Shifting arrangement for a bicycle gearbox, gearbox unit and shifting method

Abstract

A shift arrangement (200) for a vehicle transmission (18), in particular for a transmission of a vehicle powered by muscle power and / or engine power, comprising a transmission shaft arrangement (202) on which several loose gears (132) are rotatably mounted and which can be connected to the transmission shaft arrangement (202) via respective shift means (142), wherein the shift means (142) can be actuated by means of a camshaft arrangement (204) which has a first camshaft. The first camshaft (64A) and a second camshaft (64B) of the camshaft arrangement (204) are coupled to each other via a reversing gear (220) such that a rotation of one of the camshafts in a first direction of rotation (D1) leads to a rotation of the other camshaft in a second direction of rotation (D2), which is opposite to the first direction of rotation (D1).

Description

[0001] The present disclosure relates to a shifting arrangement for a vehicle transmission, in particular for a transmission of a vehicle that can be driven by muscle power and / or motor power, having a transmission shaft arrangement on which a plurality of idler gears are rotatably mounted, which idler gears can be connected to the transmission shaft arrangement via respective shifting means, wherein the shifting means can be actuated by means of a camshaft arrangement that has a first camshaft.

[0002] The present invention therefore generally relates to the field of human-powered vehicles, in particular bicycles. However, the invention also relates to vehicle transmissions such as those used in vehicles powered by human power and motor power, i.e., in e-bikes in which an electric motor provides additional drive power. However, the invention is also applicable to other types of motor-powered vehicles, for example, e-bikes without pedals. Generally, shifting arrangements of the type described are also applicable to other fields involving drive trains with spur gears that can be shifted via camshafts.

[0003] In the field of bicycles and e-bikes, a distinction is essentially made between three types of gear systems. The first type is derailleur gears, which have been around for many years. The second type is a so-called hub gear, which is often integrated into the hub of a bicycle's rear wheel. This is usually implemented as a planetary gear.

[0004] The third type of gearshift is a so-called bicycle gearshift, which is mounted in the area of ​​a bottom bracket. Such a bicycle gearshift, which is implemented in a front-end design, is known, for example, from document EP 2 379 402 B2. This document relates to a gearshift device for a transmission unit, comprising a first shaft on which a plurality of idler gears are mounted, a corresponding number of gears mounted on at least one second shaft, wherein the idler gears are connectable to the first shaft by means of respective switching means, and wherein the first shaft is designed as a hollow shaft having one or two coaxially internal camshafts (referred to therein as switching pins). The camshaft(s) is / are connected to drive means designed to rotate the camshaft(s) in order to actuate the switching means. The drive means include a speed superposition gear designed as a planetary gear.The switching means are designed as switchable freewheels. Actuating sections are formed on the camshaft, by means of which the freewheels can be actuated. The actuating sections are arranged on the camshafts in such a way that the freewheels of two consecutive gear stages can be actuated simultaneously.

[0005] If the transmission unit has two partial transmissions connected in series in the manner of a group transmission, two camshafts are provided which can be operated independently of one another.

[0006] The switching device described in document EP 2 379 402 B2 enables very short switching times.

[0007] From the document WO 2019 / 166563 A1, another transmission unit is known in which a switching arrangement is designed so that switching can take place under load.

[0008] The demands placed on such shifting arrangements are significantly higher for e-bike applications than for conventional bicycles. Due to the sometimes significantly higher torques transmitted via the transmission unit and consequently also partially via the shifting arrangement, particularly stringent requirements arise with regard to the timing of shifting operations.

[0009] Against this background, it is an object of the invention to provide an improved switching arrangement, an improved transmission unit and improved switching methods.

[0010] The above object is achieved by a switching arrangement having the features of claim 1, wherein the first camshaft and a second camshaft of the camshaft arrangement are coupled to one another via a reversing gear such that a rotation of one of the camshafts in a first direction of rotation leads to a rotation of the other camshaft in the second direction of rotation, which is opposite to the first direction of rotation.

[0011] Furthermore, the above object is achieved by a transmission unit having such a switching arrangement and by a method for carrying out gear changes by means of such a switching arrangement, with the step of rotating the camshaft arrangement by an alternating angle of rotation in order to disengage a source gear stage and to engage a target gear stage.

[0012] The shifting arrangement according to the invention is particularly suitable for use in transmission units that have two shiftable partial transmissions, wherein the idler gears of the first partial transmission and the second partial transmission are preferably mounted on a transmission shaft arrangement. This results in a situation in which the shifting means in the second partial transmission, viewed in the circumferential direction, must be aligned opposite to the shifting means of the first partial transmission (see, for example, Fig. 11 of the above-mentioned European patent EP 2 379 402 B2).

[0013] By coupling the first camshaft and the second camshaft via a reversing gear, the camshafts' directions of rotation are therefore opposite. This allows the same kinematics to be created between the two camshafts and the associated shifting devices. In other words, the shifting devices can be aligned uniformly in the circumferential direction. This allows the same, good shiftability under load to be achieved in both sub-transmissions.

[0014] However, it is still possible to drive only one of the two camshafts to perform gear shifts. Consequently, it is not necessary to control the camshafts independently of each other, which would entail increased mechanical and control complexity.

[0015] Preferably, neither of the two camshafts is axially movable. Rather, the axial positions of the camshafts are preferably fixed.

[0016] Preferably, the camshaft arrangement is actuated by means of a switching actuator, in particular in the form of an electric motor, preferably by coupling the switching actuator to only one of the first camshaft and the second camshaft.

[0017] The problem is thus completely solved.

[0018] It is particularly advantageous if the transmission shaft arrangement is designed such that a drive force can be transmitted via the transmission shaft arrangement, if both a plurality of first idler gears of a first partial transmission and a plurality of second idler gears of a second partial transmission are rotatably mounted on the transmission shaft arrangement, and if the transmission shaft arrangement is designed as a hollow shaft at least in sections.

[0019] It is particularly advantageous if the camshaft arrangement is rotatably mounted at least in sections in the transmission shaft arrangement and has the first camshaft assigned to the first idler gears and the second camshaft assigned to the second idler gears.

[0020] In general, it is conceivable to arrange the transmission shaft and the camshaft non-coaxially. However, it is particularly advantageous if the transmission shaft and the camshaft are aligned coaxially with each other.

[0021] According to a further preferred embodiment, the reversing gear has a planetary gear set, wherein it is preferred if a first member (e.g. sun gear) of the planetary gear set is coupled to the first camshaft, if a second member (e.g. planet carrier) of the planetary gear set is connected to the transmission shaft arrangement and if a third member (e.g. ring gear) of the planetary gear set is coupled to the second camshaft.

[0022] This allows for a relatively simple design for the reversing gear. Furthermore, the reversing gear can preferably be arranged coaxially with the gear shaft assembly and the camshaft assembly.

[0023] It goes without saying that the three elements of the planetary gear set can also be connected differently than described above.

[0024] As explained below, the use of the planetary gear set inevitably establishes a gear ratio between the first camshaft and the second camshaft. The gear ratio is preferably selected, as explained below, to result in an increase in torque or a reduction in speed at the second camshaft.

[0025] In this embodiment, it is preferred if only the first camshaft is driven to perform gear shifts. The second camshaft is connected to the first camshaft via the reversing gear, so that the second camshaft is indirectly driven by a member of the reversing gear, preferably by the ring gear.

[0026] Generally, it is preferred if the gearshift arrangement is used in an e-bike, in which case it is preferred if an electric motor is connected between the first sub-transmission and the second sub-transmission. Since significantly increased torques are transmitted via the second sub-transmission in this case, it is advantageous that the shifting means of the second sub-transmission are actuated with increased shifting forces due to the gear ratio of the planetary gear set.

[0027] According to a further preferred embodiment, the switching arrangement includes a speed superposition gear via which the transmission shaft arrangement and the camshaft arrangement are coupled and which can be actuated in order to achieve a relative rotation of the transmission shaft arrangement and the camshaft arrangement for switching the gear stages.

[0028] It is particularly advantageous if the speed superposition gear is coupled to the gear shaft arrangement and to one of the first and second camshafts, preferably to the first camshaft.

[0029] The basic design of the shifting arrangement is therefore such that the camshaft arrangement and the transmission shaft arrangement rotate at the same speed when a gear is engaged. To perform a gear change, a relative rotation occurs between the transmission shaft arrangement and the camshaft arrangement, so that a shifting element associated with a source gear is disengaged and a shifting element associated with a target gear is engaged. After the shifting process is completed, the transmission shaft arrangement and the camshaft arrangement rotate again at the same speed.

