Eccentric gear and process for making such a gear

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PATENT INVENTOR INVENTOR'S OFFICE NAME

Gustav Rennerfelt, Nilstorpsvägen Gustav Rennerfelt, Lidingö SE Lars Johansson Patentbyrå AB Eccentric gear and procedure for switching

181 47 Lidingö SE production of such (56) (57)

CALLED PUBLICATIONS:

SUMMARY:

An eccentric gear comprises first and second gear rings (7, 10), one of which (7) has internal, straight teeth and the other (10) a smaller number of external teeth.

At least one gear ring exhibit conical shape with a large end (AA) and a small end (BB), the second gear ring (10) having a smaller pitch diameter than the first (7) and being eccentrically stored relative to it. In addition, axes of symmetry (24,25) for the respective gear rings form an angle (and) with each other. The contact points or locations for engagement between flanks on cogs in one of the cogs and opposite flanks on the cogs of this second cog ring are mainly located along a helical line extending in axial direction between said large and small ends.

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The numbers in brackets indicate international identification code, INID code. Letters in clamps indicate international document code.

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Technical field of the invention

This invention relates to an eccentric gear of the type comprising first and second co-operating gear rings, one of which exhibits a certain number of internal, straight teeth and the other a fewer number of external, equally straight teeth, and at least one of which exhibits a conical shape with a a large end and a small end, the second gear ring having a smaller pitch diameter than the first one and being eccentrically stored relative to it to roll off on its internal teeth; each axis of symmetry for each gear ring forms an angle with each other.

An example of a field of application for an eccentric gear of the type described is tear rollers in printing machines, in particular tire rollers with a built-in mechanism for effecting an axial, reciprocating movement of the roller while rotating at high speed. Another example of a field of application is reduction gears where quiet walking is important. In medical applications, control of curtains, awnings, household appliances and the like, planetary gears of plastic material are often used. These have a high noise level, which is an inconvenience.

Prior art

WO-A-9306999 discloses a mechanism in which a slightly tapered gear ring rolls toward a cylindrical internal gear ring. The slightly tapered gear ring, which is eccentrically stored, has one or more teeth less than the cylindrical gear ring and thus a slow change in angular position between said gear rings. This change in angular position is transmitted to a cam element and a cam roller thereby providing an axial displacement of a ridge roller in which the mechanism is included.

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Furthermore, US Patent 5,030,184 discloses a tooth correction method which makes it possible to produce an eccentric gear at which the two gear rings have a tooth number that differs by one or a few teeth.

In the structural design of the cogs of the gear according to the WO-A-9306999, a cross-section is laid through the slightly tapered, eccentrically arranged gear ring in a position about the middle of the axial length of the gear ring. In this position, by means of the computational and correction method described in said US Patent 5,030,184, the tapered external gear ring is corrected so that its contact with the cylindrical internal gear ring is optimal. Thus, the resulting tooth alignment is optimal only at a cut of half the axial length of the tapered, eccentric gear ring. In the unloaded state of the gear, in this axial section, theoretically, only two gear pairs are engaged, ie one pair in each load direction. Due to gear deflection, several gear pairs engage with each other at the gear load. Normally 3-5 tooth pairs in each load direction will be loaded. In each axial cut, which is different from the optimum middle section on the conical gear ring, a tooth bearing that differs from that in this cut is optimally possible.

The objects and features of the invention

In order for optimum load transfer to be obtained in the said pair of cogs which are in engagement with each other in the unloaded state of the gear, it is desirable that the cog alignment in each axial cross section is optimal. This is accomplished by turning the blank to the slightly conical gear ring with a certain peak angle determined by the method described below. From this blank, the conical gear ring is produced with a certain gear angle determined by the method described below.

According to the invention there is obtained an eccentric gear whose straight teeth in the conical gear ring have a helical contact line against the internal, straight teeth of the cylindrical gear ring. This helical contact line gives the gear one

501 463 soft and quiet operation and contributes to the gear transmitting higher torque. Since the load is distributed over many teeth and along the entire axial length of the gear ring, the surface pressure becomes relatively low, which facilitates the possibility of maintaining a good lubricating film and thus a good wear resistance of the gear.

