WO1989009351A1

WO1989009351A1 - Cam gearing

Info

Publication number: WO1989009351A1
Application number: PCT/DE1989/000189
Authority: WO
Inventors: Peter Graf Von Ingelheim
Original assignee: Peter Graf Von Ingelheim
Priority date: 1988-03-24
Filing date: 1989-03-23
Publication date: 1989-10-05
Also published as: EP0366737A1; DE3809992A1

Abstract

The gearing disclosed comprises a first gearing element (40) having an axially oscillating first closed cam (41) which moves about the axis of the gearing element, a second gearing element (42) coaxial with the first gearing element (40) and provided with an axial groove (43), a third gearing element (44) coaxial with the first and second gearing elements (40, 42) and having a second closed cam (45) which moves about the axis, and an engaging body (46) which oscillates in the axial groove (43) and which engages in the two cams (41, 45). The two cams (41, 45) are so shaped and the axially reciprocating engaging body (46) is so constructed that the oscillating motion of the latter can be varied in order to control the torque transmitted by the gearing. In another embodiment, the torque ratio or speed ratio can be varied, a further gearing element can be driven, or automatic interlocking can be achieved.

Description

Controllable cam mechanism

The invention relates to a device for the positive transmission or conversion of forces and movements. Such mechanical devices can be used as gears, clutches, differentials (differential) and brakes.

Positive gears for the transmission of uniform rotary movements are generally designed as toothed gears. Their disadvantages are the comparatively large installation space and certain basic conditions that apply to all previous gear transmissions. For planetary gear units with shafts Wl, W2 and Ws, the main speed equation of a planetary gear unit applies: z = (nl - ns) / (n2 - ns), (1) where z = construction ratio of the gear unit and ni = speed of shaft Wi;

Ml + M2 + Ms = 0, (2) where Mi = torque on shaft Wi;

M2 / M1 = -z; (3) and the power equilibrium equation Pl + P2 + Ps = 0, (4) where Pi = power on shaft Wi.

Another possibility is to transmit forces and movements with the aid of cam mechanisms.

From DE 35 05 396 AI a gear is known in which grooves are made on the jacket of a driving drum at an angle unequal to 90 ° to the axis in which balls roll. Axially movable lifting bodies with hemispherical bearing shells in which the balls engage are mounted in the output shaft. If the input and output shaft have a speed ratio not equal to 1, the balls roll along in the grooves and move the lifting bodies back and forth. Hydraulic force acting on the lifting body allows this to be drive shaft influencing flowing torque.

A disadvantage of this transmission is that the link on which the controlled force is applied is non-rotatably connected to the reciprocating lifting bodies and therefore rotates with them. Such a control can therefore only be implemented with great effort.

A second disadvantage of this transmission is that it does not make it possible to translate the torque. The torque on the output shaft is always the same as the torque on the drive shaft. Thus, such a transmission is one

A third transmission element would be necessary for torque transmission in order to be able to initiate a supporting torque Ms; see. the above equation (2).

The invention provides a remedy here. On the one hand, it enables a controlled application of force to non-rotating ones

(but axially oscillating in the housing) gear elements, which enables easy control. On the other hand, the invention makes it possible to implement controllable transmissions with three torque shafts for real torque transmission. These gears can be built both as conventional epicyclic gears, which can even be designed without play, and as epicyclic gears, which have a different main speed equation than conventional planetary gears, or can be used in gear-planet gears for speed control of the planet gear.

This is achieved in that the transmission contains three coaxial or parallel-axis transmission elements, one of which is provided with axial slots, in which engagement members can be moved axially back and forth, and the engagement members act in such a way that (self-locking) transmissions which can be fixed in the power flow can be realized are and / or permit control of the torques transmitted by the transmission or change in the speed ratios of the transmission shafts.

