a first planetary gear  314  and a second planetary gear  316 . The first planetary gear  314  is connected in each case in a rotationally fixed fashion to the second planetary gear  316 . A first sun gear  318  is arranged coaxially with respect to the planetary carrier  92  and meshes with the first planetary gears  314 . The first sun gear  318  is connected in a rotationally fixed fashion to the output shaft  96  which is connected in a rotationally fixed fashion into the driving gear  100  of the epicyclic transmission  98 . A second sun gear  320 , which meshes with the second planetary gears  316 , is mounted coaxially with respect to the sun gear  318 . The second sun gear  320  is connected in a rotationally fixed fashion to the ring gear shaft  93  which is connected to the tension disk  94 . The first sun gear  318  and the first planetary gears  314  form a first transmission ratio which differs from a second transmission ratio of the second sun gear  320  with the second planetary gears  316 . 
     The rotation of the shaft  72  is transmitted to the driven gear  312  via the driving gear  76 . The driven gear  312  drives the planetary gear mechanisms  311 ′,  311 ″. In this context, the planetary carrier  92 , on which the planetary gear sets are mounted, is driven. The planetary gear mechanisms  311 ′,  311 ″ are formed as stepped planetary gear mechanisms. The first sun gear  318  meshes with the first planetary gears  314  and forms the output of the planetary gear mechanisms  311 ′,  311 ″. If the tension disk  94  is not activated and the second sun gear  320  is therefore at rest, the rotation of the shaft  72  is transmitted to the shifting pins  74 ′,  74 ″ via the epicyclic transmission  313 , the planetary gear mechanisms  311 ′,  311 ″ and the second epicyclic transmissions  98 ′,  98 ″. In this state, the transmission ratio is just one, with the result that the shifting pins  74 ′,  74 ″ rotate synchronously with the shaft  72  or with the same rotational speed as the shaft  72 . The second sun gear  320  serves as a further driving gear of the planetary gear mechanisms  311 ′,  311 ″. A rotation of the sun gear  320  is consequently added to the rotation of the driven gear  312 , with the result that a rotation of the tension disk  94  can be transmitted to the shifting pin  74 . 
     The method of functioning of the shifting device  310  is consequently identical to the method of functioning of the shifting device  104  from  FIG. 5 . 
       FIG. 30  illustrates an exploded illustration of the shifting device  310 . Identical elements are denoted by identical reference numbers, with only the differences or the special features being presented here. 
     The driven gear  312  and the planetary carriers  92 ′,  92 ″ are formed as a gearwheel with bearing holes and bearing pins. The driving gear  312  is mounted so as to be rotatable by means of a ball bearing on a bearing shaft  322 . The planets are formed from the planetary gears  314  and  316  which have different diameters and/or numbers of teeth. The first sun gear  318 ′ is formed as an external toothing on the output shaft  96 ′ which is formed as a hollow shaft. The second sun gear  320 ′ is formed as an external toothing on the ring gear shaft  93 ′ which is formed as a hollow shaft. The output shaft can be connected to the driving gear  100 ′ of the epicyclic transmission  98 ′. The ring gear shaft  93 ′ is connected in a rotationally fixed fashion to the tension disk  94 ′. The bearing shaft  322 , the output shaft  96 ′ and the ring gear shaft  93 ′ are formed in such a way that they can be arranged or mounted coaxially one in the other. 
     The first sun gear  318 ″ is formed as an external toothing of the output shaft  96 ″. The output shaft  96 ″ can be connected in a rotationally fixed fashion to the driving gear  100 ″. The second sun gear  320 ″ is formed as an external toothing and is connected to a gearwheel  324 ″, wherein the second sun gear  320 ″ and the gearwheel  324 ″ are preferably formed in one piece. The gearwheel  324 ″ meshes with the gearwheel  95  from  FIG. 14 , which is connected in a rotationally fixed fashion to the ring gear shaft  93 ″. 
     Alternatively, the ring gear shaft  93 ″ can be connected to a further gearwheel  95 ″ from  FIG. 28 . The gearwheel  95 ″ then meshes with a gearwheel  324 ″, which is connected in a rotationally fixed fashion to the tension disk  94 ″. As a result, both tension disks  94 ′,  94 ″ can be arranged coaxially on one side of the transmission unit. 
     As a result of this arrangement illustrated in  FIG. 30 , two stepped planetary gear mechanisms which serve to rotate the shifting pins  74 ′,  74 ″ are formed.