[0030] The speed superposition gear, as described for example in document EP 2 379 402 B2, to the disclosure content of which reference is made in this regard, can essentially be arranged on a shaft which is offset parallel to the gear shaft arrangement.

[0031] However, it is particularly preferred if the speed superposition gear is arranged coaxially with the transmission shaft assembly and the camshaft assembly. It is preferred if the speed superposition gear is arranged at a first axial end of the camshaft assembly.

[0032] The reversing gear can be arranged at any point of the gear shaft arrangement or the camshaft arrangement, viewed in the axial direction.

[0033] However, it is particularly preferred if the reversing gear is arranged at a second axial end of the camshaft arrangement.

[0034] In general, the reversing gear can be arranged on a shaft parallel to the gear shaft arrangement. Preferably, however, the speed superposition gear is arranged coaxially with it. It is particularly preferred if the reversing gear is arranged radially within the gear shaft arrangement. This results in a particularly compact design.

[0035] The first sub-gearbox is preferably arranged axially adjacent to a first axial end of the camshaft assembly. The second sub-gearbox is preferably arranged axially adjacent to a second axial end of the camshaft assembly.

[0036] In general, it is conceivable that the first camshaft and the second camshaft are adjacent to each other in the axial direction.

[0037] In a particularly preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, a second camshaft is designed as a hollow shaft and is rotatably mounted on an axial section of the first camshaft.

[0038] This results in an axially compact design. Furthermore, the connection to a reversing gearbox can be implemented in a relatively simple manner.

[0039] According to a further preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, the first camshaft and a second camshaft are coupled to one another via a transmission gear such that a rotation of one of the camshafts by a first angle of rotation leads to a rotation of the other camshaft by a second angle of rotation which differs from the first angle of rotation.

[0040] As explained above, the reversing gear is preferably designed so that it also has the function of a transmission gear, especially for slow speeds.

[0041] In general, the idea of ​​coupling the two camshafts via a transmission gear is also advantageous if the first and second camshafts generally rotate in the same direction, i.e. are not coupled to each other via a reversing gear.

[0042] It is preferred if the transmission gear is designed such that the second angle of rotation is smaller than the first angle of rotation. In this case, a higher shifting force can be realized in the second partial transmission, which is appropriate for the torque increase provided by an electric motor.

[0043] Furthermore, it is preferred if the transmission gear has a gear ratio in a range of 2 to 4, preferably a gear ratio of 3.

[0044] Furthermore, it is preferred if the second angle of rotation is equal to the first angle of rotation divided by the transmission ratio of the transmission gear, for example W2 = W1 / 3 (e.g., 20° = 60° / 3).

[0045] According to a further preferred embodiment, the first camshaft is rotatable by 360° or more and the second camshaft is rotatable over an angular range that is smaller than 360°.

[0046] Assuming a gear ratio of 3, it is therefore possible, for example, to cover all gear steps by rotating the first camshaft by 900° and by rotating the second camshaft by 300°, or by angles proportional thereto.

[0047] So if, for example, the first sub-transmission has four sub-transmission gear stages, it is advantageous if these four sub-transmission gear stages are set over an angular range of less than 360°, e.g. 180°. Gear stages 1 to 4 can then be set over the 180°. If, for example, the second sub-transmission has three sub-transmission gear stages, the first camshaft can then be rotated by a further 180°, for example, in order to shift from the fourth sub-transmission gear stage in the first sub-transmission back into the first sub-transmission gear stage of the first sub-transmission, which now represents gear stage 5 of the overall transmission when the second sub-transmission is simultaneously shifted into the second sub-transmission gear stage there.

[0048] If the first four gears are shifted to 180°, the gear changes result in a rotation angle of 60°. If the gear ratio in the transmission is 3, the shift in the first sub-transmission from sub-transmission gear 4 to sub-transmission gear 1 over 180° can correspond to a shift of 60° in the second sub-transmission.

[0049] According to a particularly preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, each gear change in the first partial transmission is consequently assigned a respective first change angle of rotation of the first camshaft, wherein each gear change in a second partial transmission is assigned a respective second change angle of rotation of a second camshaft, wherein the first change angle of rotation (e.g. 60°) of at least one partial transmission gear change in the first partial transmission is equal to the second change angle of rotation (e.g. 60°) of at least one partial transmission gear change in the second partial transmission.

[0050] In other words, the kinematics / geometry for gear changes between the first sub-transmission and the second sub-transmission can be standardized, resulting in synergy effects in the production of the switching means and the camshafts.

[0051] Preferably, the pitch, as mentioned above, is selected such that the total angle of the shifting operations in the first partial transmission is exactly 360°. Consequently, a repetition of this shifting pattern in the first partial transmission can be achieved over several revolutions of the first camshaft. For example, the pitch for four partial transmission gear stages can be set to an alternating rotation angle of 60°, so that upon reaching partial transmission gear stage 4, 180° are covered (4 partial transmission gear stages - 1 × 60°), and that the remaining angle of 180° is then rotated from gear stage 4 back to partial transmission gear stage 1.

[0052] The second camshaft is preferably designed such that the partial transmission gear stage remains set in the second partial transmission while the partial transmission gear stages in the first partial transmission are shifted through.

[0053] Generally speaking, a gear ratio between the first partial transmission (TG1) and the second partial transmission (TG2) can be calculated as follows if the pitches in Tg1 and Tg2 are to be identical: 360°−(number of gear stages TG1−1)×pitch TG1=ratio×pitch TG2

[0054] For example, with four partial transmission gear stages in TG1 and a pitch TG1 of 60° and a desired identical pitch TG2 of 60°, the ratio results in a ratio of 3: (360° - (4 - 1) × 60° = i × 60°)

[0055] With three partial transmission gear stages in TG1 and a pitch TG1 of 90° and a desired identical pitch TG2 of 90°, the result is, for example, a ratio of 2: (360°−(3−1)×90°=i×90°)

[0056] With five partial transmission gear stages in TG1 and a pitch TG1 of 45° and a desired identical pitch TG2 of 45°, the result is, for example, a ratio of 4: (360°−(5−1)×45°=i×45°)

[0057] According to a further preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, the first camshaft and a second camshaft are coupled to one another via a torque change gear such that a rotation of one of the camshafts with a first torque leads to a rotation of the other camshafts with a second torque which differs from the first torque.

[0058] As mentioned above, a transmission that slows down (so that the second camshaft rotates proportionally to the first camshaft, but more slowly) inevitably results in an increase in torque on the second partial transmission.

[0059] As mentioned, this advantageous design can be used to simplify use in e-bikes. Because the higher torque of the camshaft acting on the shifting mechanism allows the shifting mechanism to be disengaged under higher load when engaging a gear, which is particularly advantageous when using e-bikes.

[0060] Overall, it is advantageous if the transmission gear is formed by the reversing gear and / or if the torque change gear is formed by the reversing gear.

[0061] According to a further overall preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, at least one switching means is designed as a switching pawl which is mounted radially pivotably on the transmission shaft arrangement and which can be actuated between a pivoting position and a gear position by means of a functional cam body of the camshaft arrangement, wherein the functional cam body is designed as a cam body which is mounted rotatably between a first functional stop and a second functional stop.

[0062] This design makes it possible for the functional cam body to initially not rotate during rotation of the camshaft arrangement until it is forced to rotate from a certain rotational position of the camshaft arrangement due to a functional stop.

[0063] In more general terms, this preferred embodiment is defined by a separate invention in the form of a method according to claim 19 for connecting two rotary members which are rotatable relative to one another and connectable by means of a switching means, while these rotate relative to one another, namely by means of a camshaft which has a functional cam body which (i) is mounted on the camshaft so as to be resiliently rotatable relative to a basic position and which (ii) is designed to actuate the switching means, wherein the camshaft arrangement is rotated to connect the rotary members which rotate relative to one another, but the functional cam body is initially not rotated along with it until it is forcibly rotated along with it from a certain rotational position of the camshaft arrangement and is then resiliently rotated back into the basic position.

[0064] The functional cam body can be assigned to a single gear stage, but can also be formed by a cam body assigned to several gear stages (e.g. second camshaft).

[0065] Preferably, the functional cam body is initially not rotated during rotation of the camshaft arrangement by being held in place relative to one of the rotary members or relative to the switching means.

[0066] In general, two main advantages can be realized with this process.

[0067] Firstly, the risk of the shifting mechanism or the shift pawl only partially engaging the internal toothing of an idler gear when engaging a gear can be reduced. This can occur, for example, in conventional designs, particularly when the camshaft rotates comparatively slowly and the shift pawl is only hesitantly pivoted out, for example, during an upshift. This applies in particular to the second sub-transmission, whose camshaft preferably rotates a factor of one gear ratio (e.g., 3) slower than the first camshaft.