Exemplary embodiments of the invention are shown in the drawings. Show in it

1 shows a transmission in perspective, FIG. 2 shows a clutch with recovery of the previous clutch losses due to springs, FIG. 3 shows the same coupling principle in a different embodiment,

4 shows an arrangement with clutches according to FIGS. 2 and 3 to compensate for different speeds of the input and output shaft,

FIG. 5 shows an arrangement which can be used as a clutch or gearbox and has a directed power flow,

6 shows a transmission circuit with a transmission according to FIG. 5, FIG. 7 shows a transmission arrangement with the main speed equation nl-n2 = z.n3,

FIG. 8 another gear arrangement with the new main speed equation,

FIG. 9 shows a gear arrangement for a controllable locking differential,

FIG. 10 shows a gear arrangement for a continuously variable gear,

FIG. 11 shows a gear arrangement for an actively controllable differential, and FIG. 12 shows a gear arrangement for a starting clutch.

Figure 1 serves to illustrate the spatial structure of the transmission according to the invention using an epicyclic gear of the construction ratio z = -1 (differential). The two output shafts 1, 2 of the differential are provided with oppositely oscillating bead curves 3, 4 of the same number of phases. Around the two output shafts 1, 2 is the drive shaft 5 of the differential, which has a slot 6. In the slot 6, an engagement body 7 can be moved back and forth, which rolls with wheels 8 on the bead curves 3, 4 of the output shafts 1, 2.

If the drive shaft 5 is rotated and the torque on the two output shafts 1, 2 is the same, the engaging body 7 remains stationary in the slot 6. The two output shafts 1, 2 then have the rotational speed of the drive shaft 5. If one of the two output shafts has a greater torque than the other, it has the tendency to rotate more slowly. The engaging body 7 is then moved back and forth in the slot 6 and thus rotates the shaft with the lower torque forward. It should be noted that if the engagement body 7 is guided without play in the slot 6 and the wheels 8 along the bead curves 3, 4, the transmission becomes free of play.

FIG. 2 shows the principle for a clutch according to the invention with recovery of the previous clutch losses. The drive shaft 40 contains an oscillating groove 41 in which a ball 46 is located as the engagement body. The output shaft 42 is coaxial with the drive shaft 40 and engages with a tube around it. In the tube there is a slot 43 in which the ball 46 can roll back and forth. Around the input and output shaft there is a further tube 44 with an annular inner groove 45, into which the ball 46 also engages. The tube 44 can be moved axially back and forth against the force of springs. The spring force can be changed by axially adjusting adjustment rings 48 which are otherwise fixed to the housing.

This coupling works as follows. If the drive torque is sufficiently small, the transverse forces generated are not sufficient to push the springs 47 through. Input and output shaft 40, 42 are coupled. If the output torque is sufficiently large, it compresses the springs 47 and the clutch slips.

FIG. 3 shows the same coupling principle with an input shaft 50, in which sinusoidal grooves 51 are formed, and an output shaft 52, which surrounds the input shaft 50. The engagement body 53 is a ring with spikes 54 which engage in the grooves 51. The engagement body 53 slides back and forth in a slot of the output shaft 52. Furthermore, it slidably engages with an annular bead 55 in an annular groove 56 of a third gear element 57. This gear element 57 is supported by springs 49 with respect to a component that can be axially displaced in the housing 60 and engages with spikes 58 in axial grooves 59 in the housing 60; it is therefore only axially movable.

Figure 3 is primarily used to understand the following gear arrangements. Previous safety clutches (slip clutches) convert mechanical energy into unusable thermal energy when it slips. The mechanical drive power Pan = Man. nan. The mechanical output power Pab = -Man stands 2π. nab. 2π opposite. The differential power Pan - Pab is converted into thermal energy in the coupling. In the coupling principle according to FIGS. 2 and 3, the previous "loss energy" is stored by springs 47 and 49, respectively, and is returned when the springs relax.

