DE4131847C1 - Control and conversion engine etc. drive - has several rotors, each with oval gearwheels, eccentric, parallel, and symmetrical w.r.t. rotor axis - Google Patents

Abstract

The drive has a housing with two irregularly rotating rotors, forming working chambers, and non-circular toothed gears to control the rotary oscillating movement of the rotors. Eahc rotor has at least two similar eccentric, parallel, symmetrical oval gears, with an externally toothed belt on the jacket. Each oval gear (12) of the first rotor (2) meshes with two mirror-like oval gears (13) of the second rotor (3), with varying axle distances. The oval gear set controls the rotary angle differential of the two rotors. At least one oval gear is planetary guided w.r.t. the housing (1), directly or via more rotary parts. USE/ADVANTAGE - For engines, turbines etc., with compact design, small number of components and quiet run.

Description

The invention relates to a control and converter transmission, in particular for power and and work machines, with at least two in one housing Coaxially rotatably mounted, working chambers forming, non-circular rotating rotors and with the oscillatory rotations of the rotors control non-circular Gears.

The main use is drive and / or control of Oscillating rotating coaxial rotors with positive displacement for converting energy into power and work machines for liquids, vapors, gases and burning Gas-air mixtures. This requires in addition to compact Structure, high efficiency and low construction costs in addition good start-up behavior and good dynamic behavior without imbalance or play. In addition, one is from the rotor rotation angle linearly independent and Auslaßsteuerung advantage, an influenceable dependence the flow rate of the rotation angle for optimum Function indispensable.

In DE-OS 23 54 531 is a "Control for rotary engines with circular working cylinders" described. Here effect by a control drum guided, planetary driven eccentric over Push rods and lever the torsional vibrations of a central shaft and a hollow shaft. This voluminous arrangement of consistently well-known machine elements comes the demand only conditionally contrary to a dynamic drive: Imbalance occurs due to reciprocating components; Load fluctuations on the planet wheels cause noises, Wear and losses. A flexible adjustment of volume flow to angle of rotation is not possible.

The FR-PS 4 96 131 shows in addition to a precise description the segmental work spaces within two rotors an embodiment of a control gear with elliptical gears: A first rotor is connected to the working shaft, the second rotor with hollow shaft postponed. Two side by side on the hollow shaft end  and mesh elliptical gears mounted on the shaft each with one mounted on a fixed distance Secondary shaft twisted attached elliptical gear. With almost the same transmission structure continue to operate the EP-OS 00 62 087 and DE-PS 36 23 969.

In the US-PS 37 30 654 is also the same Structure as selected in FR-PS 14 96 131, different from this However, the Wälzkurvenform of a gear pair sector by sector composed of logarithmic spirals.

These almost classic gearbox versions with non-round Gears leave a certain vote of volume flow to angle of rotation too. Those used in US-PS 37 30 654 Gears achieve a better intake and Auslaßsteuerung, however, are bad to manufacture, effect sudden changes in the angular acceleration in the Dead center area and work because of steeply rising Wälzkurven with bad engagement of the teeth. all common to these transmission designs is the annoying Backlash at load change, the constructive condition occurs several times per revolution. Consequence are noises, Wear and losses.

The invention is based on the object with a compact Construction using as few as possible, purely rotatable parts as smooth as possible and backlash-free working gearbox for drive and / or control of oscillating rotating coaxial rotors, in addition, by the principle to be provided possibility the coordination of individual elements a far-reaching Variation and thus optimal adaptation of the characteristic Volume flow / angle of rotation to the respective operating purpose allows the energy converter machine.

This task is done with a control and converter gearbox for power and working machines with non-uniform circumferential, coaxial rotors and non-circular gears according to the invention solved by everyone  Rotor at least two same, point symmetrical, in the same plane and eccentric to the rotor axis, parallel and symmetrically arranged, rotatably mounted oval wheels with an externally toothed Belt on the coat has, with each oval wheel of the first rotor with two in addition mirror-image executed oval wheels a second rotor in terms of a speed-identical, alternately opposing rotation, with varying Center distances, combs, so that this with two same pairs same web and rotorwise synchronously running Oval wheels formed Ovalradsatz the rotation angle difference of two rotors controls; that the geometric shape of this point-symmetric Ovalradwälzkurven the polar equations

is sufficient, where e is the radius at the point of engagement, ε the e associated polar angle, γ the Ovalradeigenddrehung with respect to the tangent T 3 at the point of engagement, δ half the rotation angle difference of two rotors as a function of γ, δ 'the differential quotient dδ / dγ, f the fixed distance Ovalradachse Rotor axis mean; in that at least one oval wheel is forcibly guided in a planetary direction, directly or via a connecting shaft, by means of further rotating components, so that the gear has a guided rotor and at least one further loose rotor entrained via the oval gears, and in that the external mechanical coupling on one of the Rotors directly or via other rotating components and a drive shaft takes place.

Design options for a dead center operation, especially for low-speed engines are important, are by interconnecting several, segmental working chambers forming rotors given. This can be mechanical  in series one behind the other, as characterized in claims 4, 19, 22 or by pairwise parallel coupling accordingly Claim 20 or 21 take place.

Other possibilities of dead center bridging operation are characterized by those in the dependent claims 3, 17, 18 Characteristics of the formation of variable, coaxial Gear chambers by the oval wheels according to the invention given inside the rotors.

The unbalance-free functional design more important Transmission details with variants also with regard to backlash is characterized by the subclaims 5 to 14.

Characterized by the dependent claims 2, 15, 16 are two the possible embodiments of the transmission characteristics determining periodic function δ in dependence of γ, which together with the frame formulas (e) and (ε) determine the oval wheel geometries.

The advantages achieved by the invention are:

• 1) Very good dynamic behavior, achieved by pure rotating machine elements without imbalance and low backlash or play-free interaction of the components also with functional internal load changes.
• 2) functional and production-related design of Toothing provided rolling elements by avoiding Kinks and extremely rising corners.
• 3) Good adaptation to the respective purpose by combined use of several inventive Properties, also known.
• 4) Possibility of installing one of the rotor rotation angle linearly independent inlet and outlet control.
• 5) Largely influenced by Ovalradwälzkurvenvariation Transmission characteristics up to fluid-constant Motor / pump.

Further developments relate to further embodiment the variants characterized in the subclaims 4, 18, 19, 22 mechanically coupled in series one behind the other Rotors.  

With reference to the embodiments shown in the figures the invention will be explained in more detail below. In the drawing, three embodiments of the subject of the invention shown, two of them with sub-variants. It shows

Fig. 1 a control and torque converter according to the invention formed by the chambers and Ovalradansatz gear with planetary forced guidance of the oval wheels are non-circular planetary gears in the section corresponding to the line AA of FIG. 2.

Fig. 2 shows an embodiment of a control and converter transmission of Fig. 1 with planetary coupling of the oval wheels to the drive shaft via non-circular gears in an axial section, shown in simplified. Fixed parts have the same hatching.

Fig. 3 shows an embodiment of a control and converter transmission of Fig. 1 with representation of the control channels in a sub-variant with coupling of the eccentric drive shaft via non-circular gears on the loose rotor in an axial section. Fixed parts have the same hatching.