[0068] The second advantageous aspect resulting from the rotatability of the functional cam body with respect to the associated camshaft is the play compensation that can be achieved thereby, in particular between the first camshaft and the second camshaft.

[0069] In other words, asymmetric camshaft timing can be achieved by the rotatability of the functional cam body.

[0070] Such asymmetric camshaft timing means that the angle of rotation, and thus the timing, of the respective shifting mechanism actuation differs during downshifting from that during upshifting. The angle-dependent angular positions at which certain shifting mechanisms (e.g., shift pawls) pivot out or in can therefore be different during upshifting than during downshifting.

[0071] The functional cam body, which is rotatable between two functional stops relative to the associated camshaft, is preferably pre-tensioned by means of a functional spring against one of the functional stops in such a way that the functional cam body remains at this first functional stop during downshifts.

[0072] On the other hand, it is preferable if the functional cam body is rotated against the effect of such a functional spring during upshifts (i.e. is not rotated).

[0073] As a result, the functional cam body cannot initially rotate until it is forced to rotate from a certain rotational position of the camshaft (in particular when the second functional stop is reached) and is then rotated back to the basic position in a spring-elastic manner by means of the functional spring.

[0074] This makes it possible to ensure that the switching means is quickly pushed outwards by the functional cam body, which is quickly actuated by the force of the functional spring, so that reliable engagement in an internal toothing of a loose wheel is ensured even if the functional cam body otherwise rotates comparatively slowly.

[0075] According to a further preferred embodiment, the angle between the first functional stop and the second functional stop is selected such that it compensates for at least one circumferential play that exists between the functional cam body and the associated camshaft.

[0076] Such circumferential play is particularly present when the first camshaft and the second camshaft are not rigidly connected to each other, but rather via connecting means. Such play arises in particular when the first camshaft and the second camshaft are coupled via a gear, i.e., in particular, a transmission gear or a reversing gear.

[0077] According to a further preferred embodiment, the angle between the first functional stop and the second functional stop is selected such that it is adapted to a relative rotatability between the functional cam body and the associated camshaft, that is to say, for example, to a transmission ratio between the first camshaft and the second camshaft.

[0078] This angle is preferably in a range of 5° to 20°, but can also be in an angle of 50° to 80°, depending on whether this angle is also intended to compensate for a gear ratio between the first camshaft and the second camshaft.

[0079] Furthermore, it is advantageous if this angle is selected in such a way that it is possible to engage a target gear independently of a switching contour of the cam body.

[0080] As mentioned, this allows a kind of “pre-tensioning” of the switching device to take place, which can then be pressed suddenly into the internal toothing of the idler gear from the moment of forced co-rotation, without the switching device having to “roll off” a switching contour (or vice versa).

[0081] This ensures that the switching element always fully engages the internal teeth of the gear, especially right down to the tooth base. This reduces wear on the switching element and the internal teeth.

[0082] In general, such a functional cam body can be assigned to each gear stage of the first partial transmission.

[0083] However, it is particularly advantageous if the functional cam body is assigned to the highest gear stage of the first partial transmission.

[0084] According to a further preferred embodiment, which in conjunction with the preamble of claim 1 constitutes a separate invention, the second camshaft is mounted so as to be rotatable to a limited extent with respect to the first camshaft between a first coupling stop (corresponding to a first functional stop) and a second coupling stop (corresponding to a second functional stop) and is prestressed against the first coupling stop by means of a coupling spring (corresponding to a functional spring).

[0085] As a result, the second camshaft as a whole forms a functional cam body, which is particularly advantageous because the second camshaft rotates more slowly than the first camshaft and, consequently, the “pre-tensioning” in the second partial transmission is of particularly advantageous importance.

[0086] In the second partial transmission, the cam bodies are preferably all rigidly connected to the second camshaft, so that the latter forms the cam bodies for all gear stages of the second partial transmission.

[0087] It is particularly advantageous if the second camshaft can be held in a locking position in relation to the transmission shaft arrangement by means of a locking arrangement in a plurality of locking positions.

[0088] This allows the coupling spring to be equipped with a comparatively high spring force, which can ensure the rapid pivoting of a switching pawl. In this case, the second camshaft is preferably held in a respective detent position, so that the second camshaft initially does not rotate with the first camshaft during rotation until a certain rotational position of the first camshaft is reached, which causes a stop to be reached, which drives the second camshaft along and releases the detent arrangement.

[0089] In the case of a cam body which is assigned to a single idler gear arranged in the first partial transmission (such as, for example, an idler gear for the highest gear stage of the first partial transmission), either such a locking arrangement can also be provided between the cam body and the transmission shaft arrangement, and / or a torsion spring element which counteracts co-rotation of the cam body.

[0090] Overall, at least one of the following basic ideas is revealed here: - an opposite direction of rotation of camshafts of series-connected switchable partial transmissions is realized via a reversing gear, so that the camshafts do not have to be driven independently of each other, - individual cam bodies are held and pre-tensioned to achieve immediate release of a switching device when a certain angle of rotation of the camshaft is exceeded, - a cam body or a camshaft is designed in such a way that switching means can be pivoted in or out at different angles of rotation depending on the direction of rotation, thus resulting in “asymmetric camshaft timing”, - Play compensation of different camshaft directions is achieved by a pre-tensioning mechanism on cam bodies or (partial) camshafts.

[0091] It is understood that the features of the invention mentioned above and those to be explained below can be used not only in the combination specified in each case, but also in other combinations or on their own, without departing from the scope of the invention.

[0092] Further features and advantages of the invention will become apparent from the following description of preferred embodiments with reference to the drawings. Fig. 1 a schematic representation of a bicycle frame with a gear unit in the area of ​​the pedal cranks; Fig. 2 a schematic layout of a transmission unit; Fig. 3 a schematic representation of a speed superposition gear of a switching arrangement; Fig. 4 a perspective view of a switchable idler gear with internal teeth; Fig. 5 a perspective view of a switching means in the form of a switching pawl or freewheel body, which can be pivotally mounted on a transmission shaft arrangement; Fig. 6a-6f are schematic cross-sectional views of a switching arrangement for illustrating a switching process from a source gear stage to a target gear stage; Fig. 7 is a perspective view of an embodiment of a switching arrangement; Fig. 8 a schematic longitudinal sectional view of the switching arrangement of the Fig. 7; Fig. 9 is a perspective exploded view of the switching arrangement of the Fig. 7; Fig. 10a schematic exploded view of the circuit arrangement of the Fig. 7 with an enlarged view of a first partial transmission; Fig. 11 a schematic perspective view of a first camshaft of the switching arrangement of the Fig. 7; and Fig. 12a and Fig. 12b schematic representations of two camshafts and their angles of rotation in association with gear stages.

[0093] In Fig. 1 is a gear unit generally designated 10.

[0094] Fig. 1 shows a side view of a bicycle frame 12 having a gear housing 14 in which the gear unit 10 is accommodated. The gear unit 10 is only indicated schematically in this illustration and is designed as a compact unit, which is preferably arranged in a gear cage (not shown here). The gear unit 10 is described herein by way of example for use in a two-wheeler, although use in other muscle-powered vehicles is also possible. It is understood that the gear unit 10 can also be used for vehicles in which muscle power is used in combination with a drive motor to drive the vehicle, or which can be driven exclusively by means of such a drive motor. The drive motor is preferably an electric motor, and the vehicle is preferably an e-bike or a pedelec.

[0095] The gear unit 10 and the gear housing 14 together with pedal cranks 16 and 16' form a multi-speed transmission 18.

[0096] Fig. 2 shows a schematic layout of the transmission unit 10.

[0097] The transmission unit 10 has an input shaft 20 and an output shaft 22. The input shaft 20 is designed as a through shaft and can be connected to pedal cranks in a rotationally fixed manner. The output shaft 22 is designed as a hollow shaft. The input shaft 20 and the output shaft 22 are arranged coaxially with each other. The output shaft 22 is connected to a chainring 24 in a rotationally fixed manner, which forms an output member of the transmission unit 10.

[0098] The transmission unit 10 has a first sub-transmission 26 and a second sub-transmission 28. A plurality of drive gears 30, 31, 32, 33, 34, 35 are mounted on the input shaft 20. The first sub-transmission 26 has a countershaft 36. Driven gears 38, 39, 40, 41, 42, 43 are mounted on the countershaft 36. The driven gears 38 to 43 are designed as first idler gears.