The power flow can be outlined as follows. If M is the average torque that is needed to compress the springs, then it acts as a brake on the input shaft and accelerates on the output shaft. When relaxing, it works the other way round. If the time for pressing together and relaxing were the same, no power could be transferred from the input shaft to the output shaft. If tl is the time for compressing and t2 is the time for relaxing the springs, then the pulse M acts on the output shaft. tl accelerating and M. t2 wide. The average torque Md acting on the output shaft during a sine phase is then calculated

Md = M. (tl - t2) / (tl + t2). (5)

This time difference can additionally be influenced positively by different steepnesses of the sinus curves for the compression and relaxation phase. In addition, a flywheel with periodic change in the moment of inertia is shown in FIG. 4 as an energy store and explained further below.

FIG. 4 shows an application of the coupling according to the invention. In motor vehicles, a number of auxiliary units are driven by the motor shaft (lubricating oil pump, alternator, auxiliary pumps, etc.). That is, even though they have to do their job at the lowest engine speed, their speed increases linearly with the engine speed. So they take up power that would not be necessary for their satisfactory work in the engine.

Would you e.g. If a centrifugal pump is used in conjunction with a coupling according to the invention, the lubricating oil pump could be made very small and economical. The power consumption curve of a centrifugal pump increases exponentially with the pump speed. So far, it has not been suitable for an internal combustion engine with large differences in speed.

If one now closes such a centrifugal pump 61 with the clutch 62 according to the invention to a motor shaft 63, the clutch will slip from a certain pump speed. A flywheel 65 acts as an energy store, which is fixed to the pump shaft 64 and in which the masses 66 are moved radially inwards or outwards depending on the clutch or spring position, as a result of which the moment of inertia of the flywheel 65 is changed.

Figure 5 shows the principle of a transmission with directed power flow. Such mechanical transmissions have so far only been known as worm gear transmissions and other high-ratio transmissions with self-locking. This self-locking can be easily realized by different steepness (ratio of amplitude to phase length) of two sine curves. With the same wavelength range, sine curves with different number of phases and the same amplitude have different slopes. With the gearbox according to the invention it is also possible to realize clutches (1: 1 ratios) with self-locking. This serves for safety and precision.

On the drive shaft 70, a sinusoidal groove 71 is formed on a large cylinder circumference. The steepness of this sine curve is therefore low. A sinusoidal groove 73 is provided on the output shaft 72 on a small cylinder circumference. The steepness of this sine curve is great. Engagement bodies 75 engage in the input and output shafts 70, 72, which can be moved back and forth in slots 77 in the housing 76 and bridge the difference in diameter.

If you turn the drive shaft 70, the power flow takes place from the large circumference to the small one. The axial force on the engagement body 75 is a multiple of the peripheral force on the drive shaft 70. On the output shaft 72, the peripheral force is a multiple of the axial force. The output shaft 72 is therefore rotated. In the opposite direction, no power flow is possible because the transmission stops due to friction. Such a transmission can also be designed without play. It is therefore very well suited for precision drives (leaving a precisely approached position only through active adjustment).

Figure 6 shows the circuit of such a clutch in Connection with a hydraulic motor. The transmission 81 of FIG. 5 is arranged between the hydraulic motor 80 and the working machine 82. Leakage losses of the hydraulic motor 80 and the compressibility of the hydraulic oil cannot lead to the shaft of the working machine 82 adjusting itself.

FIG. 7 shows a transmission according to the invention with the main speed equation nan - nab = ns. (6)

Such a main speed equation cannot be implemented with known epicyclic gears. The drive shaft 90 has two single-phase sinusoidal grooves 91, 92 (oblique ring grooves) into which spikes 93 of an engagement body 94 engage. The engagement body 94 slides in slots 95 of the output shaft 96. An annular bead 97 can be rotated in an annular sleeve 98 of the third gear element, an annular sleeve 99. These parts thus correspond to the analog parts in FIG. 3.

The back and forth movement of the sleeve 99 at unequal speeds of the input and output shafts is used to drive a further shaft. In FIG. 7, a wheel 101 is rotated at the speed ns via a crankshaft drive 100.

