DE10030941A1 - Multiple gear shift hub has control cam carrier unit with cam turning round geometric axis at angle to plane normal to axis - Google Patents

Abstract

The hub has at least one planetary transmission with gear ratio selecting sunwheels and other free sunwheels. An axially displaceable sunwheel state changer is move by cams in a cam carrier unit with cam resting elements which can be adjusted radially by the use of a gear changing signal line to engage with and disengage from one of the cams.

Description

The invention relates to a multi-speed gear hub for bicycles or the like, comprising:

• - a hub axis with a geometric axis,
• One with at least one torque introduction element, in particular a sprocket, connectable or coupled Driver, which is rotatably mounted on the hub axle,
• - a hub sleeve rotatably mounted on the hub axle,
• - A planetary gear system arranged within the hub sleeve with at least one planetary gear for torque transmission from the driver to the hub sleeve, this planetary gear system executed with
  at least one sun gear, which can optionally be switched between a state of free rotation about the hub axis and a state of rotation blocking with respect to the hub axis,
  at least one ring gear rotatably mounted on the hub axle and
  at least one planet gear, which is rotatably mounted on a planet gear carrier rotatable about the hub axis,
  also comprehensive
  an axially movable sun gear condition converter which, depending on its axial position, the sun gear
  either causing the state to rotate freely or the state of rotation blocking,
  a gear change device for moving the sun gear state converter in the direction of the geometric axis, this gear change device in operative connection
  with a gear change signal line inserted into the multi-gear hub from a gear change signal transmitter outside the multi-gear shift hub.

Such a multi-speed shift hub is known from EP 446 966 B1. According to FIG. 9 of this document, a blocking sleeve on the hub axle is arranged in a non-rotatable but axially displaceable manner as a sun gear condition converter. Depending on the axial position, this locking sleeve is in rotationally locking engagement with either a small or a large sun gear. The axial displacement of the blocking sleeve takes place from a gear change signal transmitter, ie a gear switch, which is arranged outside the multi-gear shift hub, for example on a handlebar of the respective bicycle. By optionally determining or releasing the small or large sun gear, the transmission ratio of the planetary gear can be changed. The axial displacement of the blocking sleeve from the gear change signal transmitter must be coordinated very precisely so that the sun gears are blocked individually with maximum tooth coverage to the hub axis. Changes in length in the signal line connecting the gear change signal generator to the blocking sleeve can lead to malfunctions within the multi-speed shift hub. If an attempt is made to axially shift the blocking sleeve for the purpose of gear shifting as long as the multi-speed shifting hub is under load, i.e. as long as the pedal is in the forward direction of rotation, then large support moments occur between the hub axle and the respectively effective sun gear, the correspondingly large tooth flank friction between the respective sun gear and the hub axle. This means that the blocking sleeve can only be moved in the axial direction with great effort. and accordingly large switching forces occur on the gear change signal generator and the gear change signal line. These large forces lead to length changes in the gear change signal line, so that the aforementioned precision of the locking sleeve setting can be lost quickly.

The invention has for its object a multi-speed hub trained in such a way that the sun gear state converter (comparable to the blocking sleeve in EP 446 966 B1) increased precision can be positioned and held in position; further gear shifting by changing the switching state of one or several sun gears by means of axial displacement of the sun gear state converter be facilitated; in other words: the one on the gear change signal generator and the gear change signal line occurring switching forces should also be kept low when gear shifting under load, d. H. while stepping forward.

The following features are proposed to solve this task:

• a) a cam carrier unit rotatable about the geometric axis with at least one in the circumferential direction around the geometric axis extending and inclined to a plane normal to the axis Cam;
• b) a rotary driving connection between one when the Driver or / and when rotating the hub sleeve around the geometric Axis rotating hub component and the cam carrier unit;
• c) essentially at least one cam system element axial contact with the control curve,
• d) wherein this cam system element relative to the hub axis essentially in the direction of the hub axis and in the circumferential direction immovable and radially adjustable with respect to the geometric axis is between a radially overlapping position to the control curve and a radially non-overlapping position to the control curve,  
• e) wherein further the gear change signal line for the radial adjustment of the Control cam system element is operatively connected to this,
• f) wherein further the cam carrier unit with the sun gear stand converter is identical or in axial driving connection with the Sun gear state converter stands,
• g) and the geometric relationships are chosen such that according to radial adjustment of the cam system element in its Control curve radially overlapping position the control curve as a result of Rotation of the cam carrier unit around the geometric axis in Can come into contact with the control system element and the Control cam carrier unit and the sun gear state converter as a result Relative rotation of the control curve and the control curve system element experience an axial displacement.

In the inventive design of the multi-speed hub, the Axial position of the cam carrier unit and thus the sun gear stand converter essentially depends on the interaction a cam and a cam system element. Because the tax rate Venanlagelement in the circumferential direction and in the axial direction of the hub axis in Is essentially immovable results in a certain rotation kelstellung the cam carrier unit a clear axial definition this cam carrier unit and thus the sun gear state converter, which is largely independent of the length setting of the aisle AC signal line is. This results in the required precision in terms of the blocking and rotation release states of the at least one Sun.

Because the axial displacement of the sun gear state converter from the rotation the driver or the rotation of the hub sleeve can also be derived when shifting gears under load no large shifting forces on the Gear change signal generator, d. H. on the hand-operated corridor switch and in the gear change signal line. Gear shifting under load  is therefore relieved. At the same time, there is a risk of a change in length in the Gear change signal line reduced when shifting under load. this applies not only if the gear change signal line is purely mechanical from the Gear switch leads to the sun gear state converter, but also when for example an electrical gear change signal line for use comes, which also go somewhere in a mechanical encoder got to.

Further features of the invention emerge from the subclaims.

In a first embodiment of the invention it is provided that the Control cam assembly of the control cam carrier unit and the system elements Equipment of the hub axle for the axial displacement of the control cure vträgereinheit are formed in a single direction and that a Reset device for axial displacement of the cam carrier unit in the opposite direction is provided. In this embodiment it may prove necessary in an axial position caused by Interaction between the control cam and the cam system element is achieved will provide a holding device which the reset device, for example a return spring that withstands and is rendered ineffective, when the axial displacement is initiated in the opposite direction becomes.

According to a second embodiment it is provided that the control curves Equipping the control cam carrier unit and the system element equipment the hub axis for the axial displacement of the cam carrier unit in opposite directions are formed.

In the simplest case, the combination of a control curve and one is used Control cam system element for securing against rotation or for enabling a rotation Sun gear or for optional rotation lock or rotation release in each case one of two sun gears. In addition, it is also possible that  the control cam assembly of the control cam unit and Equipment of the hub axle for setting more than two axial positions of the cam carrier unit determined and suitable are.

The axial adjustment of the cam carrier unit or the sun gear level converter in two opposite axial directions can ins special to be set up in such a way that defining the tax rate Carrier unit in a certain axial position on the control cam Carrier unit arranged two axially opposite control cams are and that each of these axially opposite control curves at least one radially overlapping cam system element in the axial Facing direction.

To ensure that the cam carrier unit reached one Axial position according to the selection of a specific sun gear can also be maintained until another switching signal is given ensure that each of the control cams belongs to the axis normal Flat envelope circle axial distance from that of the respective Control cam in the axial direction opposite control cam system ment or affects it. In the event of the registration of a axial distance this axial distance is preferably kept small, so that Adjustment game of the cam carrier unit and accordingly also the Remains as small as possible.

Provided that two axially opposite Control cams are provided on the control cam carrier unit, one can ensure that the cam carrier unit is defined in a certain axial position each assigned to this axial position Control system element or a group of such elements in one Circumferential zone between two facing each other in the axial direction Control curves lie or lie and these control curves overlap radially  or overlap. This embodiment can be used e.g. B. ver real that the two facing control curves from the Limiting curves of a cam groove on an outer peripheral surface of the Cam carrier unit takes.

Alternatively, you can also provide that to determine the control curves Carrier unit in a certain axial position two of each other control cams one of two determined this axial position control cam system elements or one of two this axial position defining groups of cam system elements in axial Direction is opposite and the respective control curve radially overlaps. In this variant, the two can face away from each other Screw control curves, for example, from the limitation curves of a screw be formed rib extending from the peripheral surface of a Cam carrier unit protrudes out.

The switching mechanism can be designed so that with existing Determination of the cam carrier unit in a specific axial position a control curve to initiate the change in this axial position system element through the gear change signal line in the direction of overlapping position can be applied to an associated control curve such that it - depending on the current angle of rotation of the control cams Carrier unit - immediately or after turning the cam carrier unit in overlapping position to the associated control curve, with this in Engagement occurs and the cam carrier unit in its further rotation in axially shifts.

In particular, it is possible that the existing tax rate Carrier unit in a first specific axial position for induction another specific axial position at least one further control course Venanlageelement axially behind an envelope through the gear change signal line towards the radially overlapping position of this one  Envelope associated with a control curve can be acted on such that it - depending after the current angle of rotation position of the cam carrier unit - immediately or after rotation of the cam carrier unit in one compared to the a circumferential area set back from the one enveloping circle associated control curve in a radially overlapping position to this one Control curve arrives and immediately or after further rotation of the control curves Carrier unit engages with this one control cam and after another Rotation of the cam carrier unit towards the other certain axial position shifts axially. In this embodiment, can further be provided that during or after the axial displacement of the cam carrier unit in the direction of the other specific axial position tion involved in determining the first determined axial position Cam system element by interacting with the other Control curve or / and through the gear change signal line towards the radially non-overlapping position can be reset to the other control curve is.

The interaction of two control curves with this associated tax Curve system elements can lead to situations in which ver different axial positions of the cam carrier unit associated Control system elements are in engagement with the respective control curve. In order to prevent this from jamming, it is provided that the axial distance between two cams on their whole Circumferential length is dimensioned such that during the rotation of the Cam carrier an unintended simultaneous radial Overlap by cam elements under oscillating Allow axial movement of the cam carrier unit without jamming Cam system elements that are axially adjacent for two specific ones Axial positions of the cam carrier unit are responsible. At a Such a design can never result in a pinch by being together act on the cam carrier unit with the cam system elements  come. The worst thing that can happen is that the control curves Carrier unit performs an oscillating axial movement. This is an instant recognizable fault condition that cannot lead to destruction.

It is possible that a cam system element in a radial bore the hub axis is guided radially adjustable. This can be a control curve plant element to be formed by a ball. Are next to a ball for example, cam elements in the form of pins conceivable, which are spherically or conically tapered at at least one end.

A production-technically advantageous design of the control cam carrier unit arises when a control curve of a shoulder between Inner circumferential surfaces offset in relief in the radial direction the cam carrier unit is formed. The cam carrier unit can then machined on the inner circumferential surfaces. This milling can be facilitated by the fact that Control carrier unit is composed of two parts, the other each one of the two control curves is worked out.

The gear change signal line can in different ways on the Control system elements act, for. B. so that the gear change signal line is executed or connected to a control body, which in an axial cavity of the hub axle is axially displaceable and an in Axial direction control profile for support and / or radial ver push a cam system element. It is also conceivable that the cam system elements using a electrical gear change signal line electromagnetic or other Way to be acted upon.