[0099] The driven wheels 38 to 43 can each be connected to the countershaft 36 by means of first switching means (not shown). The driven wheels 38 to 43 and the drive wheels 30 to 35 each form gear pairs that have different gear ratios, so that by selectively connecting the driven wheels 38 to 43 to the countershaft 36, different gear stages can be realized in the first partial transmission 26.

[0100] The second sub-gearbox 28 has an input shaft 46. Drive gears 48, 49, and 50 are mounted on the input shaft 46. The drive gears 48 to 50 are designed as second idler gears. Drive gears 52, 53, and 54 are mounted on the output shaft 22. The driven gears 52 to 54 are each in meshing engagement with the drive gears 48 to 50.

[0101] The meshing driven gears 52 to 54 and drive gears 48 to 50 form gear pairs with different gear ratios. The drive gears 48 to 50 are rotationally fixedly connected to the input shaft 46 by means of second shifting means (not shown), thereby forming different, selectable gear stages of the second partial transmission 28.

[0102] The countershaft 36 of the first partial transmission 26 is preferably connected in a rotationally fixed manner to the input shaft 46 of the second partial transmission 28. The countershaft 36 is preferably formed integrally with the input shaft 46.

[0103] Preferably, the drive wheels 30 to 35 are each non-positively connected to the input shaft 20 by means of a clutch (not shown), and in particular frictionally connected by means of a friction clutch. The clutch is designed to limit a torque introduced into the transmission unit 10. The clutch is designed such that, if a predefined or adjustable torque is exceeded, the connection between the input shaft 20 and the corresponding drive wheel 30 to 35 slips. Such a torque limitation makes it possible to reduce the size and weight of the transmission unit, since the transmission unit can be designed for a lower maximum torque. The drive wheels 30 to 35 can also be connected to the input shaft 20 via a torque detection arrangement (not shown), for example to detect a torque introduced into the input shaft 20 by muscle power.The detected torque can be used to control an assisting drive motor.

[0104] Because the first partial transmission 26 is connected in series with the second partial transmission 28, the possible gear stages of the first partial transmission 26 are multiplied by the gear stages of the second partial transmission 28, similar to a group transmission. Fig. 2 illustrated transmission unit 10 eighteen gears can be realized.

[0105] Furthermore, it is conceivable that the input shaft 20 could be connected to the output shaft 22 in a rotationally fixed manner by means of a coupling (not shown). This would allow for an additional gear to be implemented as a direct gear.

[0106] In Fig. 3 shows a schematic layout of a switching arrangement with a rotating camshaft. In Fig. 3 is a switching arrangement generally designated 60.

[0107] The switching arrangement 60 generally serves to selectively connect idler gears (not shown) mounted on a shaft 62 in a rotationally fixed manner to the shaft 62 in a drive direction of rotation by means of switching means (not shown), such as switching pawls. The switching arrangement 60 has a camshaft 64 which is arranged coaxially in the shaft 62, which is designed as a hollow shaft, and is rotatably mounted relative thereto. At one axial end of the shaft 62 (or offset parallel thereto) there is a speed superposition gear 66 which is connected to both the shaft 62 and the camshaft 64. The speed superposition gear 66 is preferably arranged coaxially to the shaft 62. The speed superposition gear 66 includes a transmission stage 68 and a control stage 70. The control stage 70 is connected to the shaft 62 and the transmission stage 68 is connected to the camshaft 64.Alternatively, the control stage 70 can be connected to the camshaft 64 and the transmission stage 68 to the shaft 62. The transmission stage 68 is formed by a first planetary gear 68, in particular in the form of a planetary gear set. The control stage 70 is formed by a second planetary gear 70, in particular in the form of a planetary gear set. The first planetary gear 68 has a sun gear 72 that is rotationally fixedly connected to the camshaft 64. The first planetary gear 68 has planet gears 74 that are mounted on a planet carrier 76. The planet gears 74 mesh with the sun gear 72. The first planetary gear 68 has a ring gear 78, with which the planet gears 74 also mesh. The ring gear 78 is fixed to a fixed reference point 80, preferably a gear cage or gear housing (not shown), and is firmly connected thereto.

[0108] The second planetary gear 70 has a sun gear 82 that is rotationally fixedly connected to the shaft 62. The second planetary gear 70 has planetary gears 84 that are mounted on the planet carrier 76. The planetary gears 84 mesh with the sun gear 82. The second planetary gear 70 has a ring gear 86, with which the planetary gears 84 also mesh. The ring gear 86 is rotationally fixedly connected to a pulley 88, to which a Bowden cable (not shown) can be attached.

[0109] The first planetary gear 68 and the second planetary gear 70 are dimensioned such that when the ring gear 86 is stationary or held fast, the transmission ratio from the shaft 62 to the camshaft 64 is even or exactly 1, so that in this case the shaft 62 and the camshaft 64 rotate synchronously or at the same speed. The second planetary gear 70 or the control stage 70 serves to superimpose an additional speed on the speed of the shaft 62. This additional speed is transmitted via the ring gear 86 to the planet carrier 76. Because the planet gears 84 of the second planetary gear 70 and the planet gears 74 of the first planetary gear 68 are connected to one another via a planet carrier 76, a total speed as the sum of the speed of the shaft 62 and the rotation of the ring gear 86 is transmitted to the first planetary gear 68. The first planetary gear 68 orThe transmission stage 68 serves to transmit the total rotational speed to the camshaft 64. The ring gear 86 is rotationally fixedly connected to the tension disc 88 for actuation by a Bowden cable (not shown). The tension disc 88 is rotated by a specific angle of rotation in the direction of rotation of the shaft 62 or counter to the direction of rotation of the shaft 62 in order to transmit this relative movement to the camshaft 64. This allows the gears of the transmission unit 10 to be shifted by actuating the Bowden cable. The sun gear 62 is preferably formed as part of the shaft 62 or integrally with the shaft 62.

[0110] In an alternative embodiment, the two ring gears 78, 86 are mounted for free rotation and are connected to each other in a rotationally fixed manner. In this embodiment, the planet gears 74, 84 are each connected by a separate planet carrier. One of the planet carriers is connected to the pulley 88 to transmit the rotation of the pulley 88 to the camshaft 64.

[0111] In a further alternative embodiment, it is also conceivable for the sun gears to be connected to one another in a rotationally fixed manner and mounted for free rotation. In this case, shaft 62 would be connected to the planet carrier of planet gears 84, and camshaft 64 would be connected to the planet carrier of planet gears 74.

[0112] In Fig. 4 shows a switchable idler gear with internal teeth and is generally designated 132.

[0113] The idler gear 132 has external teeth 134 and internal teeth 136. The external teeth 134 are formed on the outer peripheral surface. The internal teeth are formed on an inner peripheral surface of the idler gear 132. The internal teeth 136 have sliding sections 138 and engagement sections 140. The sliding sections 138 are formed by surfaces formed in the circumferential direction of the idler gear 132. The engagement sections 140 are formed between the sliding sections 138 at an angle to the sliding sections 138.

[0114] The external toothing 134 serves to mesh with other gears. The internal toothing 136 serves to mount the idler gear 132 on the shaft 62 and to connect it to the shaft 62 in a rotationally fixed manner by means of switching means. The sliding sections 138 serve to rotatably mount the idler gear 132 on the shaft 62 and to slide on the shaft 62. The engagement sections 140 serve to enable switching means (not shown, which will be explained in more detail below) to engage the idler gear 132 and to connect the idler gear 132 to the shaft 62 in a rotationally fixed manner.

[0115] In Fig. 5 shows a switching means in the form of a freewheel body or a switching pawl for the rotationally fixed connection of the idler gear 132 to the shaft 62 and is generally designated 142. The freewheel body 142 has an actuating portion 144 formed on an underside or radial inner side of the freewheel body 142. The freewheel body 142 has a bearing portion 146 on each of two axially lateral portions. The freewheel body 142 has an engagement portion 148. The engagement portion 148 is formed at an end of the freewheel body 142 opposite the actuating portion 144 in the circumferential direction. The bearing portions 146 are arranged in the circumferential direction between the actuating portion 144 and the engagement portion 148.

[0116] The bearing sections 146 serve to mount the freewheel body 142 on a shaft (e.g. the shaft 62 in Fig. 3) to be mounted so as to be rotatable or pivotable about a rotational axis 150. The freewheel body 142 is mounted on the shaft such that the actuating portion 144 faces the interior of the shaft. Furthermore, the freewheel body 142 is preloaded by a spring element (not shown) such that the actuating portion 144 is pivoted radially inward in the unloaded state and the engaging portion 148 is pivoted radially outward. The actuating portion 144 is pressed radially outward by a cam 111 108 of the camshaft 64 in order to pivot the engaging portion 148 radially inward about the rotational axis 150.