Fig. 4 is a set of non-circular gears for coupling the off-center drive shaft to the loose rotor the section of the line BB of FIG. 3 accordingly.

Fig. 5 is an inlet and outlet control for the segmental working chambers in section along the line CC of Fig. 3. Firmly connected parts have the same Schaffur.

Fig. 6 shows a second embodiment of a control and converter transmission with two parallel coupled, offset working pairs of rotors in play-free design of forced operation and drive and the oval wheels in an axial section. Fixed parts have the same hatching.

Fig. 7 is a play-free drive of the planetary wheels over Oval crank disc in section along the line DD of Fig. 6.

Fig. 8 is a backlash-free forced guidance of the planetary wheels over Oval crank disc in section along the line EE of Fig. 6.

Fig. 9 is a sub-variant of the second embodiment shown in Fig. 6 with off-center drive shaft and double the number of discs. Fixed parts have the same hatching.

Fig. 10 shows a third embodiment of a control and converter transmission with three backlash in series successive, offset working rotors in the housing, backlash planetary positive guidance of the oval gears on crank disks, as well as planetary gear with non-circular gears coupled drive shaft in axial section. Fixed parts have the same hatching.

Fig. 11 shows the graphical representation of a first possible function δ as well as its first and second derivative of δ 'and δ''as a function of γ.

Fig. 12 shows the graph of a second possible function δ as well as its first and second derivative of δ 'and δ''as a function of γ.

FIG. 13 shows an oval wheel according to the invention which effects the first transmission characteristic illustrated in FIG. 11.

Fig. 14, the Wälzkurvenkontur an oval wheel according to the invention, which causes the position shown in Fig. 12 second possible transmission characteristic.

Fig. 15 is a non-circular gear of a gear set for planetary forced guidance and propulsion of wheels of the oval contour of FIG. 13.

Fig. 16 to 21 a Ovalradsatz invention with oval gears of FIG. 13 in different operating positions, executes moderately by increasing Fig.-Nr. orderly. The mark X serves as a guide.

Fig. 22 to 27 a according to the invention, the forced guidance of the oval wheels effecting planetary gear set having wheels according to Fig. 15 in different operating positions, by increasing Fig.-Nr. assigned to the oval wheel positions of Fig. 16 to 21.

In a first embodiment, two (Fig. 1, 2) form by a Ovalradsatz 12, 13 (Fig. 1) controlled, in a housing 1 coaxially rotatably mounted rotors 2, 3 (Fig. 1, 2) a four oval wheels 12, 13 cup-shaped enclosing coaxial gear chamber 9 ( FIG. 2), which is formed on the cylinder jacket-shaped butt joint of both rotors 2, 3 ( FIG . Within the gear chamber 9 , the four oval gears 12, 13 form two variable coaxial gear chambers, the central gear chamber 10 and the orbital transmission chamber 11 ( FIG. 1). For supporting the oval gears 12, 13 are both circular end walls of the gear chamber 9 each with two at a fixed distance f from the center symmetrically parallel to the axis mounted cylindrical bearing pins 21, 22 ( FIG. 2) provided, the supernatant into the gear chamber 9 more than half the gear width is and as well as their diameter with the dimensions of the oval wheels 12, 13 centrally mounted blind bores 23 ( Fig. 2) in the sense of a smooth, low-play, sealing slide bearing is tuned. Furthermore, the four oval wheels 12, 13 exactly equal width with plane-parallel faces are designed so that they touch on both sides of the plane-parallel executed circular frontal inner surfaces of the gear chamber 9 in the sense of a low-backlash, smooth, sealing, axial sliding. The externally toothed belt 19 (FIGS . 1, 2) provided on the oval wheel casing is formed over the entire oval wheel width, so that the annular partition between the two gear chambers 10, 11 in the four regions of the oval wheel contact is completed by the labyrinth sealing effect of two meshing gear wheels known per se. The Ovalradwälzkurven are formed according to the above frame formulas (e) and (ε) and in this embodiment by the transmission characteristics of the formula

δ = π / 4 + (δ₀-π / 4) cos (2γ)

certainly, where π = number of circles = 3.1415. , , means and for this Wälzkurvenform δ₀ = 0.9356 <0.9603 [rad].  

This function δ, which is shown graphically in FIG. 11, is chosen here such that the shapes of the oval wheel gear curves are twice axially symmetrical and thus congruent for all oval wheels. Table I contains some associated numerical values, Fig. 13 the assignment. The number of teeth is odd with symmetrical arrangement to the larger axis of symmetry, so that all four oval wheels are the same. Further technical data and information on the theory are listed below.

Outside the gear chamber 9 , the rotors 2, 3; 2, 303 in a known manner by two hollow cylindrical segments 28 ( Fig. 5) four segment-shaped working chambers 29 ( Fig. 5), which form in combination with the gear chambers 10, 11 a dead center force or work machine. For this purpose, the jacket of the outer, cup-shaped rotor 3, 303 is axially extended beyond the butt joint that terminates the gear chamber 9 and has on its inner periphery a reinforcing ring with two radially inwardly projecting segments 28 ( FIGS. 2, 5). The inner rotor 2 , which forms the cover of the gear chamber 9 , is continued axially as a cylindrical core piece with two radially outwardly projecting segments 28 ( FIG. 5) and closes off the working chambers 29 by a second circular wall. The cylindrical core of the inner rotor 2 has a central bore for storage through a central fixed journal journal 32 ( Figure 2). The two segments 28 of the inner rotor 2 are fixed relative to the two bearing journals 22 in an aligned angular position. Both bearing journals 22 have central, axial, continuous bores 26 , which in extension pierce both end walls and the segments 28 therebetween and are designed as slide bearings.

The planetary positive guidance takes place in this embodiment on both oval gears 12 of the inner rotor 2 via bore side coaxial with the oval gears 12 mounted in the bores 26 connecting shafts 16 ( FIG. 2), two mounted on the protruding ends of the connecting shafts 16 , same planetary gears 34th (Fig. 1, 2) and a rotorkoaxial in the housing 1, 301 secured, with two planetary gears 34 meshing with the first central gear 35 ((Figs. 1, 2), is attached to the inside of the central bearing journal 32. the non-circular planetary gears 34 have functional two The circumferential ratio of planetary gear 34 to central gear 35 is selected to be 2: 2 in this exemplary embodiment The rolling curves of the three-piece gearset of externally toothed non-circular gears 34, 35 are identically designed according to the polar equations

l = f * (1 + δ ') / 2 (1) and

λ = γ + δ₀ - δ (λ)

where l is the Wälzkurvenradius, λ the l associated polar angle. Table II contains some associated numerical values, Fig. 15 shows the assignment. The curve is twice axisymmetric and the number of teeth evenly divisive to four, so that when symmetrical teeth arrangement three equal non-circular gears 34, 35 are formed. Further technical data and information on the theory are listed below.