To prevent the cam system elements between the respective control curve and a radially inward control surface crushed or damaged, one can ensure that a  Control curve system element when acted upon by a control curve against a radially inward movement at least in part by a peripheral wall of the radial bore is supported. With this out leadership form there are favorable osculation between the respective cam system element and its hub axle Support.

There are a number of options for the construction of the hub construction then when the axially movable sun gear state converter from one with the cam carrier unit by a rotary connection to the common axial movement connected coupling member is formed, which is rotationally fixed, but is axially displaceable on the hub axle and in Depending on its axial position, a sun gear in the free state Keeps rotatable or in the state of rotation blocking. The coupling link can in this embodiment namely slim, for. B. as a coupling sleeve, be trained and thus without much space Bridging large longitudinal distances between the sun gear and steering wheel carrier unit can be used.

To avoid the risk of overloading in the multi-gear when switching To avoid shift hub, it is provided that the gear change signal line contains an elastic compensating member, which initiates a Gear change signal allowed even when an intended radial Adjustment of the cam system element by the control cam carrier unit is inhibited.

The resetting of a cam system element in the radial direction can basically be left to gravity if the leadership Direction of the cam system element is vertical. But since with Consider inhibitions or delays in the cam system element grease, for example, it is recommended that the minimum a control curve on at least part of its circumferential course  Profile shape and the at least one cam system element an outline possess such that through the interaction of this profile shape and this Outline a shift of the cam system elements radially inwards is supported.

When a planetary gearbox changes gear due to momentary rotation Release of all sun gears in the sense of torque transmission is not functional and if the drive of the cam carrier unit is derived from a hub component, which in such Intermediate state is not driven, this can lead to faults that are based on the fact that the cam carrier unit as a result of current Interruption of your drive not over the necessary axial distance is moved. In order to remedy this, it is envisaged that the rotating hub component for driving the cam carrier unit driven by the driver via the planetary gear system, the torque flow from the driver to the rotating hub comm component in the course of switching a sun gear between the State of free rotatability and the state of rotation blocking under is broken, and that a friction clutch is provided to in the event any interruption the drive of the rotating hub component maintain.

The planetary gear system can be of a particularly simple design be constructed in accordance with EP 446 966 so that the planetary gear system a single planetary gear with a multi - stage planet gear and a number of corresponding to the number of stages of the planet gear Includes sun gears, each one of the sun gears in the state the rotation lock and the other sun gears in the state more freely Rotatability. It can be done in a conventional manner, such as from the EP 446 966 also known, by different guidance of the torque achieved an increase in the number of gears via the only planetary gear be, e.g. B. by the fact that the driver with either Planet carrier and can be coupled to the ring gear for common rotation  is and that the hub shell optionally with the ring gear and the planet wheel carrier is in the output connection.

In such an embodiment, the cam carrier unit on the the side of the planet carrier close to the driver, and it can then the cam carrier unit in rotary driving connection with the Drivers stand.

Alternatively, it is also possible that the control cam carrier unit on the is arranged remote from the driver side of the planet carrier. In In this case, it is in particular possible for the control cam carrier unit is in rotary driving connection with the planet carrier. The Availability of these different options allows it to be optimal To create and use space requirements.

When shifting the cam carrier unit and the sun gear stand converter inhibitions can occur, for. B. in that Kupp teeth of the sun gear condition converter and the sun gears corresponding relative angle setting at the moment of moving the Collide the sun gear state converter at the end. To in such Avoid situations of overloads in the gear change signal line, it is proposed that the rotary driving connection between the rotating hub component and the cam carrier unit as one Friction connection or overload clutch is formed. This execution is also helpful because when shifting under a heavy load that Moving the sun gear state converter can be delayed until one Load reduction occurs and the necessary axial force to move the Sun gear condition converter is thereby lower. Then it can temporarily a rotational movement in the friction connection or Overload clutch take place.

Exploiting the planetary gear to increase the number of available Gears by optionally introducing the drive into a planet carrier  and derivation of the output from the ring gear can be done in such a way that from the driver either the planet carrier or the ring gear an axially displaceable coupling sleeve is driven. It is also possible that the driver through a ring gear coupling with the Ring gear and through a planet carrier clutch with the Planet carrier is connectable, the ring gear clutch and the Planet carrier drive clutch for torque transmission to the Ring gear or the planet gear carrier is formed only in the forward direction of rotation are and wherein the planet carrier clutch is switched off. It is it is then also possible that the hub sleeve from the ring gear and from Planet gear carrier forth through a ring gear driven clutch or planetary carrier-side output clutch can be driven, the ring gear side Output clutch and the planet carrier-side output clutch for torque ment transmission from the ring gear and the planet carrier to the hubs sleeve are each formed only in the forward direction of rotation and the ring gear-side output clutch can be switched off.

The multi-speed shift hub according to the invention can be equipped with a coaster brake be equipped. When applying a coaster brake is sometimes wished that the withdrawal braking distance regardless of the respective selected gear is always the same. This requirement can be met if for torque transmission from the driver to the planet carrier and from the driver to the ring gear and for torque transmission from the Planet gear carrier to the hub shell and from the ring gear to the hub shell in Forward one-way effective one-way clutches are provided and being used to transmit a brake activation torque to the brake set up a further one-way coupling effective in the reverse direction is provided. With such training, problems can arise blocking of the multi-speed gear hub occur when braked and / or when pushing backwards. This can solve these problems that at least one of the one-way clutches is prevented by Switch-off means controlled by a reverse rotation sensor system can be switched off.

The accompanying figures explain the invention using an embodiment Example. They represent:

Fig. 1 shows a longitudinal section through a first embodiment of a Mehrgangschaltnabe invention for the purpose of overview and interface registration;

Fig. 2 is an enlargement to Fig. 1;

Fig. 3 is a section along line III-III of Fig. 1;

Fig. 4 is a section along line IV-IV of Fig. 1;

Fig. 5 is a section along line VV of Fig. 1;

Fig. 6 is a section along line VI-VI of Fig. 1;

Fig. 7 is an enlargement to Figure 2, limited to the area of the cam carrier unit and the control elements interacting with the control cam of the control cam unit.

Figs. 8a-8c are schematic developments of the control curve carrier unit in different operating conditions;

Fig. 9 shows a detail of the planet carrier;

Fig. 10 shows a further detail of the planet carrier;

Fig. 11 in detail a coupling sleeve for turning off the hohlradsei term output coupling from the transfer switch off;

Fig. 12 in detail a circuit ring to the Hohlradantriebskupplung in backward sliding and intermittent brakes off;

Fig. 13 is a processing corresponding to Figure 8a in a control carrier unit with modified control curves.

FIG. 14 shows a longitudinal section corresponding to FIG. 1 in a second embodiment of the invention;

FIG. 15a, the problem of blocking the hub when reversing slide;

FIG. 15B, the problem of blocking during braking hub;

FIG. 16a is a solution to the problems according to Fig 15a and 15b.

Fig. 16b a state I when reversing slide in the case of a solution according to Fig. 16a;

Fig. 16c a state II when reversing slide in the case of a solution according to Fig. 16a;

Fig. 16d a state III when reversing slide in the case of a solution according to Fig. 16a;

Fig. 17a shows a state IV when braking in the event of a solution accordingly Fig. 16a and

Fig. 17b shows a state V when braking in the event of a solution accordingly Fig. 16a.

Fig. 1 shows a 7-speed gear hub according to the invention in a half-section overview, the sections described below are shown in detail.

For ease of description, the functionally essential area of the 7-speed gear hub according to FIG. 1 is shown enlarged in FIG. 2, namely in 7th gear or high overdrive, that is the gear ratio with the highest ratio, which is also called high overdrive.

One can see a hub axle 10 with a geometric axis A. On the hub axle 10 , a bearing ring 11 and a bearing ring 12 are fastened by screwing. On the bearing ring 11 , a driver part 13 n is rotatably mounted by means of a ball bearing 15 , which is formed with profiles 14 for attaching at least one sprocket. On the bearing ring 12 on the one hand and the Antreiberteil 13 n on the other hand are bearing insert rings 16 a, 16 b made of steel consisting of a light metal hub sleeve 16 by means of ball bearings 17 and 18 rotatably mounted. This hub sleeve 16 is drawn interrupted to simplify the illustration. A planetary gear system PGS with - in the example - a single planetary gear 19 is accommodated within the hub sleeve 16 . This planetary gear 19 comprises a planet carrier 20 a-bc (see also FIGS. 9 and 10), which is composed of a bearing neck 20 a, an intermediate ring 20 b integrally connected to the bearing neck 20 a and a front ring 20 c. Furthermore, the planetary gear 19 comprises a ring gear 21 and sun gears 22 , 23 , 24 , one of which is supported against rotation on the hub axle 10 . In the case of Fig. 2, the sun gear 24 is supported on the hub axle 10 against rotation. The means for selectively supporting one of the sun gears 22 , 23 , 24 will be discussed later. On the planet gear carrier 20 a-bc 25 planet gear sets 26-27-28 are rotatably supported by means of planet bearing bolts . Each planetary gear set 26-27-28 consists of three planet gears 26 , 27 and 28 connected for common rotation about the respective planet bearing pin 25 , of which the planet gear 26 has the largest number of teeth and the planet gear 28 has the smallest number of teeth. The planet gears 26 , 27 , 28 mesh with the sun gears 22, 23 and 24 . In addition, the planet gear 27 meshes with an internal toothing 29 of the ring gear 21 . On the planet carrier 20 a-bc, namely on the bearing neck 20 a of this planet carrier, a pawl carrier 30 for rotational driving - and with rotational play - is arranged, which carries an output pawl set 31 on the planet carrier side. On the function of the rotational play between the bearing neck 20 a and the pawl carrier 30 will be discussed in detail later in the description of a braking device 39 . The planet carrier-side output-latch set 31, together with a ratchet tooth 32 of hub sleeve 16 and that of the bearing insert ring 16 a a planet carrier-side output coupling 31-32 to the drive transmission from the planet gear carrier 20 a-bc on the hub sleeve sixteenth As evident from the section shown in Fig. 5, the pawls of the ratchet set 31 and the ratchet teeth 32 are shut each other such that the direction of rotation for driving the hub sleeve 16 in the forward VD the pawls of the ratchet set 31 in the pawl teeth 32 drivingly engage and that in Standstill of the planet carrier 20 a-bc the pawl toothing 32 of the hub sleeve 16 rotating in the forward direction of rotation VD slides over the pawls of the output pawl set 31 , ie the planet carrier-side output clutch 31-32 is overhauled.

On the ring gear 21 is a pawl-bearing pin 33, a ring gear-side output pawl set 34 is mounted, which forms a ring gear-side output clutch 34-35 together with a ratchet toothing 35 of the hub sleeve 16 . If the pawls of the pawl set 34 engage in the pawl teeth 35 of the bearing insert ring 16 b of the hub sleeve 16 and the ring gear 21 is driven in the forward direction of rotation VD, the hub sleeve 16 is carried along by the ring gear 21 in the forward direction of rotation. When the ring gear 21 is stationary and the hub sleeve 16 rotates in the forward direction of rotation VD, the pawl toothing 35 of the hub sleeve 16 slides over the pawls of the output pawl set 34 on the ring gear side. This can also be seen from the sectional view according to FIG. 3.