[0117] If the engagement portion 148 is pivoted radially outward and protrudes relative to the shaft, it can be brought into engagement with the engagement portion 140 of the internal toothing 136 of the idler gear 132 in a direction of rotation of the idler gear 132 and thus connect the idler gear to the shaft in a rotational direction of the drive.

[0118] The freewheel body 142 further includes a sliding portion 152. The sliding portion 152 serves to pivot the freewheel body 142 radially inward when the idler gear is rotated relative to the shaft in a direction opposite to the drive rotation direction, thus serving as a freewheel.

[0119] The actuating portion 144 may have a groove extending perpendicular to the rotational axis 150 or in the direction of rotation of the shaft in order to receive a spring element for preloading the freewheel body 142.

[0120] In the Fig. 6a to 6f schematically illustrate a gear change between a source gear stage (e.g., gear stage 1) and a target gear stage (e.g., gear stage 2). Radial sectional views through idler gears 132-1 and 132-2 are shown, which correspond to the source gear stage and the target gear stage, respectively, during three states ( Fig. 6a and Fig. 6b, Fig. 6c and Fig. 6d and Fig. 6e and Fig. 6f) of the gear change.

[0121] Fig. Figure 6a shows the idler gear 132-1 assigned to the source gear stage, whose internal toothing 136-1 meshes with the two associated freewheel bodies 142-1. The camshaft 64 is positioned relative to the shaft 62 in a rotational position such that cam gaps 110-1 of the camshaft 64 are arranged in the region of the actuating sections 144-1 of the freewheel bodies 142-1, thus allowing the freewheel bodies 142-1 to pivot outward. The source gear stage is engaged.

[0122] The idler gear 132-2 assigned to the target gear is in Fig. 6b. The freewheel bodies 142-2 are pivoted radially inward by cams 111-2 and are therefore not engaged with the internal toothing 136-2 of the idler gear 132-2. In this rotational position of the camshaft 64, the cam gaps 110-2 associated with the target gear stage are not arranged below the actuating sections 144-2 of the freewheel bodies 142-2, so that the actuating sections 144-2 are pushed outward by the cams 111-2. The target gear stage is disengaged.

[0123] Based on the state of the Fig. 6a and Fig. 6b, the camshaft 64 is rotated, as indicated by an arrow 168, the cam gaps 110-1 remain in a second state or intermediate state initially below the freewheel bodies 142-1, which are assigned to the idler gear 132-1. The source gear stage therefore remains engaged in the second state, as shown in Fig. 6c, and the freewheel bodies 142-1 remain pivoted out.

[0124] In Fig. Figure 6d shows the idler gear 132-2 assigned to the target gear in this second state. In this rotational position of the camshaft 64, the cam gaps 110-2 assigned to the target gear are arranged radially below the actuating sections 144-2 of the freewheel bodies 142-2, so that the actuating sections 144-2 pivot radially inward and thus the engagement sections 148-2 can pivot radially outward. As a result, the engagement sections 148-2 are brought into engagement with the internal toothing 136-2 of the idler gear 132-2. The target gear is now also engaged.

[0125] Each freewheel body 142 is assigned a spring (not shown), which preloads the corresponding freewheel body 142 such that the respective actuating portion 144 is pressed against the camshaft 64. As a result, the respective engagement portion 148 pivots radially outward if one of the cam gaps 110 is rotated below the switching pawl 142.

[0126] Since the target gear is a higher gear with a lower gear ratio, the freewheel pawls 142-2 of the target gear engage the internal toothing 136-2 of the associated idler gear 132-2 and drive the shaft 62 at a rotational speed that is lower than the rotational speed of the idler gear 132-1 of the source gear. Therefore, the idler gear 132-1 of the source gear "overtakes" the shaft 62 in this so-called intermediate state. As a result, the sliding portion 138 of the idler gear 132-1 presses against the sliding portion 152 of the freewheel body 142-1, so that the freewheel body 142-1 is deflected inward and the idler gear 132-1 slides on the shaft 62. The idler gear 132-1 of the lower gear, i.e., the first gear, is in the intermediate state of freewheeling.

[0127] In the Fig. 6e and Fig. 6f shows the state in which the target gear is fully engaged. For this purpose, the camshaft 64 has been rotated further in the direction of arrow 168, so that the freewheel bodies 142-1 of the source gear are pivoted in by the cams 111-1 of the camshaft 64, as shown in Fig. 6e is shown. In Fig. 6f shows that the freewheel bodies 142-2 of the target gear stage continue to engage with the internal toothing 136-2 of the idler gear 132-2 because the cam gaps 110-2 are arranged below the actuating sections 144-2 of the freewheel bodies 142-2.

[0128] The intermediate state, in which the freewheel bodies 142-1, 142-2 are radially pivoted out of two gear stages, enables shifting under load. Furthermore, an idle state is avoided.

[0129] When downshifting from a high source gear (e.g. gear 2 in Fig. 6f) into a low target gear (e.g. gear 1 in Fig. 6a), in the intermediate state, the sliding section 138 of the internal toothing 136-1 of the target gear stage initially slides over the freewheel bodies 142-1. The higher gear initially remains engaged. The freewheel bodies 142-2 are only pivoted in, or the source gear stage disengaged, when the load transmitted to the shaft 62 via the idler gear 132-2 is reduced. In addition, the camshaft 64 must then be rotated further so that the actuating section 144-2 is pushed outwards. The lower target gear stage is then immediately engaged, because this gear stage has already been in the intermediate state ( Fig. 6c) was in the freewheeling state. This prevents an idle state.

[0130] In Fig. 6a to 6f, the camshaft 64 is shown with exactly opposite cams 111. Alternatively, it is also conceivable for the cams 111 to be arranged relative to one another such that only one of the freewheel bodies engages with the respective internal toothing 136. This is achieved by not arranging the respective freewheel bodies 142 on the shaft 64 exactly opposite one another. As a result, the angle of rotation of the respective idler gear 132 can be reduced until the actuating portion 148 engages the internal toothing 136. In an alternative embodiment, only one freewheel body 142 is assigned to each idler gear, and only one cam 111 or one cam gap 110 is assigned to the associated actuating element.

[0131] In the Fig. 7 to 11 show an embodiment of a switching arrangement which can be used for a transmission unit as described above.

[0132] The switching arrangement 200 includes a transmission shaft arrangement 202, which, for example, corresponds to the countershaft 36 of the Fig. 2 or wave 62 of the Fig. 3 can correspond.

[0133] The switching assembly 200 further includes a camshaft assembly 204 that includes a first camshaft 64A and a second camshaft 64B.

[0134] The first camshaft 64A is designed as a solid shaft and extends substantially over the entire axial length of the switching arrangement 200. At a first end E1, the gear shaft arrangements 202 and the first camshaft 64A are coupled to one another via a speed superposition gear 66, as described with respect to Fig. 3 has been described.

[0135] A plurality of cam bodies 206a, 206b, 206c, 206d are mounted on the first camshaft 64A, which is assigned to a first sub-transmission 26. The first sub-transmission has four sub-transmission gear stages. The cam bodies 206a, 206b, 206c, which are assigned to the sub-transmission gear stages 1 to 3 of the first sub-transmission 26, are connected in a rotationally fixed manner to the first camshaft 64A.

[0136] The cam body 206d is, as will be described below, rotatable to a limited extent on the first camshaft 64A.

[0137] The cam bodies 206 each have cams 111 and cam gaps 110 therebetween, as described above with respect to Fig. 6 have been described.

[0138] Each of the cam bodies 206 is assigned a bearing section 208a, 208b, 208c, 208d on the transmission shaft arrangement 202, on which respective freewheel bodies or switching pawls are pivotally mounted, as described above with reference to Fig. 5. The pivot axes 150 of these bearing sections 208 are in Fig. 7 schematically indicated. The pivoting or rotational axes 150 are aligned parallel to an axis of the gear shaft arrangement 202.

[0139] The cam bodies 206 of the first partial transmission 26 are arranged adjacent to the first axial end E1 of the first camshaft 64A.

[0140] In Fig. 8 it can be seen that the third cam body 206c is assigned a loose wheel 132, which is, for example, the loose wheel 41 for the third partial transmission gear stage of the transmission unit 10 of the Fig. 2. It is understood that the other cam bodies 206a, 206b, 206d are each assigned corresponding idler gears 132.