The coupling to a rotor coaxially mounted drive shaft 38 ( FIG. 2) takes place in this exemplary embodiment via a three-part planetary gear set consisting of a second central gear 37 ( FIG. 2) and two planetary gear wheels 36 , whose gear curves and gear geometry are identical to the non-circular gear wheels 34, 35 are formed. For this purpose, that is with two planetary gears 36 (Fig. 2) meshing second central 37 (FIG. 2) mounted within the housing 1 between the outer rotor 3 and the housing end wall on the drive shaft 38. This is on the one hand inside the housing with its protruding end in a central blind bore of the outer, mounted on the circumference of the rotor 3 ( Fig. 2) supported, on the other hand led out of the housing 1 by a dynamically tight bearing bore. The two non-circular planetary gears 36 are by coaxial bearing bores 26 in the journal 21 (Fig. 2) guided link shafts 17 (Fig. 2) with the oval gears 13 (Fig. 2) of the outer rotor 3 is fixedly connected.

In a possible sub-variant of the first embodiment, in contrast to the above embodiment, the mechanical coupling via an eccentric, axially parallel, offset to the rotors, in the housing 301 ( FIG. 3) rotatably mounted drive shaft 338 ( FIG. 3) and a two-part set unround, externally toothed gears. The first gear wheel 339 ( FIGS. 3, 4), which is referred to as the primary wheel, is centered and fixed on a journal mounted on the front side of the outer rotor 303 ( FIG . The second gear 340 ( FIGS. 3, 4), referred to as a secondary gear, is coaxially mounted between the housing end wall and the outer rotor 303 at the inner end of the drive shaft 338 . The rolling curves of primary wheel 339 and secondary wheel 340 are embodied here in the circumferential ratio 2: 2 according to the following polar equations:

m = g / (2-δ ') (m)

μ = γ + δ₀ - δ (μ) (primary wheel)

n = g * (1-δ ') / (2-δ') (n)

ν = γ (ν) (secondary wheel)

where m and n are the Wälzkurvenradien, μ and ν the assigned Polar angle, g mean the center distance. Tab. III contains some associated numerical values. Both rolling curves are twice-axisymmetric. Further technical data and details are listed below.

functionality kinematics

Fig. 16 shows the starting position at γ = 0 of the oval wheels 12, 13 , each with alternating contact in the points of extreme Wälzkurvenradien. The half angle difference δ here has its maximum, as can be seen in Fig. 11, so that both rotors 2, 3 are in extreme angular position to each other at the same angular velocity. For the segmental working chambers 29 this results in the dead center, with two chambers minimum, maximum with two chambers. Fig. 22 shows the associated assembly-appropriate position of the planetary gear set 34, 35 , which causes the forced operation. The radii of the rolling curves at the point of engagement are the same. The direction of rotation of the planet gears 34 and thus also the Ovalradmitten and the rotors is clockwise according to the Planetenradeigendung so also and also the rotation of the two oval gears 12th The rotation of the two oval gears 13 then runs counterclockwise. This results in an enlargement of the central gear chamber 10 and a reduction of the orbital gear chamber 11 from a middle position in Fig. 16. Fig. 17 shows the oval wheels in a position at γ = 15 °. Fig. 23 shows the associated position of the planet gears. Both positions correspond to the positions of Figs. 1, 4, 5. Fig. 18 shows the oval wheels in a position at γ = 45 °. Fig. 24 shows the corresponding position of the planet gears 34 with smallest radii of the planet gears 34 and largest radii of the central wheel 35 in working position. The inner rotor 2 has its smallest, the outer rotor 3 its largest angular velocity, the central gear chamber 10 reaches its maximum, the orbital gear chamber 11 reaches its minimum.

Fig. 19 shows the oval wheels in a position at γ = 90 °: maximum and minimum Wälzkurvenradien touch each other in the engagement points; second extreme position of both rotors 2, 3 to each other at the same angular velocity is reached. Reduction of the central gear chamber 10 and enlargement of the orbital gear chamber 11 .

Fig. 25 shows the associated position of the planet gears. The radii of the rolling curves at the point of engagement are the same. Fig. 20 shows the oval wheels in a position at γ = 135 °. Fig. 26 shows the associated position of the planet gears with largest radii of the planet gears 34 and smallest radii of the central wheel 35 in working position. The inner rotor 2 has its largest, the outer rotor 3 its smallest angular velocity, the central gear chamber 10 their minimum, the orbital gear chamber 11 their maximum. Fig. 21 shows the oval wheels in a position at γ = 180 °: Maximum and minimum Wälzkurvenradien touch. Fig. 27 shows the associated position of the planet gears. The radii of the rolling curves at the point of engagement are the same. After this half revolution of the rotors a full period ("double stroke") has passed through.

dynamics

The central drive shaft 38 with coupling via a planetary gear set 36, 37 ( FIG. 2) rotates in the same direction as both rotors 2, 3 , here clockwise, permanently at exactly twice the speed as the common tangents T 3 , T 4 of the oval wheels 12 13 , which are neutral lines at equal moments of inertia of the rotors 2, 3 . In addition, since the planetary masses rotate at the same speed as the tangents T 3 , T 4 in the same direction or opposite, no additional flywheel to compensate for torque fluctuations is required.

In the sub-variant with primary wheel 339 and secondary wheel 340 ( FIGS. 3, 4), the eccentric drive shaft 338 rotates counter to the rotors permanently exactly at the same speed as the tangents T 3 , T 4 , thereby providing a similarly favorable dynamic behavior as in the preferred variant Planetary gear 36, 37 is achieved.

power flow

When operating as an engine such as internal combustion engine, gas pressure motor, hydraulic motor gear units 10, 11 and working chambers 29 act together as follows: On toothed outer shell of each oval wheel are defined by counter-engagement each two uneven asymmetric faces, which cause torque of the oval wheels at pressure difference in the gear chambers 10, 11 , The effect of pressure differences in the working chambers 29 torques between two rotors are superimposed on the bearing pins 22, 21 the oval gears 12, 13 and transformed by their mutual support in non-radial normal force components in torques on the oval wheels. These normal force components must be collected here by the toothing, which is designed for this purpose in detail accordingly. When operating as a work machine such as a compressor or pump, the gear chambers 10, 11 and working chambers 29 , which periodically move in opposite directions and offset, generate suction and pressure in separate channels ( FIG. 3) in accordance with the above explanations.

The fluid flow control of the gear chambers 10, 11 via control channels 30 ( Figure 3) and in the rotor segments 28 recessed pockets; the segment-shaped working chambers 29 ( FIG. 5) are controlled via recesses 31 ( FIG. 5) in the connecting shafts 16 operating as rotary valves.