The drive of the 7-speed shift hub is initiated from a drive chain via a sprocket, more generally via a torque input member DEG in the sprocket-closer drive part 13 n of a driver 13 . With the driver part 13 n a pawl carrier 36 a- 36 b is connected as a sprocket-remote driver part for common rotation, which consists of the two integral parts 36 a and 36 b. With this pawl carrier 36 a- 36 b, a switching ring 36 c shown in FIG. 12 in a perspective view of an individual part is in a play-free rotary driving connection; the function of this switching ring 36 c will be explained later in connection with the operating states pushing back and braking. Arranged on the pawl carrier 36 a- 36 b is a planetary pawl set 37 which is in driving engagement with a ratchet toothing 38 of the planet gear carrier 20 a-bc and together forms a planet carrier drive clutch 37-38 . The planetary pawl pawl 37 takes the planet carrier 20 a-bc in the forward direction of rotation VD when the driver part 13 n is driven in the forward direction of rotation VD. The switching ring 36 c has on its inner circumferential surface radially inward projections 70 ( FIG. 12), via which it is in rotary connection with the part 36 b of the pawl carrier 36 a- 36 b. Furthermore, the switching ring 36 c has axial features 71 , the function of which will be explained later.

In the switching state of the seventh gear shown in FIG. 2, the following torque flow occurs when the driver part 13 n rotates in the forward direction of rotation:
The driver part 13 n takes the planet carrier 20 a-bc in the forward direction of rotation VD via the pawl carrier 36 a- 36 b and the planet carrier-carrying pawl set 37 and the pawl toothing 38 . The small planet gear 28 rolls on the large sun gear 24 , which is the only one of the sun gears 22 , 23 , 24 blocked against rotation about the hub axis 10 . The planetary gear set 26-27-28 rolls with the medium-sized planet gear 27 on the internal toothing 29 of the ring gear 21 . The sun gears 22 and 23 can rotate freely about the hub axis 10 . The ring gear 21 is driven in the forward direction of rotation at a speed which is greater than the speed of the driver 13 and the planet carrier 20 a-bc. The speed of the ring gear 21 is transmitted to the hub sleeve 16 via the ring gear-side output pawl set 34 and the pawl toothing 35 , which together form a ring gear-side output clutch 34 , 35 . The hub sleeve 16 thus rotates at a speed which is translated into rapid in relation to the speed of the driver 13 . Since the large sun gear 24 and the small planet gear 28 are involved in the translation effect of the planetary gear 19 , it is the state of the largest translation. One speaks of the great overdrive.

When - again in the switching state of the seventh-speed or large fast transition according to Fig 2 - of the impeller 13 still and the hub sleeve 16 in the forward rotational direction VD turns (idle), then also the planet gear carrier 20 a-bc and the ring gear 21 of the. Planetary gear 19 silent, and the hub sleeve 16 slides with the pawl teeth 35 over the ring gear-side output pawl set 34 .

At this point it should be pointed out that between the driver part 13 n and the ring gear 21 there is another pawl set 44 , namely a pawl set 44 driving the hollow wheel, as also shown in section in FIG. 6. This pawl set 44 sits on the driver part 13 n. Its pawls engage in a pawl toothing 45 of the ring gear 21 and, together with the pawl toothing 45, form a ring gear drive clutch 44 , 45 (in FIG. 6 only with a dashed line via an actually not in the cutting plane VI- VI indicated ring gear cutout). In the state of the 7th gear or high overdrive described here, the ring gear pawl set 44 is not involved in the torque transmission. The meaning of this set of hollow wheel driving locking links will emerge in the further course of the description in connection with the discussion of a direct gear and the gear ratios 3, 2 and 1. It is sufficient to point out that in 7th gear or high overdrive when the hub sleeve 16 is driven by the driver 13 via the planet gear carrier 20 a-bc, the ring gear 21 rotates at a higher speed than the driver 13 and the planet gear carrier 20 a-bc, so that the pawl teeth 45 slide over the hollow wheel-driving pawl set 44 .

For the purpose of braking, a coaster brake 39 is provided to the left of the planetary gear 19 . The coaster brake 39 (see also FIG. 4) comprises an expandable brake jacket 40 which can be brought into frictional contact against a braking surface 41 of the hub sleeve 16 . The expandable brake jacket 40 is secured on the bearing ring 12 against rotation about the hub axis 10 , for example by means of a brake lever (not shown) which is attached to the bearing ring 12 and is supported on the frame of the respective bicycle against rotation. The impeller 13 is rotated for braking in reverse rotational direction RD when in the switching state of the 7th passage or large fast transition according to Fig. 2, so 13 accommodates the impeller on the pawl carrier 36 a-36 b by means of another on the pawl carrier 36 a-36 b arranged , brake actuating pawl set 42 (see also FIG. 3) with the planetary gear carrier 20 a-bc in the reverse direction of rotation RD, so that the bearing neck 20 a of the planetary gear carrier 20 a-bc and the pawl carrier 30 of the planetary carrier-side output pawl set 31 in the reverse direction of rotation RD be rotated. The pawl set 42 together with the pawl teeth 38 forms a further clutch, namely a brake activation clutch 42-38 .

As can be seen in detail in Fig. 4, 40 pinch rollers 43 are arranged between the bearing shank 20 a and the brake jacket, which are positioned by a pinch roller cage 30 a made in one piece with the pawl carrier 30 . The bearing collar 20 a is provided in the region of the pinch rollers 43 d, on its outer peripheral surface with riser surfaces 20, which the pinch rollers 43 press upon pivotal movement of the bearing neck 20 a in reverse rotational direction RD against the inner circumferential surface 40 a of the expandable brake shell 40 and spread open the brake casing 40 so that the hub sleeve 16 is braked by the expanding brake jacket 40 . The braking process is not hindered by the planet carrier-side output pawl set 31 and the ring gear-side output pawl set 34 , since the pawl teeth 32 and 35 can slide freely over the planet carrier-side output pawl set 31 or the ring gear-side output pawl set 34 .

It should also be pointed out that — as already mentioned above — the pawl carrier 30 and the bearing neck 20 a are mutually rotationally connected under a rotational play ã (see FIG. 5). This rotational play ã is provided by corresponding teeth on the outer circumferential surface of the bearing neck 20 a and on the inner circumferential surface of the pawl carrier 30 . In the drive state, that is, when torque is introduced via the driver 13 in the forward direction of rotation VD, the corresponding toothings are in contact with one another via driving tooth flanks 20 e or 30 b facing one another, as shown in FIG. 5, so that torque transmission from the bearing neck 20 a to the Pawl carrier 30 can be done. During braking, however, the bearing neck 20 a under partial "exhaustion" of the rotary play A is rotated relative to the pawl carrier 30, the pinch rollers 43 faces via the riser are pressed d 20 against the brake sleeve 40 and spreading it under braking of the hub sleeve sixteenth The slope of the rising surfaces 20 d is dimensioned such that the rotational play ã is not completely used up during braking. With renewed rotary drive via the driver 13 in the forward direction of rotation VD, the bearing neck 20 a and the pawl carrier 30 are rotated relative to one another until the contact state shown in FIG. 5 of the driving tooth flanks 20 e and 30 b of the bearing neck 20 a and pawl carrier 30 again sets, the brake 39 is completely released again.

If the bicycle is pushed backwards with the driver 13 at a standstill, that is to say the hub sleeve 16 rotates in the reverse direction of rotation RD, the following occurs:
The pawls of the output carrier pawl set 31 on the planet carrier side engage in the pawl toothing 32 of the hub sleeve 16 assigned to them, so that the planet gear carrier 20 a-bc by means of the pawl carrier 30 through mutual contact of the driving tooth flanks 20 e and 30 b without brake actuation in the reverse direction of rotation RD is rotated (without brake actuation because there is no relative rotation between the bearing neck 20 a and the pinch roller cage 30 a - see also FIG. 5). During this backward rotation of the planetary gear carrier 20 a-bc, the ring gear 21 is rotated at a higher rotational speed by means of the planetary gear sets 26-27-28 , of which the planetary gear set 28 rolls on the fixed sun gear 24 in the switching state according to FIG. 2 (7th gear) driven in the reverse direction of rotation RD as the planet carrier 20 a-bc. The ring gear 21 thus also rotates at a higher rotational speed in the reverse direction of rotation RD (corresponding to the planetary gear ratio) than the hub sleeve 16 , which drives the planet gear carrier 20 a-bc without gear ratio, so that the pawls of the ring gear-side output pawl set 34 on the corresponding one Pawl teeth 35 of the bearing insert ring 16 b rotating with the hub sleeve 16 slide past.

However, a locking problem occurs when pushing the bicycle backward, which is set forth below with reference to FIGS. 3 and 6, followed by a discussion of a solution to the problem. Due to their rotation in the reverse direction of rotation RD when pushing the bicycle back both the ring gear 21 (with the higher speed) via its pawl teeth 45 and the pawl set 44 as well as the planet carrier 20 a-bc (with the lower speed) via the front ring 20 c and the pawl set 37 to drive the driver 13 in the reverse direction RD. Because the translated larger rotational speed of the ring gear 21 of the Antreiberteil 13 takes n from the ring gear 21 in reverse rotational direction RD are entrained thereby perform a relative rotation to the planet gear carrier 20 a-bc, which is driven with the non-translated rotational speed of the hub shell 16 via the pawl 31 in a backward rotating direction RD is. This relative rotation would be permitted by the pawl set 37 on the planet carrier side, but is blocked by the brake activating pawl set 42 , which would lead to a blockage of the transmission and thus to the impossibility of pushing backwards. In order to avoid such a blockage, it is necessary to prevent at least one of the pawl sets 42 and 44 from being permanently engaged with the driver 13 in the event of a backward push. The following shows how the blocking can be prevented by suppressing permanent engagement between the ring gear 21 and the ring gear pawl set 44 on the driver part 13 n.