[0141] The idler gear 132, which is Fig. 8, has an internal toothing 136, into which a switching pawl or a freewheel body can engage positively in a direction of rotation, which is pivotally mounted on the bearing section 208c (in Fig. 8 not shown). The idler gear 132, which is for example driven by the idler gear 41 of the Fig. 2, is in engagement with a fixed gear 33 which is fixed in a rotationally fixed manner to the input shaft 20.

[0142] The second camshaft 64A is formed by a cam body 210, which is designed as a hollow shaft. The cam body 210 is rotatably mounted on the first camshaft 64A in an axial section A. The cam body 210 is associated with the second partial transmission 28. The second partial transmission 28 includes three partial transmission gear stages. Consequently, three cams 111e, 111f, 111g are formed on the cam body 210 (preferably integral with the cam body 210), which are associated with respective idler gears of the second partial transmission 28. Fig. 8 schematically shows that the cam 111e is assigned a loose wheel 132, which can be formed, for example, by a loose wheel of the second partial transmission, such as the loose wheel 48 of the Fig. 2. The idler gear 132 has an internal toothing 136, into which a switching pawl or a freewheel body 142 can engage in order to establish a positive connection between the idler gear 132 and the transmission shaft assembly 202 in a drive direction of rotation. It is understood that corresponding idler gears are assigned to the other cams 111f, 111g. Fig. 8 shown idler gear 132, which can be replaced, for example, by an idler gear 48 of the Fig. 2 can be realized, is, for example, in engagement with a fixed gear 52 which is fixed to an output shaft 22.

[0143] In this case, the first sub-transmission 26 contains four sub-transmission gear stages, and the second sub-transmission 28 contains three sub-transmission gear stages. The two sub-transmissions 26, 28 are connected in series like a group transmission. Consequently, a total of 4 × 3 = 12 gear stages can be set up with the sub-transmissions 26, 28.

[0144] The second partial transmission 28 is arranged adjacent to a second end E2 of the first camshaft 64A.

[0145] The first camshaft 64A is rotatably mounted in the region of the first end E1 by means of a bearing section 214 with respect to the transmission shaft arrangement 202.

[0146] In the region of the second end E2, the first camshaft 64A and the second camshaft 64B are coupled to one another via a reversing gear 220.

[0147] The reversing gear is arranged coaxially to the axis of the first camshaft 64A and the axis of the transmission shaft assembly 202 and includes a planetary gear set 222.

[0148] The planetary gear set 222 has three members that are coupled to the first camshaft 64A, the transmission shaft assembly 202, and the second camshaft 64B, respectively.

[0149] Preferably, a sun gear 224 of the planetary gear set 222 is rotationally fixedly connected to the first camshaft 64A. A planet carrier 226 of the planetary gear set 222 is preferably rotationally fixedly connected to the transmission shaft assembly 202. Preferably, the first camshaft 64A is mounted indirectly via the planet carrier 226 with respect to the transmission shaft assembly 202 in the region of the second end E2.

[0150] A ring gear 228 of the planetary gear set 222 is preferably coupled to the second camshaft 64B in the form of the cam body 210.

[0151] While the cam bodies 206a, 206b, 206c are connected to the first camshaft 64A in a rotationally fixed manner, the cam body 206d is limitedly rotatable between two Fig. 8 stops (not shown in detail) relative to the first camshaft 64A. The cam body 206d is preloaded against one of the two stops by means of a functional spring 230.

[0152] Similarly, the cam body 210 (second camshaft 64B) is limitedly rotatable between two Fig. 8 stops (not shown in detail) relative to the ring gear 228 of the second camshaft 64B. The cam body 210 is preloaded against one of the two stops by means of a functional spring 232.

[0153] Coaxial with the first camshaft 64A, a locking arrangement 234 is also provided between the first partial transmission 26 and the second partial transmission 28. The locking arrangement 234 includes a locking ring 236 arranged around the first camshaft 64A and connected in a rotationally fixed manner to the transmission shaft arrangement 202.

[0154] The locking ring 236 is in engagement with at least one locking body 238, which is mounted axially displaceably in the cam body 210, as is particularly shown in Fig. 9. More precisely, the locking body 238 is preloaded by means of a spring (not shown in more detail) towards the locking ring 236, which includes a plurality of locking recesses (not shown in more detail).

[0155] In Fig. 9, however, a plurality of locking points 240 are schematically shown, which correspond to the respective locking recesses.

[0156] As it is in Fig. 9, a rotation of the first camshaft 64A in a first rotational direction D1 with a first torque T1 leads to a rotation of the second camshaft 64B in a second rotational direction D2 with a second torque T2.

[0157] The directions of rotation D1, D2 are opposite to each other.

[0158] The planetary gear set 222, which is connected between the first camshaft 64A and the second camshaft 64B, establishes a gear ratio of 3:1 between them. Rotation of the first camshaft 64A by a certain angle of rotation consequently results in rotation of the second camshaft in the opposite direction of rotation by an angle that is one-third of the angle of rotation of the first camshaft 64A.

[0159] The first camshaft 64A is freely rotatable through 360°. The second camshaft 64B is rotatable through a maximum angle of rotation that is less than 360° and is preferably configured by rotation limiting stops 242 formed on the locking ring 236, as shown in Fig. 10. Preferably, the maximum angle of rotation of the second camshaft 64B is limited to 300° with respect to the transmission shaft assembly 202.

[0160] In Fig. 10 it can also be seen that the locking ring 236 is fixed in the circumferential direction with respect to the gear shaft arrangement 202 by means of an anti-rotation screw 244.

[0161] In Fig. 11 further shows that at least one functional stop 250 is formed on the cam body 206c for the third gear stage, which is adjacent to the cam body 206d. An unspecified axial projection of the cam body 206d forms at least one functional stop 252, wherein the maximum angle of rotation between the first camshaft 64A (which is connected in a rotationally fixed manner to the cam body 206c) and the cam body 206d is established by the functional stops 250, 252.

[0162] The cam body 206d is also preloaded against one of the functional stops 250, 252 by means of the functional spring 230.

[0163] Fig. 12 schematically shows the first camshaft 64A and the second camshaft 64B and the assignment of the twelve gear stages realized thereby to individual angles of rotation.

[0164] At 12 o'clock (0°) in Fig. 12a shows a basic position of the first camshaft 64A in which the first partial transmission gear stage 1 of the first partial transmission is engaged.

[0165] In a corresponding manner, Fig. 12b shows the second camshaft 64B, in which a first partial transmission gear stage 1 is also shown at 12 o'clock (0°).

[0166] At W1, it is shown that when the first camshaft 64A is rotated by a first angle of rotation (= W1), the second camshaft 64B is rotated by a second angle of rotation W2 due to the reversing gear or the transmission gear, which is smaller than the first angle of rotation W1. For example, with a transmission ratio of 3:1, the second angle of rotation W2 is 20° when the first angle of rotation W1 is 60°.

[0167] In Fig. 12b also shows the rotation limit of the second camshaft 64B to 300°, which is set by rotation limit stops 242.

[0168] To engage gear 2, the first camshaft 64A is rotated by a first change angle of rotation WD1, which in this case is equal to the first angle of rotation W1, i.e., 60°. The second camshaft 64B is rotated by 20°. In the first sub-transmission 64A, the gear change 1-2 means disengaging sub-transmission gear 1 (consequently, pivoting in of a positive-locking body or switching cam 142) and engaging the second sub-transmission gear (consequently, pivoting out of the associated freewheel body or switching cam 142 by means of the second cam body 206b). A further rotation of the first camshaft 64A by another 60° leads to the engagement of gear 3. A further rotation of the first camshaft 64A by another 60° leads to the engagement of gear 4.In this case, the second camshaft 64B is rotated by 20° each time, so that gear stage 4 corresponds to a rotation of the second camshaft 64B by 60° (one third of the angle of 180° by which the first camshaft 64A has been rotated).

[0169] To engage gear stage 5, the first camshaft 64A is now rotated by 180°, as shown in Fig. 12a. Here, the fourth partial transmission gear stage is disengaged and the first partial transmission gear stage of the first partial transmission 26 is re-engaged. At the same time, the first partial transmission gear stage of the second partial transmission 28 is disengaged and the second partial transmission gear stage of the second partial transmission 28 is engaged. Since the first camshaft 64A is rotated by 180°, the second camshaft 64B is rotated by 60°. This corresponds to a second alternating rotation angle WD2 of 60° in Fig. 12b.