FIG. 8 shows a similar gear, in which the engagement body 115 engages in an oscillating groove 119 in the drive shaft 113. Furthermore, a sleeve 110 with spikes engages in an oscillating curve 111 in the housing 112. By reciprocating the engagement body 115 with relative rotation of the drive shaft 113 and the output shaft 114, the sleeve 110 is moved along in the oscillating curve 111 and therefore rotated. The sleeve 110 in turn engages with engagement bodies 116 in slots 117 of the support shaft 118 rotating at the speed ns. Thus, only the engagement body 115 and the sleeve 110 execute oscillating and rotating movements. The three shafts 113, 114, 118 of the transmission are exclusively rotated. Such transmissions are predestined for use in torque converters. If the output shaft 96 or 114 of the clutch is non-positively connected to the adjusting motor of a hydrostatic transmission and the support shaft 111 or 118 with a constant pump that delivers to a hydraulic motor, an extremely economical hydraulic transmission can then be realized. If the hydraulic motor is in the zero position (swallowing quantity = 0), the pump cannot be turned. The gearbox is clutched. When the hydraulic motor is swung out, the output shaft rotates more slowly than the drive shaft and the pump is rotated.

The advantage of such a transmission arrangement lies in the fact that only a part of the drive power flows through the hydrostatic transmission and the rest as coupling power with high transmission efficiency via a compact mechanical transmission. Therefore, a small hydrostatic transmission can be used, and the transmission efficiency is very high due to the high proportion of clutch power. FIG. 9 shows the use of the transmission according to the invention for controlling a locking differential. The drive shaft 121 of the differential, which is driven by the cardan shaft 120, is connected to the drive shaft 122 of the transmission 123 according to the invention, and the one output shaft 124 of the differential is the output shaft of the transmission 123. The support wheel

125 is stored in the housing 126. At the same speed of the input and output shafts 121, 122 of the differential, the support wheel 125 stands still. If the input and output shafts 121, 122 of the differential rotate relative to one another, the support wheel 125 also rotates. The support wheel speed and thus the relative rotation of the input and output shafts 121, 122 can be regulated via a brake or friction clutch .

FIG. 10 shows an arrangement for controlling the planet gear speed of a gear planet gear with the aid of the gear according to the invention. This is not possible with conventional transmissions or only with an unacceptable effort.

In previous planetary gears, the planet gears are considered to run freely. There is therefore an equilibrium of moments due to the moments brought in by the sun gears. This results e.g. the unconditional validity of the torque equations for planetary gears.

By controlling the planet gear speed by means of the transmission according to the invention, additional torque can be generated be introduced. In particular, it can be shown that there is a functional relationship between the planetary gear speed and the shaft shaft speed. Therefore, if the control shaft for the planetary gear rotation and the spline shaft are positively connected to one another, an inexpensive continuously variable branching transmission can be realized, in which only a small power share is achieved via the continuously variable transmission (hereinafter "CVT transmission" = continuous) Variable transmission) flows.

FIG. 10 shows a concept for an adjusting gear according to the invention with a power split. A four-wheel planetary gear 165 with two sun gear shafts 166, 167, two planet gears 168, 169 and a web 170 is used. A pair of wheels 171 is mounted on the web 170, which is driven by the planet wheels 168, 169 and rotates with the web 170. This pair of wheels 171 in turn can drive a sun gear 172 coaxial with the sun gear shafts 166, 167, _' which contains a slot 174 in its shaft 173, in which an engagement body 175 is axially displaceably mounted. The web shaft 170 is axially elongated and contains an oscillating groove 176, in which the engagement body 175 engages with a mandrel 177. With an annular bead 178, the engaging body 175 also engages in an annular groove 179 of a ring 180 surrounding the three shafts 166, 170 and 173. The ring 180 in turn engages with an arbor 182 in an oscillating groove 183 in the housing 185 the ring 180 also in

Slots 187 of a fourth sun gear 190 and is in rotationally fixed contact.