Referring to FIG. 6, it can be seen that upon rotation of the ring gear 21 in the reverse rotation direction RD (with the greater speed of rotation), the pawl teeth 45 of the ring gear 21 tries, the pawls of the ratchet set 44 n on Antreiberteil 13 take. In axial overlap with these pawls 44 , the features 71 of the switching ring 36 c are provided (see also FIGS. 2 and 12), the switching ring 36 c with its projections 70 being in play-free rotational driving connection with the pawl carrier 36 a- 36 b. The speed difference between the fast rotating ring gear 21 on the one hand and the pawl carrier 36 a- 36 b rotating slowly in unity with the switching ring 36 c on the other hand leads to a relative rotation between the unit 36 a- 36 b- 36 c driven by the planet gear carrier 20 a-bc ( consisting of the pawl carrier 36 a- 36 b and the switching ring 36 c with the features 71 ) and the driver part 13 n driven by the ring gear 21. This relative rotation is only possible within the scope of a rotational play a which has a flank spacing 13 f varying between 0 and a - 36 g between a tooth flank 13 f on the driver part 13 n and a tooth flank 36 g on the pawl carrier 36 a- 36 b (which is again to be understood as a driver part further from the chain sprocket DEG). Due to the engagement of the pawls of the pawl set 44 on the driver 13 and the pawl toothing 45 on the ring gear 21 rotating rapidly in the reverse direction of rotation RD, the driver part 13 n is carried along with the fast rotational speed of the ring gear 21 in the reverse direction of rotation RD, and the pawls of the pawl set 44 become underneath the manifestations 71 on the slower, namely with the planet carrier 20 a-bc, rotating in the reverse direction RD switching ring 36 c pushed. The driver part 13 n is carried along with the ring gear 21 by increasing the flank distance 13 f - 36 g. In the intermediate state shown in FIG. 6, the flank distance 13 f- 36 g corresponds to approximately half of the rotational play ä. Before the flank distance 13 f- 36 g reaches its maximum corresponding to the rotational play spiel, the pawls of the pawl set 44 are so far below that as blocking suppression member BS acting 71 on the slower rotating switching ring 36 c pushed that the mutual engagement between the pawls of the pawl set 44 and the pawl teeth 45 of the ring gear 21 is canceled. At this time, there is no more torque transmission from the ring gear 21 to the driver part 13 n. The ring gear 21 can thus rotate freely, ie without driving the driver part 13 n, past the pawls of the pawl set 44 . In this state, the driver part 13 n is temporarily without rotary drive, since also from the planet wheel carrier 20 a-bc ago via the pawl set 37 and the pawl carrier 36 a- 36 b, as long as no torque transmission to the driver part 13 n can take place until the edge spacing 13 f- 36 g has returned to zero, but in any case has reached its minimum. During the reduction of the flank spacing 13 f- 36 g, the pawl carrier 36 a- 36 b, which rotates with the planetary gear carrier 20 a-bc in the reverse direction of rotation RD through the mediation of the pawl set 37 , takes the switching ring 36 c over the projections 70 in the reverse direction of rotation RD , which ultimately leads to the pawls of the pawl set 44 coming out again under the features 71 and immersing again in the pawl toothing 45 on the ring gear 21 , specifically in another tooth gap, so that the above-described process (taking the pawls 44 and inserting them under the occurrences 71 etc.) repeated.

To avoid blocking the gearbox due to the different speeds of the ring gear 21 on the one hand and the planet carrier 20a-bc on the other hand, the pawls of the pawl set 44 between the driver part 13 n and the ring gear 21 are switched on and off intermittently, so that the reverse driving engagement of the ring gear 21 with the driver part 13 n via the pawls 44 is not constantly interrupted, but is intermittent; that the pawls dip one after the other in successive tooth gaps corresponds to the rotation of the ring gear 21 in the reverse direction RD relative to the driver part 13 n, which is now possible.

It should be added to the braking process with reference to FIG. 6 that the pawls of the pawl set 44 are brought out of engagement with the pawl toothing 45 between the driver 13 and the ring gear 21 in order to avoid blocking even when braking. When the driver part 13 n rotates in the reverse direction of rotation RD for the purpose of braking, the planet gear carrier 20 a-bc is carried along via the pawls of the brake-actuating pawl set 42 of the planetary gear 19 in the reverse direction of rotation RD; the planetary gear carrier 20 a-bc drives the ring gear 21 via one of the planetary gear sets 26 , 27 , 28 , specifically with an increased rotational speed in the reverse direction RD compared to the driver part 13 n. To prevent the transmission from jamming due to the relative rotation between the slower rotating driver part 13 n and the faster rotating ring gear 21 (both in the reverse direction of rotation), the pawls of the pawl set 44 also slide under the features 71 when braking, with the flanks spacing 13 f- 36 g increases. Only when the maximum edge spacing 13 f 36 g corresponding to the rotation game is ä reached, there is a torque transmission from the Antreiberteil 13 n to the gear rack 20 a-bc and ultimately to a brake applying rotational movement of the planet wheel carrier 20 a-bc in reverse rotational direction RD, which can now no longer be blocked because the pawls 44 have been switched off and can be continued until the brake 39 engages. Subsequent rotation of the driver 13 in the forward direction of rotation VD restores the initial conditions.

One can understand the flanks 13 f and 36 g and flanks 13 g and 36 h opposing in the circumferential direction as a rotary driving coupling MKS between two drive parts, namely between the drive sprocket-closer drive part 13 n and the sprocket-free drive part 36 a- 36 b, the is designed as a pawl carrier 36 a- 36 b for the pawls 37 and 42 .

Understanding the danger of locking when pushing the bicycle backwards and braking is facilitated by considering FIGS. 15a and 15b. Both figures represent a schematic development of the multi-speed shift hub according to FIGS . 1 to 13. The individual schematically illustrated hub components are designated with the same reference numerals as in FIGS. 1 to 13.

If the hub sleeve 16 in FIG. 15a is rotated at the reverse speed r in the reverse direction RD by pushing the bicycle backwards, then the hub sleeve 16 takes the planet carrier 20 a-bc in the reverse direction RD via the planet carrier-side drive clutch 31-32 , likewise at the speed r , The speed r is translated into high speed by the planetary gear 19 , so that the ring gear 21 rotates in the reverse direction RD at the speed 2r (the speed ratio 2 is arbitrary and only chosen for the sake of a simplified schematic representation). The ring gear 21 wants to take the driver 13 with it via the ring gear drive clutch 44-45 , ie drive it at the speed (2r) in the reverse direction RD. However, since the driver 13 is also to be taken along via the planetary gear drive clutch 37-38 , the PGS planetary gear system is blocked. (So far the problem 1)

A blockage also occurs if, according to FIG. 15b, the chain sprocket DEG and with it the driver 13 are rotated backwards at the braking speed b for the purpose of braking. The driver 13 takes over the brake activation clutch 42-38 the planet carrier 20 a-bc in the reverse direction of rotation RD, which should lead to a braking of the hub sleeve 16 by the braking device 39 . The hub sleeve 16 continues to rotate in the forward direction VD at a speed that decreases in accordance with the braking process. The reverse speed b given to the planetary gear carrier 20 a-bc by the driver 13 in the reverse direction RD is transmitted via the planetary gear 19 to the ring gear 21 , namely with gear ratio, so that the ring gear 21 rotates at the reverse speed 2b (for the gear ratio 2 see above) , The ring gear driven driven clutch 34 , 35 is overhauled. However, the reverse speed 2b of the ring gear 21 is transmitted to the driver 13 through the ring gear drive clutch 44 , 45 . However, since the reverse speed b is also introduced into the driver 13 by the sprocket DEG, there is a blockage. (Problem 2)

To suppress the blocking when pushing backwards and braking, a reversing sensor system MKS is provided according to FIG. 16a between the driver part 13 n near the chain sprocket and the driver part 36 a- 36 b remote from the chain sprocket, in conjunction with a blocking suppressor BS. The reverse rotation sensor system MKS is designed as a rotary drive coupling MKS with rotational play. In Fig. 16a you can see the rotational play Ä. Ä corresponds to the maximum value of the angular distance 13 f- 36 g between the tooth flanks 13 f and 36 d of the driver parts 13 n and 36 a- 36 b connected to one another by the rotary driving coupling MKS. In Fig. 16a the large overdrive is shown in the forward operation according to Fig. 1 and 2. The angular spacing 13 f 36 g is set to zero. The chain sprocket further Antreiberteil n assumes 13 via the tooth flanks 13 f and 36 g the chain sprocket distant Antreiberteil 36 a- 36 b, the pawl carrier 36 di a- 36 b, with. The drive gear 36 a- 36 b, which is remote from the sprocket, drives the planet gear carrier 20 a-bc in the forward direction VD at the forward speed v via the planet carrier drive coupling 37-38 . At the same time, the ring gear 21 is driven via the planetary gear 19 with the translated forward speed 2v and drives the hub sleeve 16 at the forward speed 2v via the ring gear-side output clutch 34-35 . The output clutch 31-32 and the ring gear drive clutch 44-45 are overhauled. The forward drive in the forward direction of rotation VD is therefore possible without any problems.

In Fig. 16b 1 and 2 in turn is of the quick response corresponding to FIGS., Specifically generated in the reverse slide in a state I. The reverse slide on the hub sleeve 16 r a reverse speed. This reverse speed r is transmitted via the planet carrier-side output clutch 31-32 to the planet carrier 20 a-bc, which consequently rotates at the reverse speed r in the reverse direction of rotation and via the planet carrier drive clutch 37-38 the pawl carrier 36 a- 36 b, that is, the driving sprocket further from the driver part of the driver 13 . From the planet carrier 20 a-bc, the ring gear 21 is driven with gear ratio, namely with the reverse speed 2r. This reverse speed 2r of the ring gear 21 is transmitted via the ring gear drive clutch 44-45 to the part 13 n of the driver 13 closer to the sprocket. The tooth flanks 13 f and 36 g move away from each other because the driver part 13 n is driven at a higher speed 2r in the reverse direction of rotation than the pawl carrier 36 a- 36 b; the angular distance 13 f- 36 g increases until the pawl 44 is actuated. This state is shown in Fig. 16c (state II of pushing backwards). In the transition from state I of the backward sliding of Fig. 16b in the state II shown in FIG. 16c is by the difference in speed 2 r n minus r between the chain sprocket close Antreiberteil 13 (which initially is driven by the Hohlradantriebskupplung 44-45) and the chain sprocket distant Antreiberteil 36 a- b (this clutch driven by the planet carrier drive 37-38) the locking pawl into the Hohlradantriebskupplung moved 44 44-45 under the blocking suppression element BS, so that the ring gear driving clutch 44-45 of FIG. 16c is turned off 36th This means that, according to FIG. 16c, the drive part 13 n, which is closer to the sprocket, is stationary. The pawl carrier 36 a- 36 b continues to run with the planet carrier 20 a-bc backwards at the reverse speed r. This is possible without the driver part 13 n closer to the sprocket being taken along in the reverse direction of rotation RD because the angular distance 13 f- 36 g has increased according to FIG. 16c. If, with the driver part 13 n at a standstill, the pawl carrier 36 a- 36 b changes from the state of FIG. 16c to the state III according to FIG. 16d, the blocking suppressor BS lifts off the pawl 44 of the ring gear drive clutch 44-45 , so that the pawl 44 engages again in the pawl toothing 45 , but in a different tooth gap, as the comparison of FIGS. 16b and 16d shows.

The following can thus be seen: When pushing the bicycle backwards, the ring gear drive clutch 44-45 is alternately switched from switching state I according to FIG. 16b to switching state II according to FIG. 16c and then from switching state II according to FIG. 16c to switching state III according to FIG. 16d ( the latter according to the switching state according to FIG. 16b), the pawl 44 jumping over one tooth of the pawl toothing 45 . This also means that the chain sprocket near Antreiberteil 13 n alternately with the rotation speed 2 r in the reverse rotational direction RD running back (Fig. 16b) then stops (Fig. 16c), and then again with the speed of 2r in reverse rotational direction RD running (Fig. 16d ).

The blockage found when viewing FIG. 15a when pushing backwards is thus suppressed.