[0170] Starting from this position, the first camshaft 64A is rotated again in 60° increments to engage the forward gears 6, 7, 8, until the first camshaft 64A has rotated 360° + 180° = 540°. The second camshaft 64B is then at 180° (540° divided by 3). A double gear shift now occurs again, in which the fourth sub-transmission gear of the first sub-transmission is disengaged and the first sub-transmission gear of the first sub-transmission 1 is engaged. At the same time, the second sub-transmission gear of the second sub-transmission 28 is disengaged and the third sub-transmission gear of the second sub-transmission is engaged, thus shifting from gear 8 of the overall transmission to gear 9. This is done on the first partial transmission 26 by turning the camshaft 64A again by 180° (see Fig. 12a), and consequently by 60° in the second partial transmission according to Fig. 12b. Consequently, the first camshaft 64A has rotated by 720° upon reaching gear stage 9 of the overall transmission, and the second camshaft 64B has rotated by 240°.

[0171] Based on this, the first camshaft 64A is again rotated further in 60° steps to engage gear stages 10, 11 and 12, respectively, up to a maximum rotation of the first camshaft 64A of 900° (corresponding to a maximum rotation of 300° of the second camshaft 64B, which is limited by the stop 242).

[0172] Downshifting occurs in the same manner. An overview of gear changes and their respective rotation angles is shown below in Table 1. Table 1 Gangwechsel Nockenwellen-Drehwinkel (TG1) bei Ende Schaltvorgang Drehwinkel W1 von TG1 inkrementell TG1 Gang Drehwinkel W2 von TG2 inkrementell TG2 Gang 1->2 60° 60° 1->2 20° 1 2->3 120° 60° 2->3 40° 1 3->4 180° 60° 3->4 60° 1 4->5 360° 180° 4->1 120° 1->2 5->6 420° 60° 1->2 140° 2 6->7 480° 60° 2->3 160° 2 7->8 540° 60° 3->4 180° 2 8->9 720° 180° 4->1 240° 2->3 9->10 780° 60° 1->2 260° 3 10->11 840° 60° 2->3 280° 3 11->12 900° 60° 3->4 300° 3

[0173] For the sake of simplicity, the above representation of the angles of rotation W1, W2 was made without taking into account the rotatability of the cam body 206d and the cam body 210 (of the second camshaft 64B).

[0174] However, this rotatability allows for so-called asymmetric camshaft timing, such that the alternating rotation angle differs during downshifts and upshifts. In other words, the rotation angle-dependent angular positions at which the shift pawls or freewheel bodies pivot out or in are different during upshifts than during downshifts.

[0175] This can be advantageous, as the first reason is that the shift pawl, in many designs, engages differently when upshifting than when downshifting. When upshifting, the principle is such that when the shift pawl of the higher gear is released, the higher and faster gear immediately takes over power transmission as soon as it engages with the shift pawl of the higher gear, i.e., the target gear.

[0176] When downshifting, however, the target gear—the lower gear—only takes over power transmission when the higher (faster) gear is engaged. The timing of the target gear's takeover of power transmission is therefore not dependent on the pawl state of the target gear (as with upshifting), but rather on the pawl state of the source gear.

[0177] This different behavior leads to the risk of only partial engagement of the shift pawl with the internal teeth of the corresponding gear when upshifting: The shift pawl is gradually released by the rotation of the camshaft and takes over power transmission as soon as it partially engages the internal teeth of the gear. How quickly the shift pawl is released depends on the angular speed of the camshaft. The probability of full or partial engagement of the shift pawl with the internal teeth of the idler gear also depends on the angular speed of the camshaft.

[0178] This problem doesn't occur when downshifting: the shift pawl of the target gear (lower gear) is already in a state where it is fully released (freewheeling state) and can therefore fully engage the internal gearing at the moment the shift pawl of the higher gear is released. Only when the higher (faster) gear is disengaged does the lower (slower) gear take over power transmission. Until this point, the shift pawl is in the freewheeling state.

[0179] In other words, with the switching arrangement disclosed here, a jump actuation of the switching pawl can be achieved despite a continuous camshaft rotation.

[0180] The need for such a mechanism is particularly evident when the camshaft is rotated continuously and less abruptly, as is the case with a shift actuator, for example. The actuator has a fixed speed, which ultimately determines the release speed of the shift pawl.

[0181] The requirement therefore increases the lower the angular speed of the camshaft. This is particularly the case when using the previously described reversing gear, which in this specific case reduces the angular speed of the second camshaft 64B in the second sub-gear 28 by a factor of 3.

[0182] In general, it is possible for such a pre-tensioning mechanism, realized by two stops, a functional spring 230 and a locking arrangement 234, to be assigned to each gear or each switchable idler gear of each partial transmission, or to any partial number thereof.

[0183] A second advantageous reason for asymmetric camshaft timing is the following: If, for example, cam bodies of different gears are not integral with the camshaft (such as cam bodies 210 or 206d), play may occur between the cam bodies due to manufacturing tolerances.

[0184] A particularly large play arises, for example, in the present construction, in which the camshaft 64B of partial transmission 2 is connected to the camshaft 64A of partial transmission 1 via a reversing gear 220.

[0185] To solve this problem, a cam body 206a or 210 of the affected gear is also preloaded in one direction.

[0186] In the present embodiment, for example, the cam body 206d of gear stage 4 is designed to be rotatable on the camshaft 64A. The gear has stops 250, 252 in both directions of rotation relative to the camshaft, which are designed to provide a defined torsional play of the cam body 206d of gear 4 relative to the camshaft 64A.

[0187] A torsion spring (leg spring) 230 ensures that the cam body 206d always rests against one of the two stops.

[0188] In the present embodiment, for example, the cam body 206d of sub-transmission gear stage 4 in sub-transmission 1 can be rotated relative to the camshaft 64A between two stops 250, 252. When shifting, for example, from gear 4 to gear 5 of a 12-speed transmission, the gear shifting takes place in sub-transmission 1 from sub-transmission gear 4 to sub-transmission gear 1 and in sub-transmission 2 from sub-transmission gear 1 to sub-transmission gear 2. For perfect function in both shifting directions, it is therefore necessary that the engagement / disengagement of sub-transmission gear stages 4 or 1 in sub-transmission 1 takes place as synchronously as possible with the engagement / disengagement of sub-transmission gear stages 1 or 2 in sub-transmission 2.

[0189] Overall, the disclosed switching arrangement can achieve almost simultaneous pivoting in and out of the switching pawls in both partial transmissions and directions of rotation, despite play between components due to the principle and manufacturing. List of reference symbols 10 Gear unit 12 bicycle frames 14 Gearbox housing 16.16' cranks 18 multi-speed gearboxes 20 Input shaft 22 Output shaft 24 chainring 26 first partial transmission 28 second partial transmission 30, 31, 32, 33, 34, 35 drive wheels 36 Countershaft 38, 39, 40, 41, 42, 43 driven wheels / first idler gears 46 Input shaft 48, 49, 50 Drive wheels / second idler gears 52, 53, 54 driven wheels 56 Spur gear set EM / 36 60 switching arrangement 62 Wave 64 camshaft 66 speed superposition gears 68 Transmission stage / first planetary gear 70 Control stage / second planetary gear 72 sun gear (68) 74 planetary gears (68) 76 planet carriers (68) 78 ring gear (68) 80 fixed reference point 82 sun gear (70) 84 planetary gears (70) 86 ring gear (70) 88 Pulley 88 110 cam gap 111 cams 132 switchable idler gear 134 external gearing 136 internal gearing 138 sliding sections 140 intervention sections 142 Freewheel body / shift pawl 144 operating section 146 storage sections 148 intervention section 150 rotation axis 152 sliding section 168 Arrow 200 switching arrangement 202 Gearbox shaft arrangement / countershaft 204 Camshaft arrangement (64A,64B) 206a / b / c / d cam body TG1 208a / b / c / d storage sections for 142 TG1 110a-d cam gaps 206 111a-d cam 206 210 Cam body TG2 110e / f / g cam gaps 210 (TG2) 111a / f / g cam 210 (TG2) 212e / f / g storage sections for 142 TG2 110e-g cam gaps 210 111e-g cam 210 214 storage section 220 reversing gear 222 planetary gear set 224 Sun gear 226 planet carrier 228 ring gear 230 Function spring for 206d 232 Function spring for 210 234 locking arrangement 236 locking ring 238 locking bodies 240 rest points 242 anti-twist stops 244 Anti-rotation screw 250 first function stop 252 second function stop 254 cam surface Alpha flank angle 136 Beta pressure angle 148 E1 first axial end 204 E2 second axial end 204 A axial section D1 first direction of rotation 64A D2 second direction of rotation 64B W1 first angle of rotation 64A W2 second rotation angle 64B T1 first torque 64A T2 second torque 64B WD1 first alternating angle of rotation 64A WD2 second alternating angle of rotation 64B EM electric motor QUOTES CONTAINED IN THE DESCRIPTION