In a second embodiment, two axially successively arranged, parallel coupled, offset working rotor pairs 102, 103; 104, 105 ( FIG. 6) are controlled by two oval wheel sets 112, 113 ( FIG. 6) in a play-free design. For this purpose, the oval gears 112, 113 are each provided with a toothed belt 119 ( FIG. 6) and an axially offset smooth belt 120 ( FIG. 6) on the jacket. The toothed belt 119 serves primarily to transmit tangential forces, while the smooth belt 120 serves to receive and convert normal forces and torques. The outer contour of the smooth belt 120 is congruent to the pitch curve of the toothed belt 119th Since the smooth belt 120 of each oval wheel is supported in any operating position of each of the smooth belt surfaces of the two adjacent, associated with the respective counter rotor oval wheels in the contact line, permanent contact and thus backlash is guaranteed; Optionally, the oval wheels can also be operated with preload the smooth belt areas. In the critical operating points of the load direction reversal ( FIGS. 16, 19, 21), the oval wheels thus act like toggle levers in the dead center position, absorbing the longitudinal force components occurring between their bearing points in both rotors, caused by torques between the two rotors, without these components being critical Operating points as torques on the oval gears and thus as a load change in the teeth are effective. The shapes of the Ovalradwälzkurven and gears selected here are the same as those of the first embodiment. For storage of the wide-ranging oval gears 112, 113 , the rotors 102, 103, which are connected in a circular-disk-like manner on both end faces, are each provided with two bearing pins 121, 122 ( FIG. 6) mounted symmetrically parallel to the axis at a fixed distance f from the center. Their projection into the space is more than half the distance between the two rotor walls from each other and is matched as well as their diameter with the dimensions of the oval wheels 112, 113 concentric unilaterally mounted blind holes 123 ( Fig. 6) in the sense of a smooth low-friction sliding bearing. All bearing journals 121, 122 are provided with coaxial through holes 126 ( FIG. 6). Outside the centrifugal circle of the oval gears 112, 113 , an annular region is enclosed between the circular disk-shaped parts of the rotors widening here on the periphery, in which segment-shaped working chambers are formed in a known manner. In this outer region, the storage and axial guidance of the rotors 102, 103 in the housing 101 takes place . Mirrored to the transmission center plane, which is about the amount of oval wheel width left of the circular disk-shaped wall of the left rotor 103 , the second, mirror image executed rotor pair 104, 105 is stored, rotor 104 with the oval wheels 112 left outside, rotor 105 with the oval gears 113th towards the center of the gear.

The positive guidance of the oval wheels in this exemplary embodiment is carried out in sets, separated from the housing end sides, to the oval wheels 112 of the outer rotors 102, 104 in play-free design as follows: In the four bores 126 of the two outer rotors 102, 104 four connecting shafts 116 ( FIG ) rotatably mounted to the oval gears 112 coaxially. Outside the rotors, circular support plates 146 ( FIG. 6) are fastened to the connecting shafts 116 , on which eccentric, block-wise aligned, outwardly pointing eccentrics 144 ( FIG. 6) are mounted. These protrude each setwise for play-free entrainment in the symmetrically arranged receiving holes of an annular crank disk 147 ( Fig. 6, 8), which is eccentrically mounted in the housing end wall eccentrically rotatable about its center. The connecting line of both oval wheel centers, the connecting line of both eccentric centers and the two crank arms of the respective oval wheel center to the corresponding eccentric center form a parallelogram with the connecting line from rotor center to crank center center as center parallel. The smooth axial guidance of the crank disks 147 carried by the support plates 146 on the one hand and the housing end walls on the other.

For backlash-free, planetary coupling of both offset Ovalradsätze to a common drive shaft 138 ( Fig. 6) are for their storage in both end walls of the housing 101 ( Fig. 6) and in all four rotors 102, 103, 104, 105 coaxial, continuous bearing bores appropriate. The end faces of a central thickening of the continuous drive shaft 138 guide them axially between the two inner rotors 103, 105 . On this thickening, the drive shaft 138 has two same disc-shaped central eccentric 143 ( FIGS. 6, 7) which are twisted by 180 °. On each central eccentric 143 is ever an annular crank disk 148 ( Fig. 6, 7) in its central bore easily rotatable, free of play. Connected to the four oval gears 113 of the inner rotors 103, 105 , similarly as described above for the oval gears 112 of the outer rotors 102, 104 , are connecting shafts 116 ( FIG. 6) with carrier plate 146 and eccentric 144 . The latter each protrude in sets of play-free driving in symmetrically mounted in the crank disks 148 receiving bores. The sliding in the middle of the housing crank disks 148 are guided by the support plates 146 of the connecting shafts 16 from both sides axially together smoothly. In order to achieve a staggered working of the rotor pairs by 1/4 period, the bearing points of the two crank disks 147 in the housing 101 must be arranged rotated by 90 ° because of the similar eccentric positions relative to the associated oval wheels and the relative Zentralzzenterverdrehung of 180 ° ( Fig. 7, 8th).

Disturbing here is the asymmetric load distribution.  

In a first sub-variant of the second embodiment, a uniform load distribution with twice the number of crank pulleys, so a total of eight worked. Instead of the eight circular support plates 146 then eight disc-shaped eccentric 545 occur ( Fig. 9), whose center axes are positioned offset from the central axes of the pin-shaped eccentric 544 on the connecting shafts by 180 °, and which the added crank pulleys 549, 550 ( FIG. take over holes.

In a second sub-variant, the crankshafts arranged in the central region, deviating from the above preferred embodiment, are mounted peripherally in the eccentric bores of a hollow central eccentric 643 ( FIG. 9), which in turn is rotatably mounted rotatably in the housing 601 rotatably on its periphery so as to be free of play. A ring gear 642 coaxially mounted on its circumference meshes with a pinion 641 ( Figure 9) which is mounted on a drive shaft 638 ( Figure 9) rotatably mounted in the housing 601 eccentrically outside the rotor circuit.

functionality Kinematics, dynamics

The runaway configurations of the oval gear set formed from the oval gears 112, 113 are optically the same as for the first embodiment ( Figures 16-21) but with reversed rollers by opposite oval gear rotation directions as follows: The oval gears 112 positively constrained by crank disks remain in place during of the circulation constant direction, ie relative to one of the clockwise rotating with the angular velocity tangents T 3 , T 4 ( Fig. 17) rotate the oval gears 112 counterclockwise, in terms of magnitude at the same angular velocity. The internal rotations of the oval gears 113 are then inevitably directed relative to the circumferential tangents in the clockwise direction; The angular velocities of the oval gears 113 relative to the circumferential tangent amount also, so that their absolute intrinsic rotation = 2. This means at a half Tangents circulation full rotation of the oval gears 113 ( Fig. 16-21, mark X). The angular velocity of the drive shaft 138 is then also 2, the direction of rotation as rotors. In the second sub-variant this applies to the hollow central eccentric 143 ; the drive shaft 638 coupled via the constant ratio pinion 641 rotates faster and opposite. In the case of the second rotor pair 104, 105 , the same proportional ratios of the angular speeds of the drive shaft, tangent and planetary masses prevail, which are essential for smoothness and torque constancy.

power flow

When operating as an engine (combustion, compressed air, Hydraulic motor) are made by the between two rotors acting torque between the oval wheel bearing journal longitudinal forces effective, which, except in the dead center, through mutual support of the oval wheels on their smooth Belt areas transformed into torques of the oval wheels become. Forced operation via two connecting shafts with Carrier plate and eccentric and a crank disk is sufficient already, in order to achieve a rotation of both rotors. The Equidistant to each other running eccentric loose oval wheels are thus indirectly also forced, and thus also a crank disk mounted on both eccentrics, which then in turn is capable of a Zentralexzenter take. The dead center bridging takes place through the second rotor pair.

Conversely, when operating as a working machine (compressor, pump) driven by the central eccentric crankshaft takes over their eccentric bores with the eccentric loose oval wheels, which are thus offset together with the positively driven oval wheels just because of the positive guidance in rotation. By the mutual support of the oval gears on the smooth belt areas, the moments of the oval wheels as force components in the bearing pins 122, 121 are effective, which in turn build up a torque between the rotors, which increases very greatly towards the dead center.