The blockage during braking observed when viewing FIG. 15b is also suppressed by the construction according to FIG. 16a. This results from FIGS. 17a and 17b, which represent two successive states, brakes IV and brakes V. In the brake IV state according to FIG. 17a, the drive part 13 n closer to the sprocket is rotated backwards in the reverse direction RD. It is assumed that, prior to the initiation of braking, the fast forward gear according to FIG. 16a took place, ie the angular distance 13 f- 36 g is set to zero, ie the two tooth flanks 13 f and 36 g abut one another. This means that the drive sprocket part 13 n closer to the sprocket in accordance with FIG. 17 a can be rotated in the reverse direction of rotation RD without taking the pawl carrier 36 a- 36 b with it and the braking process can thus be initiated until the angular distance 13 f- 36 g is maximal and Angular distance 13 g - 36 h O is. When the angular distance 13 g - 36 h is used up corresponding to the state of braking V in FIG. 17 b, the pawl carrier 36 a - 36 b is taken along by the chain tooth drive part 13 n closer to the chain tooth via the tooth flanks 13 g and 36 h, so that the planet gear carrier 20 a-bc on the brake activation clutch 42-38 in the reverse direction RD with the brake speed b is taken and thereby the coaster brake 39 is activated to slow the speed to be braked u of the hub sleeve 16 . This movement of the planet carrier 20 a-bc in the reverse direction of rotation RD is transmitted via the planetary gear 19 to the ring gear 21 with gear ratio, so that the ring gear 21 rotates with the translated brake speed 2 b in the reverse direction of rotation RD. In the reverse movement of the ring gear 21 thus occurring, however, as shown in FIG. 17b, the ring gear drive clutch 44-45 is switched off, since the pawl 44 under the blocking suppression element already during the transition from state IV according to FIG. 17a to state V according to FIG. 17b BS has come ( Fig. 17b). The braking movement b in the reverse direction RD can thus continue until the coaster brake 39 exerts a braking effect on the hub sleeve 16 . The state shown in Fig. 17b is retained, even if the braking movement b in the reverse direction of rotation RD has come to a standstill by gripping the coaster brake 39 , as long as is stepped forward again.

So much for the principle and behavior of the 7-speed gear hub shown in FIG. 2 in 7th gear, which corresponds to the largest gear ratio and is therefore called high overdrive.

If it is desired to switch to 6th gear, a medium-sized gear ratio or medium overdrive gear, the medium-sized sun gear 23 is blocked for rotation about the hub axis 10 , and the small sun gear 22 and the large sun gear 24 are prevented from rotating about the hub axis 10 released . How this switchover of the sun gear conditions occurs will be discussed in more detail later in the description. The shifting state of the 6th gear differs from the shifting state of the 7th gear only in that the gear ratio of the planetary gear 19 is smaller: the hub sleeve 16 is therefore still driven at an increased speed compared to the speed of the driver 13 , but the ratio the speed of the hub sleeve 16 to the speed of the driver 13 has become lower than in 7th gear or high overdrive. Incidentally, the movements are when driving (rotation of the driver 13 in the forward direction of rotation), when idling (standstill of the driver 13 and rotation of the hub sleeve 16 in the forward direction of rotation), when braking (rotation of the driver 13 in the reverse direction of rotation and rotation of the hub sleeve 16 in the forward direction of rotation) and when pushing the bicycle backwards (standstill of the driver 13 and rotation of the hub sleeve 16 in the reverse direction of rotation) the same as described above for the 7th gear be.

If it is desired to switch to 5th gear, ie the small overdrive, the small sun gear 22 is blocked against rotation about the hub axis 10 , and the two sun gears 23 and 24 can rotate freely about the hub axis 10 . The processes when switching the sun gears in the sense of a blocking of the small sun gear 22 and a rotation release of the sun gears 23 and 24 will be described later. If the small sun gear 22 is blocked, the ring gear 21 still rotates at an increased speed compared to the speed of the driver 13 and the planet gear carrier 20 a-bc, which via the ring gear-side output pawl set 34 and the pawl toothing 35 on the hub sleeve 16 is transmitted. The ratio of the speed of the ring gear 21 and the hub sleeve 16 to the speed of the driver 13 has become even smaller than in 6th gear. Incidentally, the movement conditions are also in 5th gear when pedaling forward (drive of the driver 13 in the forward direction of rotation), in freewheeling (standstill of the driver 13 and rotation of the hub sleeve 16 in the forward direction of rotation), when braking (rotation of the driver 13 in the reverse direction of rotation and rotation of the hub sleeve 16 in the forward direction of rotation) and when pushing the bicycle backwards (standstill of the driver 13 and rotation of the hub sleeve 16 in the reverse direction of rotation) the same as described above with respect to the 7th gear.

In the 4th Gear, the hub sleeve 16 is driven at the same speed as the driver 13 . One speaks of the direct passage. This is where the previously mentioned hollow wheel pawl set 44 and the pawl toothing 45 belonging to it on ring gear 21 come into play. The direct gear corresponding to the 4th gear is present when the ring gear 21 is driven directly by the driver 13 via the ring gear pawl set 44 and the ring gear 21 drives the hub sleeve 16 via the ring gear driven output pawl set 34 . The hub sleeve 16 then rotates at the same speed as the driver 13 . However, in order to transfer the speed of the driver 13 directly to the ring gear 21 via the ring gear pawl set 44 , the torque transmission from the driver 13 to the planet carrier 20 a-bc must be switched off, which previously existed thanks to the planetary pawl set 37 . Specifically, as long as the planet carrier 20 a-bc is driven by the pawl set 37 driving the planet carrier, the ring gear 21 rotates at a higher speed than the driver 13 due to the action of the planetary gear 19 , the pawl toothing 45 sliding over the pawl set 44 driving the hollow wheel so that the ring gear 21 cannot rotate at the speed of the driver 13 . For the purpose of bringing about the 4th gear, ie the direct gear, the torque transmission from the driver 13 to the planet carrier 20 a-bc must be switched off via the pawl set 37 driving the planet carrier. This is done in the following way:
A thrust block 48 is arranged to be displaceable in the axial direction within a diametral window 49 of the hub axis 10 . When a control signal is introduced via a signal line SL, the thrust block 48 is shifted into an axial central position within the diametral window 49 . The pushing block 48 is displaced with the support of a compression spring 72 . During the displacement, coupling sleeves 73 and 74 connected to the thrust block 48 are also displaced. The coupling sleeve 73 is connected to the thrust block 48 and consequently does not rotate; The coupling sleeve 74 , which is shown in perspective in a single part in FIG. 11, is coupled via connecting tongues 75 to the thrust block 48 and the coupling sleeve 73 in such a way that it can be displaced in the axial direction with the thrust block 48 , but with respect to the latter and the coupling sleeve 73 rotates. The connector tabs 75 extend through diametrically namely window 76 of the front ring 20 c of the planet carrier 20 a-bc (see also Fig. 10). The interaction of the connecting tongues 75 and the diametral window 76 results in a rotary drive from the planet gear carrier 20 a-bc to the coupling sleeve 74 . As can also be seen from FIG. 11, axial projections 77 are provided on the coupling sleeve 74 . Axial projections 78 are also attached to the coupling sleeve 73 . If one now pushes the thrust block 48 into its central position within the diametral window 49 , the non-rotating coupling sleeve 73 with its axial projections 78 is pushed into the area of the pawls of the pawl set 37 . This displacement of the thrust block 48 and the coupling sleeve 73 is only possible if the radially inward-pointing control tails of the pawls of the pawl set 37 (see FIG. 3) do not stand in the way of the axial projections 78 which follow one another in the circumferential direction. Otherwise, the thrust block 48 together with the coupling sleeve 73 remains in the starting position shown in FIG. 2 until, due to the relative rotation of the pawl carrier 36 a- 36 b to the non-rotating coupling sleeve 73 , a gap between two successive axial projections 78 with a control tail of a pawl 37 flees; Only then can the coupling sleeve 73 and the thrust block 48 be displaced by means of the compression spring 72 . With continued relative rotation of the pawl carrier 36 a- 36 b to the non-rotating coupling sleeve 73 , the control tails of the pawls 37 then slide onto the axial projections 78 and are lifted by them, so that the radially outer engagement ends of the pawls 37 (see FIG. 3) are disengaged are "lifted" with the pawl toothing 38 on the planetary gear carrier 20 a-bc and remain raised on a closed annular surface of the coupling sleeve 73 due to a bearing of the control tails occurring in the course of further axial displacement. The force required to disengage the pawls of the pawl set 37 is thus derived solely from the torque of the rotary pawl carrier 36 a- 36 b. The force to be applied by the compression spring 72 or by a manual switching operation is only required for the axial adjustment of the thrust block 48 and the coupling sleeve 73 to initiate the switching operation, but not for lifting the pawls 37 out of the corresponding pawl toothing 38 . This allows shifting from 5th to 4th gear under load without requiring a particularly powerful 31836 00070 552 001000280000000200012000285913172500040 0002010030941 00004 31717 spring 72 or a large amount of force via the signal line SL.

If freewheeling operation is to take place in the 4th or direct gear, ie when the driver 13 is stationary and the hub sleeve 16 rotates in the forward direction of rotation, the pawl toothing 35 slides over the output pawl set 34 on the ring gear side.

If braking is to take place in 4th gear, the driver 13 is rotated in the reverse direction of rotation, while the hub sleeve 16 continues to rotate in the forward direction of rotation. The planet gear carrier 20 a-bc is carried along by the driver 13 in the reverse direction of rotation via the pawl set 42 , so that, analogously to the braking process in 7th gear, the brake jacket 40 is spread again and the hub sleeve 16 is thereby braked.

If the bicycle is pushed backwards in 4th gear, ie the hub sleeve 16 rotates in the reverse direction of rotation, then the same conditions arise as were explained in detail above for 7th gear with reference to FIG. 6.

In the 3rd Gear is reduced by means of the planetary gear 19, the speed of the driver 13 in slow motion, ie the hub sleeve 16 rotates in the forward direction of rotation at a speed which is less than the speed of the driver 13 , which is introduced into it by the drive chain of the bicycle. Referring again to the illustration of FIG. 2, the sun wheel is again blocked so 22 as in the 5th gear against rotation about the hub axle 10, while the sun gears 23 and 24 are freely rotatable on the hub axle 10. The driver 13 drives the ring gear 21 in the 3rd gear via the ring gear driving pawl set 44 and the pawl toothing 45 in the forward direction of rotation; the planetary pawl set 37 is switched off again, as described in connection with the 4th gear or direct gear. In addition, the output pawl set 34 on the ring gear side is also switched off in 4th gear, so that the hub sleeve is no longer carried along by the ring gear 21 . This "disengagement" of the output pawl set 34 on the ring gear side is effected by a further displacement of the thrust block 48 by means of the compression spring 72 into its rightmost position. The clutch sleeve 73 is again shifted further to the right, the pawl set 37 driving the planet carrier remaining unchanged. With this axial displacement of the thrust block 48, the coupling sleeve 74 is also moved further to the right. The axial projections 77 first come into the area of the pawl set 34 as soon as the relative angular position permits this and then, thanks to the relative rotation between the ring gear 21 and the planet gear carrier 20 a-bc, the pawls 34 out of engagement with the ratchet toothing 35 on the hub sleeve 16 . This means that this changeover from 4th to 3rd gear under load can also take place without the need for a particularly strong spring 72 or a large amount of force being applied via the signal line SL.