[0000] This list of documents submitted by the applicant was generated automatically and is included solely for the convenience of the reader. This list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions. Cited patent literature

[0000] EP 2 379 402 B2 [0004, 0006, 0012, 0030] WO 2019 / 166563 A1

[0007]

Claims

[1] Shifting arrangement (200) for a vehicle transmission (18), in particular for a transmission of a vehicle driven by muscle power and / or motor power, with a transmission shaft arrangement (202) on which a plurality of idler gears (132) are rotatably mounted, which are connectable to the transmission shaft arrangement (202) via respective shifting means (142), wherein the shifting means (142) are actuated by means of a camshaft arrangement (204) having a first camshaft, characterized by in that the first camshaft (64A) and a second camshaft (64B) of the camshaft arrangement (204) are coupled to one another via a reversing gear (220) such that a rotation of one of the camshafts in a first direction of rotation (D1) leads to a rotation of the other camshaft in a second direction of rotation (D2) which is opposite to the first direction of rotation (D1). [2] Switching arrangement according to claim 1, with - the transmission shaft arrangement (202), (i) via which a drive force can be transmitted, (ii) on which a plurality of first idler gears (41) of a first partial transmission (26) and a plurality of second idler gears (48) of a second partial transmission (28) are rotatably mounted, which serve to set up different gear stages, and (iii) which is designed at least in sections as a hollow shaft, wherein the camshaft arrangement (204) is preferably rotatably mounted at least in sections in the transmission shaft arrangement (202) and has the first camshaft (64A) assigned to the first idler gears (41) and the second camshaft (64B) assigned to the second idler gears (48). [3] Switching arrangement according to claim 1 or 2, characterized bythat the reversing gear (220) has a planetary gear set (222), wherein it is preferred if a first member (224) of the planetary gear set is coupled to the first camshaft (64A), a second member (226) of the planetary gear set is coupled to the transmission shaft arrangement (202), and a third member (228) of the planetary gear set is coupled to the second camshaft (64B). [4] Shifting arrangement according to one of claims 2-3, with a speed superposition gear (66) via which the transmission shaft arrangement (202) and the camshaft arrangement (204) are coupled and which can be actuated in order to achieve a relative rotation of the transmission shaft arrangement (202) and the camshaft arrangement (204) for switching the gear stages, wherein the speed superposition gear (66) is preferably coupled to the transmission shaft arrangement (202) and to one (64A) of the first and second camshafts (64A, 64B). [5] Switching arrangement according to one of claims 1-4, characterized by , that - the speed superposition gear (66) is arranged at a first axial end (E1) of the camshaft arrangement (204), and / or - the reversing gear (220) is arranged at a second axial end (E1) of the camshaft arrangement (204), and / or - the reversing gear (220) is arranged radially inside the gear shaft arrangement (202), and / or - the first partial transmission (26) is arranged axially adjacent to a first axial end (E1) of the camshaft arrangement (204), and / or - the second partial transmission (28) is arranged axially adjacent to a second axial end (E2) of the camshaft arrangement (204). [6] Switching arrangement according to one of claims 1-5 or according to the preamble of claim 1, characterized by that the second camshaft (64B) is designed as a hollow shaft and is rotatably mounted on an axial section (A) of the first camshaft (64A). [7] Switching arrangement according to one of claims 1-6 or according to the preamble of claim 1, characterized by in that the first camshaft (64A) and the second camshaft (64B) are coupled to one another via a transmission gear (220) such that a rotation of one of the camshafts by a first angle of rotation (W1) leads to a rotation of the other camshaft by a second angle of rotation (W2) which differs from the first angle of rotation (W1). [8] Switching arrangement according to claim 7, characterized by , that - the transmission gear is designed such that the second angle of rotation (W2) is smaller than the first angle of rotation (W2) and / or - the transmission gear (220) has a transmission ratio in a range of 2 to 4, preferably a transmission ratio of 3 and / or - the second angle of rotation (W2) is equal to the first angle of rotation (W1) divided by the ratio of the transmission gear (220). [9] Switching arrangement according to one of claims 1-8, characterized by that the first camshaft (64A) is rotatable through 360° and the second camshaft (64B) is rotatable over an angular range that is smaller than 360°. [10] Switching arrangement according to one of claims 1-9 or according to the preamble of claim 1, characterized by that each gear change in the first partial transmission (26) is assigned a respective first change angle of rotation of the first camshaft (64A), wherein each gear change in the second partial transmission (28) is assigned a respective second change angle of rotation of the second camshaft (64B), wherein the first change angle of rotation (WD1) of at least one gear change in the first partial transmission (26) is equal to the second change angle of rotation (WD2) of at least one gear change in the second partial transmission (28). [11] Switching arrangement according to one of claims 1-10 or according to the preamble of claim 1, characterized byin that the first camshaft (64A) and the second camshaft (64B) are coupled to one another via a torque change gear (220) such that a rotation of one of the camshafts with a first torque (T1) leads to a rotation of the other camshaft with a second torque (T2) which differs from the first torque (T1). [12] Switching arrangement according to claim 1 and according to claim 7 and / or 11, characterized by that the transmission gear is formed by the reversing gear (220) and / or wherein the torque change gear is formed by the reversing gear (220). [13] Switching arrangement according to one of claims 1-12 or according to the preamble of claim 1, characterized byin that at least one switching means is designed as a switching pawl (142) which is mounted radially pivotably on the transmission shaft arrangement (202) and which can be actuated between a pivoting position and a gear position by means of a functional cam body (206d; 210) of the camshaft arrangement (202), wherein the functional cam body (206d; 210) is designed as a cam body (206d; 210) which is mounted rotatably between a first functional stop (250) and a second functional stop (252). [14] Switching arrangement according to claim 13, characterized by , that - the functional cam body (206d; 210) is pre-tensioned against the first functional stop (250) by means of a functional spring (230; 232) such that it remains on the first functional stop (250) during downshifts, and / or - the functional cam body (206d; 210) is rotated against the action of a functional spring (230; 232) during upshifts. [15] Switching arrangement according to claim 13 or 14, characterized by that the angle between the first functional stop (250) and the second functional stop (252) - is selected so that it compensates for at least one circumferential play that exists between the functional cam body (206d; 210) and the associated camshaft (64B; 64A), and / or - is adapted to a relative rotatability between the functional cam body (206d; 210) and the associated camshaft (64B; 64A), and / or - is in a range of 5° to 20° or in a range of 50° to 80° and / or - is selected in such a way that engagement of a target gear is possible independently of a shift contour. [16] Switching arrangement according to one of claims 13-15, characterized by that the functional cam body (206d) is assigned to the highest gear stage of the first partial transmission (26). [17] Switching arrangement according to one of claims 1-16 or according to the preamble of claim 1, characterized by in that the second camshaft (64B) is mounted so as to be rotatable to a limited extent with respect to the first camshaft (64A) between a first coupling stop and a second coupling stop and is prestressed against the first coupling stop by means of a coupling spring (232), wherein the second camshaft (64B) is preferably held in a locking manner in a plurality of locking positions (240) with respect to the transmission shaft arrangement (202) by means of a locking arrangement (234). [18] Transmission unit (10) with a switching arrangement (200) according to one of claims 1 -17. [19] Method for connecting two rotary members (202, 41; 202, 48) which are rotatable relative to one another and can be connected by means of a switching means (142), while they rotate relative to one another, by means of a camshaft which has a functional cam body (206d; 210) which (i) is mounted on the camshaft so as to be rotatable relative to a basic position and which (ii) is designed to actuate the switching means (142), wherein the camshaft arrangement (204) is rotated to connect the rotary members (202, 41; 202, 48) which are rotating relative to one another, but the functional cam body (206d; 210) is initially not rotated along with the camshaft arrangement until it is forcibly rotated along with the camshaft arrangement from a certain rotational position of the camshaft arrangement (204) and is then rotated back into the basic position. [20] Method according to claim 19, wherein the functional cam body is initially not rotated during rotation of the camshaft arrangement by being held in place relative to one of the rotary members. [21] Method according to claim 19, wherein the functional cam body is initially not rotated during rotation of the camshaft arrangement by being held fixed relative to the switching means. [22] Method for performing gear changes by means of a shift arrangement (200) according to one of claims 1-17, comprising the step of rotating the camshaft arrangement (204) by a change angle of rotation (WD1) in order to disengage a source gear stage and to engage a target gear stage.

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