The arrangement of web 170 with groove 176, sun gear 172 with shaft 173 and slot 174, engagement body 175 with mandrel 177 and bead 178, ring 180 with groove 179, mandrel 182 and engagement bodies 186, housing 185 with groove 183, and sun gear 190 with slots 187 corresponds exactly to the transmission of Figure 8 with the new main speed equation (6).

If the sun gear 190 is non-positively connected to the shaft shaft 170 by means of a hydrostatic adjustment gear with an adjustment motor 191 and an adjustment pump 192, then an enormously inexpensive continuously variable adjustment gear is obtained. If z = construction ratio of the gear planetary gear, i = gear ratio set = nan / nab = = Mab / Man, then

Ms = torque on the web = (i-1) .Man, ns = speed of the web shaft = (z.nab - nan) / (z-l). The power Ps flowing over the web is then:

Ps = Ms.ns = (i-1). (z.nab/nan-1) / (z-1) .Man.nan =

= Man.nan. (i-1). ((z / i) -l) / (z-l). (7)

It can similarly be shown briefly that the power flow via the planet gears obeys precisely this formula.

Note: In the torque formulas above, the transmission ratio i can now be used instead of the construction ratio z, since the planet gears no longer have to reach an equilibrium state, but are controlled in a targeted manner by an additional torque.

As can be seen, the power flow through the hydrostatic transmission is low. For i = 1 and i = z, no power flows through the hydrostatic transmission. In any case, the power flow through the CVT transmission is low.

An example is a continuously variable transmission with a gear ratio of 1: 1 to 1: 4 (z = 4). The following applies to this gearbox:

Gear ratio: 1: 1.5 1: 2 1: 3 1: 3.5

Power flow via hydrostatic transmission (% of total power): 0.28 0.33 0.22 0.12

For a converter with a gear ratio range from 1: 1 to 1: 2, the following applies (z = 2):

Gear ratio: 1: 1.2 1: 1.4 1: 1.6 1: 1.8 power flow "over

Hydrostatics (in%) 0.1333 0.1714 0.15 0.08

FIG. 11 shows a controllable differential for the targeted regulation of the speeds of the two output shafts. Such a differential provides advantages over "controllable limited slip differentials", especially when cornering. When cornering, an accelerating force vector from the road acts on the wheel on the inside of the curve. This reduces the torque on the two output shafts and the desired power can no longer be brought to the road. With a targeted speed control depending on the vehicle speed and the

Steering wheel lock can always bring the desired performance to the road even in curves. If the possibility of self-locking is taken into account when determining the differential, a different torque on the vehicle wheels cannot change the set speeds.

Again, a transmission according to the invention with the new main speed equation (6) is used. The propeller shaft 200 of the vehicle rotates the drive shaft 201 of the differential. In the slots 202 of this drive shaft 201 there are four engagement bodies, of which only the lower 203 and the upper 204 can be seen in FIG. 11, which engage with thorns in oscillating grooves 222, 223 of the two output shafts 207, 208. The two further (not visible) engagement bodies are in the direction of view in front of or behind the output shafts 207, 208 and engage in further oscillating grooves 224, 22-5 in these. With another mandrel or annular bead, the engagement bodies 203, 204 engage in annular grooves of sleeves 212, 213, while the two further engagement bodies, not visible in FIG. 11, engage in the sleeves 210, 211. The sleeves 210 ... 213 are axially displaceable in axial slots 215 of the housing 216. With mandrels 218, 219 they also engage in oscillating grooves 220 of a control shaft 221 coaxially surrounding all shafts. The control shaft 221 can be rotated by an adjustable hydraulic motor.

The differential works as follows. If the control shaft 221 is not rotated, the two output shafts 207, 208 of the differential have the same speed (the differential is locked). If the control shaft 221 is rotated in one direction, e.g. the right output shaft 208 faster than the left; if the control shaft 221 is rotated in the other direction, the left output shaft 207 rotates faster than the right. The speed differences of the two output shafts 207, 208 depend on the speed of rotation of the control shaft 221.