In the 3rd Gear is driven by the ring gear 21 via the planetary gear 19 of the planetary gear carrier 20 a-bc at a speed which is reduced in relation to the ring gear 21 and the driver 13 ; it rotates in the forward direction of rotation at a lower speed than the driver 13 . Since the smallest sun gear 22 is blocked on the hub axle 10 and the largest planet gear 26 meshes with the sun gear 22 , the lowest reduction ratio is selected. Accordingly, the 3rd gear is also called the low-geared slow gear or the low-geared mountain gear. The step-down rotary movement of the planetary gear carrier 20 a-bc is transmitted to the hub sleeve 16 via the output pawl set 31 and the ratchet toothing 32 on the planet carrier side, so that the hub sleeve 16 rotates at a speed which is slightly reduced compared to the speed of the driver 13 .

When freewheeling takes place in 3rd gear, ie when the driver 13 and with it the planetary gear 19 is stationary, the pawl toothing 32 slides over the pawl set 31 on the planet carrier side.

If braking is to take place in 3rd gear, the driver 13 is rotated in the reverse direction of rotation, and the hub sleeve 16 continues to rotate in the forward direction of rotation with decreasing speed. By rotating the driver 13 in the reverse direction of rotation, the planetary gear carrier 20a-bc is rotated in the reverse direction by means of the brake-actuating pawl set 42 , namely at the same speed as the driver 13 . The pinch rollers 43 are effective so that the reverse rotation of the bearing neck 20 a of the planet carrier 20a-bc and the pinch rollers 43 spreads the brake jacket 40 and presses it frictionally against the braking surface 41 . The torque curve when braking through pedaling backwards is unchanged as in 7th to 4th gear.

When the bicycle is pushed backwards in 3rd gear, the driver 13 stands still and the hub sleeve 16 rotates in the reverse direction of rotation. The same conditions occur as have already been explained for 7th gear with reference to FIG. 6.

In the 2nd The medium-sized sun gear 23 is blocked against rotation about the hub axis 10 and the sun gears 22 and 24 are freely rotatable. The planetary gear 19 then has a medium reduction effect. The processes in forward pedaling are the same as in the case of the third gear. The same processes occur in freewheeling, braking and pushing the bicycle backwards as in 3rd gear.

In the 1st Gear, the strongly reduced slow gear or mountain gear, the sun gear 24 is blocked against rotation about the hub axis 10 , and the sun gears 22 , 23 are freely rotatable. The planet gear carrier 20a-bc is driven by the driver 13 via the ring gear-side drive pawl set 44 , the pawl toothing 45 , the ring gear 21 and the planetary gear 19 at a strongly reduced speed in the forward direction of rotation and takes the hub sleeve 16 via the planet carrier-side output pawl set 31 and the pawl toothing 32 in the forward direction with. In free-wheeling mode, in braking mode and when pushing the bike backwards, the same movement conditions result as described in connection with 3rd gear.

It should be pointed out that the torque when braking in all gears is transmitted from the driver 13 via the planetary gear carrier 20 a-bc to the coaster brake 39 with the same brake transmission ratio (corresponding to the direct gear).

The optional fixing of the sun gears 22 , 23 and 24 against rotation about the hub axis 10 is carried out by means of a locking sleeve 47 acting as a sun gear condition, which can be axially displaced in the axial direction A of the hub axle 10 and by interacting toothings of the hub axle 10 and the locking sleeve 47 on rotating is hindered about the hub axis 10 . In addition to the diametral window 49 of the hub axle 10 , the thrust block 48 also passes through diametral slots 50 of the blocking sleeve 47 . The blocking sleeve 47 , which is blocked against rotation about the hub axis 10 , has on its left end an external toothing 51 , which is in blocking engagement with an internal toothing 54 of the sun gear 24 in 7th gear according to FIG. 2; this external toothing 51 can be brought by axial displacement of the blocking sleeve 47 also for blocking engagement with an internal toothing 53 of the sun gear 23 and with an internal toothing 52 of the sun gear 22 , so that either one of the sun gears 22 , 23 , 24 is blocked. The mechanism (part of the gear change device GW) for axially displacing the blocking sleeve 47 is shown enlarged in FIG. 2 in area B of FIG. 7.

The hub axle 10 has three diametrical openings b1, b2 and b3 in the region B. The diametrical openings b1 to b3 intersect with an axial bore 54-1 of the hub axle 10 , which is open at least towards the right end of the hub axle 10 . Two balls k1, k2 and k3 are accommodated in the diamond openings b1, b2 and b3. The balls k1, k2, k3 rest with their radially inner poles on a coil-shaped control body 55 which is axially displaceable within the axial bore 54-1 . For the axial displacement of the control body 55 , a displacement mechanism (not shown) is provided to the right of the latter on the hub axle 10 , which displacement mechanism is part of the signal line SL or is coupled to it. The signal line SL comes from a gear switch GS ( FIG. 2), which is arranged, for example, on the handlebar of the respective bicycle. By actuating this gear switch GS, the control body 55 can be moved in the axial direction A. The balls k2 and k3 face with their radially outer poles a cam carrier unit 56 , while the ball k1 faces the inner circumferential surface of the blocking sleeve 47 functioning as a sun gear condition converter. The control cam carrier unit 56 is guided axially displaceably and non-rotatably in a sliding sleeve 62 . The sliding sleeve 62 is in turn connected to the driver 13 via a friction sleeve 63 . The cam carrier unit 56 is thus taken through the friction sleeve 63 by the driver 13 in the direction of rotation, the friction sleeve 63 acting as an overload clutch. The cam carrier unit 56, as a part rotating with the driver 13, is connected to the axially displaceable but not rotatable blocking sleeve 47 by means of a rotary connection 57 for common axial movement. The cam carrier unit 56 is composed of two parts 56 a and 56 b, which collide in the area of a cam groove 58 . In the position of the control body 55 according to FIG. 7, the ball k3 is immersed in the control groove 58 of the control venträger 56 under the action of the thickening 61 of the control body 55 which is trapezoidal in longitudinal section. The cam groove 58 is, as shown in the development of the radially inner side of the cam carrier 56 according to FIG. 8a, delimited by two cams 59 and 60 which are directed towards one another, with each of the parts 56 a and 56 b in each case one of the cams 59 and 60 is designed as a transition shoulder between inner circumferential surfaces 56-1 and 56-2 offset in relief. The control curves 59 , 60 are intended to interact with the balls k1, k2 and k3. As can be seen from FIG. 8a, the control cams 59 and 60 have an axial minimum distance a at each point along the circumference such that two adjacent balls can be exchanged into the control cave groove 58 at the same time without blocking the control cage support unit 56 from rotating , . In the representation according to Fig 7 the ball is k3 with its radially inner pole in contact with the trapezoidal thickening 61 of the control body 55 so that the ball k3 from the associated Diametralboh tion is immersed in the cam groove 58 63, that is - when viewed in the axial direction - with its radially outer pole area in radial overlap with the control cam 60 . As can be seen from FIG. 8a, the control curves 59 and 60 run on part of their circumference with a slope á with respect to a normal surface N to the hub axis A. It can also be seen from FIG. 7 that half the diameter of the balls k1, k2 and k3 is greater than the depth of the cam groove 58 . In addition, it can be seen in FIG. 7 that, viewed in the axis-parallel section, the profiles 59 a, 60 a of the control curve 59 , 60 run against the axis-normal surface N such that the ball k3 (the same also applies to the balls k1 and k2 under certain conditions) can climb up over these profiles. The profiles 59 a and 60 a are chamfered in sections, so that when one of these profiles interacts with one of the balls k1, k2 or k3, the respective ball at t1, t2, t3 on the wall of the holes b1, b2 or b b3 is pressed and damage to the control body 55 can thus be excluded. This minimizes radial forces at n1, n2, n3.

. With simultaneous consideration of Figures 7 and 8a it can be seen the following:
The ball k3 is located in a continuous circumferentially middle zone 58 a of the cam groove 58, which extends substantially parallel to the axis-normal surface N and which is of two dashed lines drawn Hüllkreisen (in the settlement Hüllgeraden) I1 and I2 limited. These enveloping circles I1 and I2 have a distance b. The distance b is slightly larger than the diameter d of the exchanged in the cam groove 58 spherical segment of the balls k1, k2, k3 (here ball k3), so that in the instantaneous state of FIG. 7 and 8a with a rotational movement of the cam carrier 56 in the forward rotational direction VD relative to the ball k3 do not set any constraining forces between the curve 60 and the ball k3 and the cam carrier 56 thus maintains its axial position according to FIGS. 7 and 8a with the result that the external toothing 51 of the blocking sleeve 47 is in rotationally blocking engagement with the Sun gear 24 remains. Even if, in the course of further rotation of the control cam unit 56 in the direction of the arrow VD according to FIG. 8a, the peripheral section 60 n is axially opposite the ball k3, there is no constraint between the ball k3 and the control cam 60 . Furthermore, in the state according to FIGS. 7 and 8a, when the control cam carrier 56 rotates, the control cam 59 does not come into compelling contact with the ball k3, so that no axial displacement of the control cam carrier unit 56 with respect to the ball k3 occurs due to the interaction of the ball k3 and the control cam 59 takes place, which in turn has the consequence that the locking sleeve 47 maintains its position shown in Fig. 7 axial position, and thus the sun gear 24 remains blocked. The balls K1 and K2 are in the state of FIG. 7 and 8a narrowed to the diameter portion 64 a of the control body 55 at. The ball k1 is constantly held within its bore b1 by the inner peripheral surface of the blocking sleeve 47 . The ball k2 could then move radially outwards if it is located in the peripheral portion 59 n of the control curve 59 , but will then, when it comes into engagement with the peripheral portion 59 m of the control curve 59 , due to its large diameter and the profile 59 a Control curve 59 causes climbing over the profiling 59 a to move radially inward again until it abuts on the diameter-narrowed section 64 a of the control body 55 . In conclusion, it can be determined again that in the position of the control body 55 according to FIG. 7, none of the balls k1 to k3 engages with the control cams 59 and 60 in such a way that an axial displacement of the control cam carrier unit 56 and thus the blocking sleeve 47 could occur. The sun gear 24 remains locked.

If the control body 55 is shifted to the left by the schematically drawn signal line SL in FIG. 7, the trapezoidal thickening 61 tries to push the ball k2 radially outward. A displacement of the ball k2 radially outward within the diametral bore b2 can only occur if, in the course of the rotation of the control cam carrier unit 56 in the forward direction of rotation VD, the state according to FIG. 8a is reached, ie if the peripheral section 59 n of the control cam 59 in located axially opposite the ball k2. The ball k2 can then enter the cam groove 58 unchecked and thus come into radial overlap with the cam 59 . If then during further rotation of the control curve carrier unit 56 in the forward rotational direction VD of Fig. 8a, the portion 59 of the cam 59 m, the ball approaches k2, so must the control curve carrier unit 56 to evade to the left, ie is shifted axially to the left and displaces about the pivotal connection 57 the blocking sleeve 47 also to the left, with the result that the external toothing 51 comes out of the blocking engagement with the internal toothing 54 of the sun gear 24 and engages with the internal toothing 53 of the sun gear 23 . As a result of the axial displacement of the cam carrier unit 56 to the left, the peripheral portion approaches 60 60 n of the cam upon rotation of the cam carrier unit 56 in the forward rotational direction of the ball to k3. This means that the peripheral portion 60 m of the cam 60 comes as a consequence of the rotary motion of the cam carrier unit 56 in the forward rotational direction VD and this rotational movement superimposed axial displacement of the cam carrier unit 56 to the left in circumferential alignment with the ball k3 and that when the peripheral portion 60 m after a further rotation of the cam carrier unit 56 strikes the ball k3, this by the control cam 60 is again radially inward narrowed in the direction of the diameter portion 64 b of the control body 55 is shifted so that, after further rotation of the in Figure 8b. state shown adjusts and henceforth with further rotation of the cam carrier unit 56, this maintains its axial position, which it has achieved through the interaction of the radially extended ball k2 and the cam 59 . The engagement between the external teeth 51 of the locking sleeve 47 and the sun gear 23 is then made, as the ball passes in this state, in the "neutral" zone 58 a k2.