Another distinctive and important sub - variant of the second embodiment is obtained by training the oval gear curves according to the frame formulas (e) and (ε) at a transmission characteristic of

δ = (1/4-δ₀ / π) · (24γ-6sin4γ-sin³4γ) / 6 + δ₀

valid for 0γπ / 2, defined in the remaining three quadrants by symmetry in γ = π / 2, γ = π, and with the initial value δ₀ = 0.92016 <0.95106. This function δ graphically illustrated in Fig. 12 is chosen here so that apart from doubled axis symmetry of the Ovalradwälzkurven ( Figure 14) a drive shaft rotational angle proportional flow behavior is effected without any ripple, so that with this arrangement of two parallel connected by 1/4 Period offset working controlled pairs of rotors 102, 103; 104, 105 a constant-pump / motor is formed. Some numerical values for ε and e are given in Tab. IV.

In a third exemplary embodiment, three rotors 206, 207, 208 ( FIG. 10) coupled mechanically in series one behind the other are provided by two clearance-free, offset oval gear sets 212, 213; 214, 215 ( FIG. 10), at whose two common connecting shafts 218 ( FIG. 10) both the backlash-free positive guidance via eccentric 245 and crank disks 249 and the planetary coupling of a central drive shaft 238 take place via a three-part planetary gear set. For this purpose, the oval gears 212, 213; 214, 215 each having a toothed belt 219 ( Fig. 10) and an axially offset, smooth belt 220 ( Fig. 10) provided on the jacket. Further details regarding the operation of both belts, the remarks in the second embodiment can be found here.

The structure of the gearbox is symmetrical. The three-part housing 201 ( FIG. 10) is formed from a central, perpendicular to the axis, annular core wall with a large central passage and two bell-shaped side parts. In the core wall, larger in diameter than the central opening, short, coaxial, cylindrical bearing bores are mounted from both sides, in which the middle, two-part rotor 207 , the inner edge of the core wall from the center embracing axially guided and rotatably mounted on its circumference is. On both sides outside the disc-shaped, the bearing serving parts of the central rotor operating within the smaller centrifugal oval wheels, here with a smooth belt to the transmission center, arranged axially guided by the end walls of the central rotor 207 and the outer rotors 206, 208th Their radial bearing, in turn, optionally takes place in the housing 201 or, as shown in FIG. 10, on the circumference in the central rotor 207 , whose two circular end walls are continued for this purpose at the periphery in axial extension as enclosing hollow cylinder beyond the Ovalradsätze. The axial guidance of the rotors 206, 208, not illustrated in FIG. 10 , by interengaging rotor segments 228 forming working chambers 229 is supplemented by axial guidance over the end walls of the housing 201 . In the middle rotor 207 227 (Fig. 10) are at a distance f from the rotor axis, by 180 ° offset from each other by two end walls continuous axially parallel bearing bores provided in which the same two left-sided oval wheels 215 co-axial with a respective one of two identical right-side oval wheels 213 rotatably mounted on each projecting Ovalradlagerzapfen 225 and connected via two connecting shafts 218 at the same angle twisted firmly connected and thus axially backlash-free and easily secured. For backlash-free, smooth-running pivotal mounting of the four loosely running, centrally drilled oval wheels 212, 214 , the two outer rotors 206, 208 are provided with two at a distance f from the rotor axis, offset by 180 °, cylindrical, axis-parallel bearing pin 222 provided, the supernatant something less than the oval wheel width. Between two disc-shaped end walls serving for storage, the middle rotor is recessed except for two connecting bridges 253 ( FIG. 10) described in greater detail below. In this space for synchronous positive guidance of the two Ovalradsätze on the connecting shafts near the end wall each two eccentric 245 twisted by 180 ° attached, such that each arranged in pairs in the same plane eccentric rotate equidistantly synchronously, each with play-free, smooth, common web guide by two symmetrical bearing bores of one of the two annular crank disks 249 . These are eccentric at its outer periphery in the core wall of the housing 201 , offset by 180 °, rotatably mounted. Two eccentric blind bores driven into the core wall within the rotor bearing diameter from both sides beyond the center form a passage in the center and a clearance in the interior of the core wall. In this middle space between the crank disks of the three-piece planetary gear set for coupling the central, continuous drive shaft 238 is housed; it consists of two identical non-circular planetary gears 236 and a non-circular central gear 237th The latter is coaxially mounted on the drive shaft 238 , which in turn is rotatably mounted in continuous, central bearing bores of the rotors 206, 207, 208 and the housing end walls. The two planet gears 236 meshing with the central gear 237 are fastened coaxially in the same direction on the connecting shafts 218 rotating at a constant distance from the rotor axis. On both sides provided on the planet gears 236 in the area of cam 245 start-up steps are used, in cooperation with the end faces of the oval wheel bearing pin 225 of the axial guide of the circular crank disks 249th These have, in addition to their central through- bore 251 ( FIG. 10), which permits contact-free circulation around the drive shaft 238 , and the two symmetrical bearing bores two through-bores 252 ( FIG. 10) offset at 90.degree peg-shaped connecting bridges 253 ( FIG. 10) fastened to both disc-shaped parts of the central rotor are passed, which permit a contact-free, circular, directionally constant relative movement of the crank disks 249 . The accommodation of the segment-shaped working chambers 229 formed in a known manner can be done in free spaces within the housing 201 by direct attachment to the rotors 206, 207, 208 .

functionality

The operation of a zero-clearance Ovalradsatzes has already been explained in the second embodiment. The direct mechanical coupling of the Ovalradsätze with each other by two torsionally rigid connecting shafts 218 for dead-center bridging reduces the number of movable components, but in this embodiment leads to more complex gear wheel geometries and special dynamic conditions: Neither of the rotors 206, 207, 208 nor one of the circumferential tangents (T 3 , T 4 ) of one or the other Ovalradsatzes are torque-neutral reference objects. With the same design of both Ovalradsätze 212, 213; 214, 215 , equal moments of inertia of the outer rotors 206, 208 and twice the moment of inertia of the central rotor 207 , the torque-neutral reference line runs approximately in the middle between two rotating with angular difference circumferential tangents of the two Ovalradsätze. Double axisymmetric oval gears are no longer optimal here, they give an asymmetrical movement; a delivery constant, as described in a sub-variant of the second embodiment, so you do not reach here. The transmission characteristic δ as a function of γ, which determines the oval gear curves together with the frame formulas (e) and (e), is adapted to the respective requirements. Once established, the pitch curves of the planet gears 236 and the central gear 237 are also determined according to formulas (a), (b) below. With proper coordination of the gear geometry with the moments of inertia, the backlash between the central gear 237 and planet gears 236 is not disturbing, since there no function-related periodic load fluctuations or changes occur.

This version is backlash-free, very compact and torsionally rigid; preferred embodiment for internal combustion engine control.  

Further applications of the transmission are limited to its use as a purely mechanical converter gearbox for converting rotational movements into angle-limited torsional vibrations. The commercial applications are varied and range from tool and machine tool Range up to the servo for rotary valves and butterfly valves. Depending on the intended use, the transmission characteristics δ individually adapted by appropriate shaping of Oval wheels.