Figure 12 shows a clutch for torque-free starting of motors on work machines. A high torque is often required to overcome the static friction of stationary machines necessary. However, electric motors already require considerable energy to start up. This often results in high switching impulses.

Centrifugal clutches are known which produce the frictional connection between two shafts only from a certain engine speed. Such centrifugal clutches can only be used at high shaft speeds.

The transmission according to the invention according to FIG. 12 enables a starting clutch even for low shaft speeds. The drive shaft 230 has an oscillating groove 231 in its circumference. The output shaft 232 has axial slots 233 in which balls 234 can be moved back and forth as engagement bodies. The third coaxial gear element is a ring 235 with a likewise oscillating groove 236, which lies around the input and output shaft 230, 232. The balls 234 engage in the two grooves 231, 236. The output shaft> 232 and the ring 235 are connected to one another via spring elements 237.

The clutch works as follows. In the start-up phase, when a high torque acts on the output shaft 232, the ring 235 rotates around the two shafts since it does not yet provide any support torque. Only when the springs 237 are tensioned to such an extent that the ring 235 and the output shaft 232 are connected with a sufficiently high positive connection is the output shaft 232 moved as well. Such a clutch therefore also enables a very smooth start-up, since the transmitted torque gradually increases (start-up ramp).

Claims

1. Gearbox with a first gear element (40; 70; 90) which has a closed curve (41; 71; 91, 92) which oscillates in the axial direction and runs around the axis of the gear element, one with an axial slot (43; 77; 95 ) provided with the first gear element (40; 70; 90) coaxial or parallel axially second gear element (42; 76; 96), a third gear element (44; 72; 99 coaxial or parallel to the first and second gear element) ), which is provided with a second closed curve (45; 73; 98) running around the axis, and one which can be moved back and forth in the axial slot (43; 77; 95) and into the two curves (41, 45; 71, 73; 91, 92, 98) engaging engaging body (46; 75; 94), characterized in that the two curves (41, 45; 71, 73; 91, 92, 98) are shaped so that the axially and movable engaging body (46; 75; 94) is constructed or provided with a device such that the reciprocating movement of the engaging body body (46; 75; 94) for controlling the torque transmitted by the transmission or the torque or speed ratio or for driving a further transmission element (101) or for self-locking.

2. Transmission according to claim 1, characterized in that the two curves (3, 4) on the outer circumference of the first and the third transmission element (1, 2) are arranged.

3. Transmission according to claim 1 or 2, characterized in that the first transmission element (70) form the drive shaft and the third transmission element (72) form the output shaft of the transmission, that on the drive shaft (70) and on the output shaft ( 72) provided curves (71, 73) oscillate in the axial direction, and that the diameter of the drive shaft (70) is much larger than the diameter of the output shaft (72) and the curve (71) provided on it has a slope which effects self-locking .

4. Transmission according to one of claims 1 to 3, characterized ge indicates that the engagement body (46) is connected to a periodically acting energy store (e.g. springs 47, flywheel 65).

5. Transmission according to claim 4, characterized in that it is inserted between a rotating at a variable speed drive shaft (63) and an output shaft (64) as a speed of the output shaft (64) limiting clutch.

6. Transmission according to one of claims 1 to 5, characterized ge indicates that 'it is a first shaft (113) rotatable with the speed nan, a second rotatable with the speed n

Shaft (114) and a third shaft rotatable at the speed ns

(118) and the speeds of the main speed equation

are sufficient, where z is the construction ratio of the transmission determined by the phase ratio of the curves (119, 111).

7. Transmission according to one of claims 1 to 6, characterized ge indicates that it is used for controlling a differential gear.

8. Transmission according to one of claims 1 to 7, characterized ge indicates that it is used for supplying or removing power to or from the planet gears (168, 169) of a planetary gear.

9. Transmission according to one of claims 1 to 8, characterized ge indicates that the first, second and third gear elements (230, 232, 235) are rotatable and two of these gear elements (232, 235) via one Spring (237) are connected.