If the sun gear 22 is to be blocked against rotation about the hub axis 10 , the control body 55 is shifted further to the left, so that the ball k1 can now be pressed radially outward. This is possible because the blocking sleeve 47 has stepped to the left until the external toothing 51 engages in the internal toothing 53 of the sun gear 23 . The ball k1 can then move radially outwards when the circumferential section 59 n of the control cam 59 has reached the ball k1 in axial opposition. The processes that were described when the control body 55 was first displaced for the purpose of pushing the ball k2 radially outward are then repeated. Finally, the state according to FIG. 8c is established.

The switching back of the blocking sleeve 47 from the position according to FIG. 8c to the position according to FIG. 8b can be initiated by shifting the control body 55 to the right, so that the ball k2 is again immersed in the control groove 58 and the ball k1 is removed from the control groove again returned to its hole b1. The same applies to switching back from the position shown in FIG. 8b to the position shown in FIG. 8a.

The pole spacing of the balls k1, k2 and k3 must correspond to the center-to-center spacing of the toothings 52 , 53 , 54 . The axial plateau length x of the circumferential surface 61 x of the trapezoidal thickening 61 must, on the one hand, be long, so that the unintentional switching operation is not initiated in the event of an unintentional axial movement of the control body 55 . On the other hand, the axial plateau length x must be dimensioned such that when the trapezoidal thickening 61 begins to contact the ball k1, the ball k2 with its radially inner pole still rests on the plateau 61 x, so that the displacement of the control body 55 occurs before the radial outward displacement of the ball k1 is initiated is ensured in a position in which the displacement of the cam unit 56 by the ball k2 has come to an end.

It should be noted that the pushing of the blocking sleeve 47 back and forth is not brought about by a large force signal on the signal line SL, but by the relative rotation of the cam carrier unit 56 with respect to the hub axis 10 , which as a result of the rotational driving of the cam carrier unit 56 by means of the friction sleeve 63 occurs. The force for displacing the control body 55 only has to be so great that the balls k1, k2, k3 are displaced radially by means of the trapezoidal thickening 61 . It is therefore possible to shift under load even with those gear changes that occur due to the blocking of different sun gears, without the need for large forces to be introduced from outside via the signal line SL. After it has already been stated earlier that no large forces have to be introduced from the outside via the signal line SL for the gear change processes which are brought about by switching off the pawl sets 34 and 37 , it can thus be stated that all the switching processes can be carried out under load ,

The retraction of the control body 55 can be effected by the signal line SL. However, it is also possible for the control body 55 to be switched back by means of a helical compression spring 83 , which is accommodated and prestressed as a compression spring in the left region of FIG. 2 within the hub sleeve 10 and acts on the control body 55 .

The signal line SL consists of two strands, one of which acts on the control body 55 and the other of which acts on the thrust block 48 . As can be seen from the above description, the two strands are acted on at different times. Nevertheless, it is possible to activate the two strands at different times if desired from a single gear switch GS, for example by twisting two different cam disks NS1, NS2 by means of a single shift wire SD or a shift linkage from the gear switch GS on the handlebar , The two cam discs NS1, NS2 are profiled differently. One cam disc acts on the control body 55 , the other cam disc on the thrust block 48 . By under different profiles of the two cams, the mutual coordination of the displacements of thrust block 48 and control body 55 is achieved.

The signal line SL can be designed with a resilient transmission element (not shown), so that the control body 55 can temporarily stand still when shifting to the left, if a ball k2 or k1 to be displaced radially outwards cannot be displaced radially outwards yet because the peripheral section 59 n has not yet come into axial comparison with the respective ball k2 or k1.

Fig. 13 is a development showing a cam carrier unit 56 'with a modified control cam shape. It is noteworthy that only the forward direction of rotation VD pointing cam regions 59 'v 60' v of the cams 59 ', 60' are compared with the normal N inclined surface with an angle á. The trailing flank areas 59 'n, 60 ' n can be inclined at a right angle â. With such an asymmetrical geometry of the control curves 59 ', 60 ', more control curve periods P can be accommodated on the circumference of the control body unit 56 ', which leads to a faster response of the circuit to a switching signal entered by the cyclist: the more periods P are arranged on the circumference , the faster the control carrier unit 56 is switched after the control body 55 has been moved . It should be pointed out, however, that even with a control cam design according to FIG. 13, there is an imperative requirement that two balls can be immersed in the control cam groove 58 ′ at the same time without the control cam carrier 56 ′ being locked against rotation. This condition limits the number of within the control curve carrier unit 56 'maximum storable cam periods P. Incidentally, the operation of the control curve carrier unit 56' corresponds to the to the control curve carrier unit 56 has been said.

A second embodiment of the 7-speed shift hub is explained below with reference to FIG. 14. To avoid repetition, only the differences from the 7-speed shift hub explained with reference to FIGS. 1-13 are explained. The same reference numerals as for the first embodiment are used for components of the same type or function, but the number 100 is preceded by these reference numerals.

The mechanism for fixing the sun gears 122 , 123 , 124 with the aid of the blocking sleeve 147 is arranged to the left of the planetary gear sets 126 , 127 , 128 . The control cam carrier unit 156 has a toothing 180 radially on the outside, via which it is connected to a sliding sleeve 162 ; the sliding sleeve 162 is driven by means of a friction sleeve 163 of the bearing neck 120 a of the planet carrier 120 a-bc.

The coupling sleeve 173 can be arranged directly on the hub axle 110 , since the mechanism for fixing the sun gears is moved to the left side of the planetary gear sets 126-127-128 . With such a structure, the multi-speed hub can be made more compact, in particular because the functional groups requiring space for fixing the sun gears on the one hand and for controlling the pawl sets 134 and 137 on the other hand are arranged on different sides of the planetary gear sets 126-127-128 are.

The rotary drive of the cam carrier unit 156 is no longer derived directly from the driver, but from the planet gear carrier 120 a-bc via the friction sleeve 163 and the sliding sleeve 162 . To "switch" from one sun gear to the other by moving the blocking sleeve 147 , a rotary drive of the cam carrier unit 156 by the planet gear carrier 120 a-bc is required. To ensure a reliable switching process between two sun gears, it is therefore necessary to ensure a permanent rotary movement of the planet gear carrier 120 a-bc. The following problem arises: Is the blocking sleeve 147 with its external teeth 151 just in a non-engagement intermediate position between the internal teeth of two adjacent sun gears, z. B. the sun gears 123 and 124 , no rotational movement is transmitted to the planet carrier 120 a-bc. To remedy this, a friction ring 129 is inserted between the ring gear 121 and the planetary gear carrier 120 a-bc, which, even if all sun gears 122 , 123 , 124 are not blocked, for driving the planetary gear carrier 120 a-bc (and thus the cam carrier 156 ) with it Ring gear 121 ensures and thus ensures a complete displacement of the blocking sleeve 147 .

Claims (33)

1. Multi-speed shift hub for bicycles or the like, comprising:
a hub axis ( 10 ) with a geometric axis (A),
one with at least one torque input member (DEG), in particular a chain sprocket (DEG), coupleable or coupled driver (13) which is rotatably supported on the hub axle (10),
a hub sleeve ( 16 ) rotatably mounted on the hub axle ( 10 ),
a within the hub sleeve ( 16 ) planetary gear system (PGS) with at least one planetary gear ( 19 ) for torque transmission from the driver ( 13 ) to the hub sleeve ( 16 ), this planetary gear system (PGS) executed with
at least one sun gear ( 22 , 23 , 24 ), which can optionally be switched between a state of free rotation about the hub axis ( 10 ) and a state of rotation blocking relative to the hub axis ( 10 ),
at least one ring gear ( 21 ) rotatably mounted on the hub axle ( 10 ) and
at least one planet gear ( 26 , 27 , 28 ) which is rotatably mounted on at least one planet gear carrier ( 20 a-bc) which can be rotated about the hub axis ( 10 ),
also comprehensive
an axially movable sun gear state converter ( 47 ) which, depending on its axial position, causes the sun gear ( 22 , 23 , 24 ) to either assume the state of free rotation or the state of the rotation lock,
a gear change device (GW) for moving the sun gear state converter ( 47 ) in the direction of the geometric axis (A), this gear change device (GW) in operative connection
with at least one gear change signal line (SL) inserted into the multi-gear hub from at least one gear change signal generator (GS),
marked by

• a) a control cam carrier unit ( 56 ) rotatable about the geometric axis (A) with at least one control cam ( 59 , 60 ) running in the circumferential direction about the geometric axis (A) and inclined to an axis-normal plane (N);
• b) a rotary driving connection ( 62 , 63 ) between a hub component ( 13 ) rotating about the geometric axis (A) when the driver ( 13 ) and / or and the hub sleeve ( 16 ) rotates, and the cam carrier unit ( 56 );
• c) at least one control cam contact element (k1, k2, k3) in axial comparison with the control cam ( 59 , 60 ),
• d) wherein this cam system element (k1, k2, k3) is essentially immovable with respect to the hub axis ( 10 ) in the direction of the geometric axis (A) and in the circumferential direction and is adjustable in the radial direction with respect to the geometric axis (A) between a radially overlapping position (at k3) to the control curve ( 59 , 60 ) and a radially non-overlapping position (at k1, k2) to the control curve ( 59 , 60 ),
• e) the gear change signal line (SL) for the radial adjustment of the cam system element (k1, k2, k3) being operatively connected to the latter,
• f) wherein the control cam carrier unit ( 56 ) with the sun gear condition converter ( 47 ) is identical or in axial driving connection with the sun gear condition converter ( 47 ),
• g) and the geometric relationships are selected so that after the radial adjustment of the cam system element (k1, k2, k3) into its position radially overlapping the cam ( 59 , 60 ) (at k3) the cam ( 59 , 60 ) as a result of the rotation the control curve carrier unit (56) about the geometric axis (A) in engagement with the Steuerkur venanlageelement (k1, k2, k3) may occur and the control curve carrier unit (56) and the Sonnenradzustandwandler (47) due to relative rotation of the control cam (59, 60) and the control curve contact element (k1, k2, k3) experience an axial displacement.