The drive is in the simplest case on one of the oval wheels direct, but also planetary of one central drive shaft can be effected from, with low dynamic requirements the use of round gears enable.

The performed mathematical processing includes as a first step the Derivation of the frame formulas (e) and (ε) for determination the oval wheel geometries:

For the rolling of two rolling elements on each other Generally two conditions to consider:

• 1) When rolling, no sliding should occur.
• 2) The two rolling elements must not strive away from each other and do not penetrate each other.

For the rolling elements come the following design-related requirements added:

• 3a) The distances of all rolling element pivot points from one common, fixed, central point are constant.
• 3b) All distances rolling element pivot point - central point are equal to each other.
• 4) Each rolling element rolls simultaneously on two tangents adjacent rolling elements; the number of rolling elements is even.
• 5) The connecting lines of two opposing rolling element pivot points form each with the central point a constant angle of 180 °.

The number of rolling elements that make up a set is fixed here with four, which is not necessarily the case  must be, but requires the least amount of construction. The Conditions 1, 2, 3b, 4, 5 are based on logical considerations without computational effort by two reformulated, mathematically better detectable conditions replaced:

• 6) A right-angled axbox forms the common tangents T 3 , T 4 of all four rolling elements in the contact points, each with one rolling element per quadrant.
• 7) Two adjacent rolling elements are mirror images of each other, opposite rolling elements are the same; There are thus two identical pairs of rolling elements educated.

From 6) and 3a) follow mathematically exact formulated conditions, which, when used, is more generally accepted mathematical laws and rules of differential geometry lead to the frame formulas (e) and (ε).

From 6) follow in this context constraints, the the periodic behavior of the half-angle difference δ in Depending on the self-rotation γ concern and in the Determining this the transmission characteristic determining Function δ must be considered:

• 8) δ⟨γ + π / 2⟩ = π / 2-δ⟨γ⟩ (symmetry of the third kind).
• 9) δ⟨γ + π⟩ = δ⟨γ⟩ ⇒ the period length is π180 °, so that two mid-point and two-far areas Oval, which leads to the name "Ovalrund".
• 10) Continuity of the function δ and its first three derivatives after γ, so that kink-free course still at Angular acceleration is guaranteed.
• 11) The additional condition δ⟨-γ⟩ = δ⟨ + γ⟩ provides a function δ with a symmetry of the fourth kind, too two axis-symmetrical oval wheels leads.

The conditions 8, 9, 10, 11 allow a multiplicity of possible functions δ⟨γ⟩; such a function ( Figure 11) has been defined and shown graphically for the concrete shaping of the oval gears ( Figure 13) and the minor non-circular gears of the transmission ( Figures 15, 1, 4). The additional requirement of constant flow rate at constant drive shaft speed leads to a further specification of the symmetry properties:

• 12) δ'⟨γ⟩ + δ'⟨γ + π / 4⟩ = constant.

This additional requirement also leaves room for variations, one of which has been defined and graphically represented for the concrete shaping of the oval wheels ( FIG. 14) ( FIG. 12).

The optimal design and adaptation of the forced operation and the drive shaft coupling is for good dynamics important. For slowly rotating gearboxes, two are the same Planetary gear sets with round gears sufficient, as the Rotor masses in such training of forced operation and Drive shaft coupling in any case run unbalance. Imbalance arises in this case due to the non-circular, relatively small, planetary masses. usable are planetary gear sets formed with non-circular gears with circumferential ratios of central to planetary of 4: 2, 6: 2, 8: 2. , . also with internal toothing of the central wheel. The preferred embodiment with circumference ratios of 2: 2 offers in addition to full balance following Advantages: standardization (three equal wheels); - largest variation width for the oval wheels without extreme geometric degeneration of the planet gears; - Combinability with crank pulley forcing / drive.

Also, the wheelset 339, 340 for coupling an off-center drive shaft to a rotor can be varied. Possible circumferential ratios of both wheels are dependent on the Planetenradübersetzung and also executable with internal toothing. The preferred embodiment with a circumference ratio of 2: 2 allows not only full balance balanced proportions of non-circular gears without strong degeneracy with a large range of variation for the oval wheels. The calculation of these wheelsets of non-circular gears is carried out with the help of the known, mathematically derived frame formulas for rolling curves with constant center distance c:

The constructional relationships between oval wheel rotation γ and respective angles of rotation α and β of the respective Wheels are in the preferred embodiments at a constant speed of the planet. masses:

Planetary Coupling / Forced Operation:
Central wheel: α = γ - (δ₀-δ) (Zr)
Planetary gear: β = γ + (δ₀-δ) (Pl)
Coupling of the eccentric drive shaft to the loose [guided] rotor:
Primary wheel: α = γ + [-] (δ₀-δ) (Pri)
Secondary wheel: β = γ (Sec)

The differentiation according to γ is then given in each case and, which, inserted into the associated formulas (a) and (b), give the angles α and β the associated radii. In the course of a realistic implementation, the fine tuning of the max. Half-angle difference δ₀ mathematically made so that are used at a predetermined distance f for the oval gears and the wheels of the planetary gear set with c = f standard moduli. In the wheel set of primary wheel 339 and secondary wheel 340 , conversely, the fine tuning of the axial distance (c = g) was carried out after the module and the number of teeth had been determined.

ε ° e (mm) 0 38.638 2,787 38.602 5,587 38.492 8,414 38.410 11.296 38.052 14.271 37.714 17.388 37.292 20.712 36.777 24.315 36.161 28.280 35.440 32.688 34.612 37.626 33.691 43.170 32.699 49.386 31.680 56.312 30.692 63.943 29.805 72.212 29.097 80.971 28.639 90 28.480

f = 48 (mm)
δ₀ = 0.9356 53.6555 ° <0.96031 55.0216 °
z = 45 (number of teeth)
m = 1.5 (mm)

Table II

f = 48 (mm); Oval wheel acc. Tab. I
δ₀ = 0.9356 (rad.) 53.6555 °
z = 50 (number of teeth)
m = 1.0 (mm)

Table III

g = 50.3556 (mm)
δ₀ = 0.9356 (rad)
zp = zs = 34 (number of teeth)
m = 1.50 (mm)

Table IV

f = 48 (mm)
δ₀ = 0.92016 52.72128 ° <0.95106 54.4918 °
Curve length = 67.5 π (mm)

Claims (23)