2. Multi-speed hub according to claim 1, characterized in that the control cam assembly of the cam carrier unit ( 56 ) and the contact element equipment of the hub axle ( 10 ) for the axial displacement of the cam carrier unit ( 56 ) are formed in a single direction and that a reset device for the axial displacement of the cam carrier unit ( 56 ) is provided in the opposite direction. 3. Multi-speed hub according to claim 1, characterized in that the control cam assembly of the cam carrier unit ( 56 ) and the contact element equipment of the hub axle ( 10 ) for the axial displacement of the cam carrier unit ( 56 ) are formed in opposite directions. 4. Multi-speed hub according to one of claims 1-3, characterized in that the control cam assembly of the cam carrier unit ( 56 ) and the contact element equipment of the hub axle ( 10 ) for the setting of at least two axial positions of the cam carrier unit ( 56 ) are determined and suitable. 5. Multi-speed hub according to claim 3 or 4, characterized in that two axially oppositely directed control cams ( 59 , 60 ) are arranged to fix the cam carrier unit ( 56 ) in a certain axial position ( Fig. 8a) on the cam carrier unit ( 56 ) and that each this axially oppositely directed control curve is opposed in the axial direction by at least one radially overlapping control curve contact element (k3). 6. Multi-speed hub according to claim 5, characterized in that the envelopes ( 11 , 12 ) belonging to the control cams ( 59 , 60 ) and lying in the axis-normal plane have an axial distance from the control cam support element ( 56 ) corresponding to the respective position of the control cam support unit ( 56 ). k3) or affects it. 7. Multi-speed hub according to one of claims 5 and 6, characterized in that to fix the cam carrier unit ( 56 ) in a certain axial position ( Fig. 8a) each associated with this axial position control system element (k1, k2, k3) or a group of such elements in a circumferential zone ( 58 a) lies or lie between two control cams ( 59 , 60 ) facing each other in the axial direction and these control cams ( 59 , 60 ) radially overlap or overlap. 8. multi-speed hub according to one of claims 5 and 6, characterized, that to define the cam carrier unit in a particular Axial position two control cams facing away from each other one of two control course determining this axial position valve system elements or one of two determine this axial position  groups of control system elements in axial Direction faces and the respective control curve radially overlaps. 9. multi-speed hub according to one of claims 1-8, characterized, that with the existing definition of the cam carrier unit in a certain axial position to initiate the change of this Axial position of a cam system element by changing gear signal line in the direction of an overlapping position to one The relevant control curve can be acted on in such a way that - depending on current angular position of the cam carrier unit - immediately or after rotation of the cam carrier unit in overlapping Position to the associated control cam comes into engagement with this occurs and the cam carrier unit in its further rotation in axially shifts. 10. Multi-speed hub according to one of claims 6-8, characterized in that when the control cam carrier unit ( 56 ) is fixed in a first specific axial position ( FIG. 8a) to bring about another specific axial position ( FIG. 8b), at least one further cam system element (k2 ) can be acted upon axially behind an enveloping circle ( 11 ) by the gear change signal line (SL) in the direction of a radially overlapping position towards the one control cam ( 59 ) belonging to this one enveloping circle (I1) such that it - depending on the instantaneous angular position of the cam carrier unit ( 56 ) - the (I1) associated control curve (59) passes immediately or after rotation of the control curve carrier unit (56) in a opposite to the one enveloping circle (I1) recessed peripheral region (59 n) of an envelope circle in radially overlapping position to said one control cam (59) and immediately or after further rotation of the cam carrier unit ( 56 ) m it engages one control cam ( 59 ) and, after further rotation of the cam support unit ( 56 ), axially displaces it in the direction of the other specific axial position ( FIG. 8b). 11. Multi-speed gear hub according to claim 10, characterized in that during or after the axial displacement of the control vender unit ( 56 ) in the direction of the other specific Axialposi tion ( Fig. 8b) involved in the determination of the first specific axial position ( Fig. 8a) Control system element (k3) can be reset by interaction with the other control curve ( 60 ) and / or by the gear change signal line (SL) in the direction of the radially non-overlapping position to the other control curve ( 60 ). 12. Multi-speed shift hub according to any one of claims 5-11, characterized in that the axial distance (a) between two control cams ( 59 , 60 ) is dimensioned over its entire circumferential length such that it unintentionally simultaneous during the rotation of the control cam carrier unit ( 56 ) radial overlap by cam abutment members (K2, K3) under oscillating axial movement of the control curve carrier unit (56) allow klemmungsfrei that the control curve carrier unit (56) for two particular axially adjacent axial positions (Fig. 8a, Fig. 8b) responsible. 13. Multi-speed hub according to one of claims 1-12, characterized in that a cam system element (k1, k2, k3) in a radial opening (b1, b2, b3) of the hub axle ( 10 ) is guided radially adjustable. 14. multi-speed hub according to one of claims 1-13,  characterized, that a cam system element (k1, k2, k3) from a ball (k1, k2, k3) is formed. 15. Multi-speed hub according to one of claims 1-14, characterized in that a control cam ( 59 , 60 ) of a shoulder ( 59 , 60 ) between the inner circumference offset in relief in the radial direction ( 56-1 , 56-2 ) of the cam carrier unit ( 56 ) is formed. 16. Multi-speed hub according to one of claims 1-15, characterized in that the gear change signal line (SL) is designed or connected to a control body ( 55 ) which is axially displaceable in an axial cavity ( 54-1 ) of the hub axle ( 10 ) and has a control profile ( 61 , 64 a, 64 b) extending in the axial direction for supporting and / or radially displacing a cam system element (k1, k2, k3). 17. Multi-speed hub according to one of claims 13-16, characterized in that a control cam system element (k1, k2, k3) when acted upon by a control cam ( 59 , 60 ) against a radially inward movement at least partially of a peripheral wall of the radial openings ( b1, b2, b3) is supported. 18. Multi-speed hub according to one of claims 1-17, characterized in that the axially movable but non-rotatably arranged on the hub axle ( 10 ) sun gear condition converter ( 47 ) with the cam carrier unit ( 56 ) by a coupling member ( 57 ) has a rotatable connection for the common axial movement of Sun gear condition converter ( 47 ) and cam carrier unit ( 56 ) and, depending on its axial position, holds a sun gear ( 22 , 23 , 24 ) in the state of free rotation or in the state of rotation blocking with respect to the hub axle ( 10 ). 19. Multi-speed hub according to one of claims 1-18, characterized in that the gear change signal line (SL) contains an elastic compensating element which allows the initiation of a gear change signal even when an intended radial adjustment of the cam system element (k1, k2, k3) by the Control cam carrier unit ( 56 ) is inhibited. 20. Multi-speed hub according to one of claims 1-19, characterized in that the at least one cam ( 59 , 60 ) has a profile shape ( 59 a, 60 a) and the at least one cam system element (k1, k2, k3) at least on part of its circumferential profile ) have an outline such that the interaction of this profile shape ( 59 a, 60 a) and this outline supports a shift of the control system elements radially inwards. 21. Multi-speed hub according to one of claims 1-20, characterized in that the rotating hub component ( 120 a) for driving the cam carrier unit is driven by the driver ( 113 ) via the planetary gear system (PGS), the torque flow being driven by the driver ( 113 ) to the rotating hub component ( 120 a) in the course of switching a sun gear ( 122 , 123 , 124 ) between the state of free rotation and the state of rotation lock is interrupted, and that a friction clutch ( 129 ) is provided in order In the event of an interruption, maintain the drive of the rotating hub component ( 120 a). 22. Multi-speed shift hub according to one of claims 1-21, characterized in that the planetary gear system (PGS) is a single planetary gear ( 19 ) with a multi-stage planet gear ( 26 , 27 , 28 ) and one of the number of stages of the planet gear ( 26 , 27 , 28 ) corresponding number of sun gears ( 22 , 23 , 24 ), wherein one of the sun gears ( 22 , 23 , 24 ) are in the state of rotation blocking and the other sun gears are in the state of free rotation. 23. Multi-speed shift hub according to claim 22, characterized in that the driver ( 13 ) can optionally be coupled with the planet gear carrier ( 20 a-bc) and with the ring gear ( 21 ) for common rotation, and that the hub sleeve ( 16 ) optionally with the ring gear ( 21 ) and the planet carrier ( 20 a-bc) is in the output connection 24. Multi-speed shift hub according to claim 23, characterized in that the control cam carrier unit ( 56 ) is located on the side of the planet carrier ( 20 a-bc) close to the driver ( 13 ). 25. Multi-speed shift hub according to claim 24, characterized in that the control cam carrier unit ( 56 ) is in rotary driving connection with the driver ( 13 ). 26. Multi-speed gear hub according to claim 23, characterized in that the control cam carrier unit ( 156 ) is arranged on the side of the planet gear carrier ( 120 a-bc) remote from the driver ( 113 ). 27. Multi-speed shift hub according to claim 26, characterized in that the control cam carrier unit ( 156 ) is in rotary driving connection with the planet gear carrier ( 120 a-bc). 28. Multi-speed shift hub according to one of claims 1-27, characterized in that the rotary driving connection ( 63 or 163 ) between the rotating hub component ( 13 or 120 a-bc) and the control vender unit ( 56 or 156 ) as a friction connection or Overload clutch is formed. 29. Multi-speed shift hub according to claim 23-28, characterized in that the driver ( 13 ) through a ring gear drive clutch ( 44-45 ) with the ring gear ( 21 ) and through a planet carrier drive clutch ( 37-38 ) with the planet gear carrier ( 20 a-bc) is connectable, the ring gear drive clutch ( 44-45 ) and the planet carrier drive clutch ( 37-38 ) are each designed for torque transmission to the ring gear ( 21 ) or the planet gear carrier ( 20 a-bc) only in the forward direction of rotation (VD) and wherein Planet carrier drive clutch ( 37-38 ) can be switched off. 30. Multi-speed hub according to one of claims 23-29, characterized in that the hub sleeve ( 16 ) from the ring gear ( 21 ) and from the planet carrier ( 20 a-bc) ago by means of an output clutch on the internal gear side ( 34-35 ) or an output clutch on the planet carrier side ( 31-32 ) can be driven, the ring gear-side output clutch ( 34-35 ) and the planet carrier-side output clutch ( 31-32 ) for torque transmission from the ring gear ( 21 ) and the planetary gear carrier ( 20 a-bc) to the hub sleeve ( 16 ), respectively are formed only in the forward direction of rotation (VD) and the output clutch ( 34-35 ) on the ring gear side can be switched off. 31. Multi-speed shift hub according to one of claims 1-30, characterized in that it is equipped with a coaster brake ( 39 ). 32. Mehrgangschaltnabe according to claim 31, characterized in that the coaster brake (39) is operable by a brake activating coupling (42-38) extending from the driver (13) produces a brake activation moment only in the reverse rotation direction (RD) of the coaster brake (39) can transmit. 33. Multi-speed hub according to claim 31, wherein for torque transmission from the driver ( 13 ) to the planet carrier ( 20 a-bc) and from the driver ( 13 ) to the ring gear ( 21 ) and for torque transmission from the planet carrier ( 20 a-bc) to the hub sleeve ( 16 ) and from the ring gear ( 21 ) to the hub sleeve ( 16 ) in the forward direction of rotation (VD) effective one-way clutches ( 37-38 , 44-45 , 31-32-34-35 ) are provided and for transmitting a brake activation torque the braking device ( 39 ) is provided a further ON-WAY clutch ( 42-38 ) which is effective in the reverse direction of rotation (RD), characterized in that at least one ( 44-45 ) of the ON-WAY couplings is provided by a reverse rotation sensor system (MKS) controlled switch-off means (BS) can be switched off.