1. control and converter transmissions, in particular for power and working machines, with at least two coaxially rotatably mounted in a housing, working chambers forming, non-uniform rotating rotors and with the oscillatory rotations of the rotors controlling non-circular gears, characterized
in that each rotor has at least two equal, point-symmetrical, in-plane and eccentric, parallel and symmetrical, rotatably mounted oval gears with an externally toothed belt on the shell, each oval wheel ( 12, 112, 412, 212 ) of the first rotor ( 2, 102, 208 ) with two mirror-symmetrically designed oval wheels ( 13, 313, 113, 413, 213 ) of a second rotor ( 3, 303, 103, 207 ) in the sense of a speed-like, alternately counter-rotating rotation, with varying center distances, meshes in such a way that this pair of oval gears ( 12, 13, 112, 113, 412, 212, 213 ) formed with two identical pairs of web-like and rotor-like synchronously running oval gears, the rotational angle difference between two rotors ( 2, 3; 2, 303; 102, 103; 207 ) controls;
that the geometric shape of these point-symmetric Ovalradwälzkurven the polar equations is sufficient, where e is the radius at the point of engagement, ε the e associated polar angle, γ the Ovalradeigenddrehung with respect to the tangent (T 3 ) at the point of engagement, δ half the rotation angle difference of two rotors as a function of γ, δ 'the differential quotient dδ / dγ, f the fixed Mean the distance between the oval wheel axis and the rotor axis; in that at least one oval wheel is forcibly guided in a planetary direction directly or via a connecting shaft by means of further rotating components relative to the housing ( 1, 301; 101, 601; 201 ), so that the transmission has a guided rotor ( 2; 102; 207 ) and at least one further has loose rotor ( 3, 103, 206, 208 ) entrained in the oval gears, and in that the external mechanical coupling on one of the rotors takes place directly or via further rotating components and a drive shaft ( 38, 338, 138, 638, 238 ) , 2. control and converter transmission according to claim 1, characterized
that the Ovalradwälzkurven are formed symmetrically in addition to their symmetry to two mutually perpendicular in their origin symmetry axes, so that four congruent Wälzkurven are formed;
that the number of teeth of the oval gears is odd and that the teeth are arranged symmetrically to one of the two axes of symmetry, so that a combing according to claim 1 oval gear is formed with four identical oval wheels. 3. control and converter transmission according to claim 1, characterized
in that the two rotors ( 2, 3, 2, 303 ) controlled by the set of oval wheels enclose this pot-like manner with an annularly dynamic seal at its butt joint, such that the circular end walls of the cup-shaped gear chamber ( 9 ) thus formed on the inner sides each have two bearing journals ( 21st , 22 ) for the dynamically tight bearing of the unilaterally provided with central blind holes ( 23 ) oval gears and coaxially through the bearing pins ( 21, 22 ) extending bores ( 26 ) for passing the connecting shafts ( 16, 17 ) of the not loosely revolving oval wheels ( 12, 13 ) respectively;
that the oval gears on the one hand wide and plane-parallel in the sense of a two-sided axial, low-play, sealing sliding through the inside plane-parallel end walls of the gear compartment and on the other hand made with an over the entire Ovalradbreite made externally toothed belt with labyrinth seal, so that within the gear compartment ( 9 ) has two variable counteracting, coaxial working chambers, the central gear chamber ( 10 ) and the orbital gear chamber ( 11 ) are formed. 4. control and converter transmission according to claim 1, characterized
in that more than two rotors ( 206-208 ) are mechanically coupled in series one behind the other in such a way that the rotors are stacked alternately axially with oval wheel sets ( 212, 213/214, 215 ) as coupling members, each with a rotor ( 206, 208 ) concluding on both sides;
that both left-side and both right-side oval gears ( 215, 213 ) of each inner rotor ( 207 ) are mounted coaxially with each other in pairs and are identical in rotation by one connecting shaft ( 218 );
that either at an arbitrary number of inner rotors both left-hand oval wheels opposite to two right-hand oval wheels are fixed identically twisted, so that the Ovalradsätze work out of phase;
that the Ovalradsätze ( 212, 213/214, 215 ) are either the same, symmetrical or different. 5. control and converter transmission according to claim 1, characterized
in that the four oval wheels of an oval gear set are each divided into two axially offset belts, a belt ( 119, 419; 219 ) having a pitch curve shape corresponding to the frame formulas in claim 1, in particular for transmitting tangential forces, and an untoothed, smooth belt ( 120, 420, 220 ) with a rolling curve of the respective toothing congruent outer contour for receiving and converting normal forces and torques, wherein the toothed and the untoothed Ovalradteil, either directly or at an axial distance via an intermediate piece are firmly connected. 6. control and converter transmission according to claim 1, characterized
in that each two oval wheels ( 112, 412; 213 ) of the guided rotor ( 102, 207 ) are mounted on an eccentric ( 144, 544, 545; 245 ) aligned directly or parallel to connecting shafts ( 116, 218 ) by an axially parallel crank disk ( 147; 547 , 548; 249 ) are coupled with two symmetrically mounted bores for driving the two eccentrics, such that the connecting line of both Ovalradmitten, the line connecting both eccentric centers and the two crank arms of each Ovalradmitte to the eccentric center form a parallelogram with the line connecting the rotor center to Crankshaft center as center parallel and
the crank disk is rotatably mounted about its center in the housing ( 101, 601; 201 ), so that both oval wheels are positively guided in a planetary manner. 7. control and converter transmission according to claim 1, characterized
in that the two oval wheels ( 113 ) of the loose rotor ( 103 ) are coupled to one another via directly or on connecting shafts ( 17 ) aligned eccentric ( 144 ) by an axially parallel crank disc ( 148 ) with two symmetrically mounted holes for driving the two eccentrics ( 144 ) are;
that the crank disc ( 148 ) rotatable about its center, by a rotorkoaxial in the housing ( 101 ) rotatably mounted central eccentric ( 143 ) is supported and taken, such that the connecting line of the three eccentric pivots, the line connecting the three eccentric centers and the two crank arms of respective oval wheel center to the corresponding eccentric center form a parallelogram with the connecting line of Zentralezzenterdrehpunkt central eccentric center as a medium-parallel crank arm. 8. control and converter transmission according to claim 1, alternatively to claim 6, characterized
that the oval wheels ( 12 ) of the guided rotor ( 2 ) via Ver connecting shafts ( 16 ), two mounted on the connecting shafts ( 16 ), the same planetary gears ( 34 ) and a rotorkoaxial stored, with two planetary gears ( 34 ) meshing, first central wheel ( 35 ) are centrally controllable, with optional momentum balance of the planetary mass moved by appropriate training of the planet gears ( 34 ) and the central wheel ( 35 ) as a non-round;
in that the first central gear ( 35 ) is fixed in a rotationally fixed manner to the counter-torque support on the housing ( 1, 301 ), so that a positive guidance of both rotors ( 2, 3 ) is formed in the sense of alternately increasing and decreasing rotational speeds. 9. control and converter transmission according to claim 1, alternatively to claim 7, characterized
that the two oval wheels ( 13 ) of the loose rotor ( 3 ) via connecting shafts ( 17 ), two on the connecting shafts ( 17 ) fixed, same planetary gears ( 36 ) and a centrally mounted, with two planetary gears ( 36 ) meshing second central wheel ( 37 ) are coupled with a rotorkoaxial in the housing ( 1 ) mounted drive shaft ( 38 ), in optional momentum balance of the planetary mass moved by appropriate design of the planetary gears ( 36 ) and the central wheel ( 37 ) as a round. 10. control and converter gear according to claim 8 and 9, characterized
that the circumferential ratio of central wheel ( 35, 37 ) to Plane tenzahnrad ( 34, 36 ) is 2: 2, that is, a period corresponding to 180 °;
that each planetary gear set of three congruent non-circular gears ( 34, 35, 36, 37 ) with the polar equations l = f · (1 + δ ') / 2 (l) and
λ = γ + δ₀-δ (λ), where l is the rolling curve radius, λ the l associated polar angle, δ₀ half the initial rotation angle difference of two rotors at γ = 0 and f, γ, δ, δ 'correspond to the definition of claim 1, so that at constant drive shaft speed, the internal rotations of the planetary moving masses are constant. 11. control and converter gear according to claim 7 and 8, characterized
that both the forced operation and the external coupling is carried out deviating from claim 7 at the oval wheels of the guided rotor, so that the positive guidance with non-circular gears according to claim 8/10, the external coupling is effected according to claim 7 via eccentric, crank and Zentralzcenter run along the two oval wheels of the loose rotor as loose wheels. 12. control and converter gear according to claim 6 and 9, characterized
that both the forced operation and the external coupling deviates from claim 9 on the oval wheels ( 213, 215 ) of the guided rotor ( 207 ), such that the forced operation according to claim 6 via eccentric ( 245 ) and crank disc ( 249 ) takes place external coupling according to claim 9/10 with non-circular gears ( 236, 237 ) is effected, the two oval wheels ( 212, 214 ) of the respective loose rotor ( 206, 208 ) run as idler wheels. 13. Control and converter gearbox according to claim 1 or 6 or 8 as an alternative to 7 and 9, characterized
that the external coupling of the transmission to a non-coaxial, axially parallel outside the rotors ( 2, 303 ) in the housing ( 301 ) mounted drive shaft ( 338 ) via a two-piece gear set with, if necessary, non-circular gears, such that the one gear, as the primary wheel ( 339 ), with one of the two rotors ( 2, 303 ) coaxial rotatably connected, the other gear ( 340 ), referred to as a secondary, rotatably coaxially mounted on the drive shaft ( 338 ). 14. Control and converter gear according to claim 13, 10 and 1, characterized
that primary wheel ( 339 ) and secondary wheel ( 340 ) in the Umfangsver ratio 2: 2 according to the polar equations m = g / (2-δ ') (m) μ = γ + δ₀-δ (μ) n = g · (1-δ ') / (2-δ') (n) ν = γ (ν) are formed, where m and n are the Wälzkurvenradien, μ and ν the associated polar angle, g the center distance and δ₀, γ, δ, δ 'the definitions Of claim 10 and 1, so that thereby optionally combined use with a positive guidance of the rotors according to claim 6 or 10, the internal rotations of the planetary masses are constant, the torque fluctuation of the drive shaft at the same rotor moments of inertia is zero. 15. Control and converter transmission according to claim 2, characterized
that the dependence of the half rotational angle difference δ of the Ovalradeigenddrehung γ by the formula δ = π / 4 + (δ₀-π / 4) cos (2 γ) is given, where δ₀ half the initial rotational angle difference at γ = 0, freely selectable within π / 4 <δ₀ <0.96031 and π = circular number = 3.1415. , , mean. 16. Control and converter transmission according to claim 2, alternatively to claim 15, characterized in
the dependence of the half rotational angle difference δ on the oval-wave rotation γ is given by the formula δ = (1/4-δ₀ / π) · (24 γ-6 sin 4 γ-sin³ 4 γ) 6 + δ₀ in the range 0γπ / 2, where the symbols of the definition of claim 1 and 15 correspond, the shape of the curve in the remaining three quadrants from the symmetry in γ = π / 2 and γ = π follows, δ₀ half the initial rotation angle difference, freely selectable π / 4 <δ₀ <0 , 95106 and π = number of circles = 3.1415. , , mean. 17. Control and converter transmission according to claim 2 and 3, characterized
that the four working chambers ( 29 ) formed in a manner known per se by the segments ( 28 ) of the two rotors ( 2, 3; 2, 303 ) are operated in combination with the two phase-offset and counter-rotating gear chambers ( 10, 11 ), such that a dead-center fluid converter transmission is formed with only one stage;
that the fluid control of the gear chambers ( 10, 11 ) via channels ( 30 ) and recesses in the rotor segments ( 28 ) is effected, the fluid control of the rotor working chambers ( 29 ) via the acting as a rotary valve, provided with recesses connecting shafts ( 16 ) of the oval wheels, so that a quick reversing takes place at the respective dead centers without the use of additional components. 18. Control and converter gear according to claim 3 and 4, characterized
that according to claim 4 more than two mechanically successively coupled rotors the oval wheel sets located between them pot-like enclose and thus each oval wheel set form two variable coaxial Geriebekammern which twisted by according to claim 4 set twisted on the connecting shafts attached oval gears in pairs out of phase with each other, so that is formed by parallel connection of the phase-controlled at least four gear chambers, a dead-point converter transmission. 19. Control and converter transmission according to claim 4 and 5, characterized
that according to claim 4 more than two mechanically in series coupled rotors by setwise twisted on the connecting shafts mounted oval wheels are controlled in phase with two each according to claim 5 differently shaped belt areas, so that a dead-point transmission is formed. 20. Control and converter gear according to claim 2, 5, 6, 7 and 15, characterized
that in the housing ( 101 ) four rotors ( 102, 103, 104, 105 ) are rotatably mounted coaxially, such that each two rotors, as described in claim 1, are controlled by four oval wheels;
that the eight oval wheels according to claim 5 are each divided into two axially offset belt areas and are formed according to the formula in claim 15;
in that the four oval gears ( 112, 113 ) of the first rotor pair ( 102, 103 ) are offset by 1/4 period from the four oval gears ( 112, 113 ) of the second rotor pair ( 104, 105 ) for dead-center bridging, the gear parts for positive guidance and external coupling according to claim 6 and 7 are formed without play in duplicate with the exception of the common drive shaft ( 138 ) to which the two central eccentric ( 143 ) are mounted offset by 180 °. 21. Control and converter transmission according to claim 2, 5, 6, 7 and 16, characterized
that the transmission corresponds in construction to the embodiment according to claim 20;
that the Wälzkurvenformen of the oval wheels, however - deviating from claim 20 and 15 - are designed in the sense of a flow constant according to claim 16. 22. Control and converter gear according to claim 12 and 19, characterized
in that three rotors ( 206, 207, 208 ) of different moments of inertia mechanically coupled in series in series are controlled by two oval gear sets ( 212, 213/214, 215 ) whose positive guidance and shaft coupling take place according to claim 12. 23. Control and converter gearboxes with two coaxially rotatably mounted in a housing rotors and with the torsional vibrations of the rotors controlling non-circular gears, characterized
in that each rotor has two equal, point symmetrical, in the same plane and to the rotor axis eccentric, parallel and symmetrically arranged, rotatably mounted oval wheels with an externally toothed belt on the shell, each oval wheel of the first rotor with two mirror-symmetrically designed oval wheels of a second rotor in the Meaning of a speed-like, alternately opposite rotational vibration, with varying center distances, meshes, so that this two pairs of identical webs and rotorwise synchronously running oval gears formed Ovalradsatz controls the rotational angle difference of the two rotors;
that the geometric shape of these point-symmetrical Oval wheelwheel curves satisfies the polar equations (e) and (ε) of claim 1;
in that at least one oval wheel is driven via a connecting shaft and optionally further gear parts.