Figure Description

An embodiment of the invention will be described in more detail below with reference to the accompanying drawings, in which: Fig. 1 is a longitudinal section of a known eccentric gear assembly in a tear roll; Fig. 2 is a partial cross-sectional view showing the tooth engagement between a toothed ring with internal teeth and a toothed ring with external teeth before profile offset and stub; Figure 3 is an analogous section after profile offset Fig. 4 is an enlarged and simplified longitudinal section showing angles, stubble of the externally toothed wreath and stumping of the internally toothed wreath. Figure 5 is a cross section A - A of Figure 4, Figure 6 is a cross section B - B of Figure 5, Figure 7 is a longitudinal section through a reduction gear where an eccentric gear is made of plastic to obtain optimum quiet operation, and Fig. 8 is a longitudinal section of a reduction gear according to Fig. 7, where also the internal gear ring is conical to allow for elimination of gaps.

Further description of the prior art

Figure 1 shows an eccentric gear assembly intended to be inserted into a tear roller for a printing press to assign a reciprocating motion to it. The assembly is of the type described in said WO-A-9306999. The assembly is inserted and fixed at one end by a groove roller, which includes a rubber-coated outer tube. The shaft 1 of the unit is attached

501 463 at the printing press stand. The unit thus constitutes the one bearing for the tear roll. The roller is driven at high speed by rolling it off against another roller in the printing press. The task of the unit is, in addition to being a storage unit, to give the ridge roller an axially reciprocating, slow movement. Usually, the motion is sinusoidal with an amplitude in the range of +/- 5-20 mm. The speed of the Riwals can be 1,000 - 2,000 rpm and the frequency of axial movement to 0.5 - 2 H_z.

To obtain this sinusoidal axial movement, the assembly is constructed as follows:

On a stationary shaft 1, a cylinder 2 is rotatably mounted by gables 3, 4 and needle bearings 5, 6. A cam ball element 8 is obliquely mounted on the stationary shaft 1 and has a conical gear ring 10 with external teeth that rolls off against an internal cylindrical gear ring. 7 on the cylinder 2. The other end of the cam ball element is provided with an axially directed cam groove 9 in which a cam roller 14 rolls off. The cam roller is mounted on a shaft 15 attached to the cylinder 2.

The conical gear ring 10 and the cylindrical gear ring 7 are parts of an eccentric gear. The cogs in these wreaths have been corrected according to the method described above. Thus, in the construction of the cogs of the known gear, a tooth cross-section located approximately in the middle of the axial length of the conical gear 10 is selected. The cross-section of the conical gear 10 in this mid-axial length position provides optimal abutment relative to the internal teeth of the cylindrical gear ring 7 by a calculation and correction method according to Figures 2 and 3.

As the input data for the tooth correction method, the desired gear ratio and the approximate partition diameter Ds as well as the desired value m of the tooth module are used. From these values, a first approximate value of the pitch diameter of the gear ring 10A is calculated, and a first value

501 463 on the eccentricity e_O, which is approximately equal to half the pitch diameter Ds of the gear ring 10 A less than half the pitch diameter of the gear ring 7 A. Furthermore, the gear number Z is calculated_10A and Z_7A for the two cogs 10A and 7 A. If the cog number does not become an integer, the pitch diameter is changed and the procedure is repeated. When the number of cogs becomes an integer, preferably in CAD technology, the cogs of the two cogs are drawn. The result is shown in Figure 2. Figure 2 shows the gear engagement between a gear ring 10A with 89 external teeth and a gear ring 7A with 90 internal teeth. The cogs on the cogs 7A and 10A are made by a conventional cog processing method. As can be seen in Figure 2, there is gear interference between 06:00 and 09:00. In these areas the flanks of the teeth collide. A gear made with such gear rings would be blocked.

To avoid such tooth interference profile, the teeth of the gear rings 7A and 10A are displaced, whereby a new value of the eccentricity is obtained according to the following equation: e₁= e_Q+ f · (m * x₁ , mx₂) where m is the gear module, x₁ and x₂ profile offset coefficients and where f (....) is a function of mx ^ and m * x₂.

The cogs of the two cogwheels are redrawn again with the cogs corrected by profile offset. By this adaptation of the eccentricity to the profile displacement, an optimal abutment is obtained between the flanks of the gear pairs which (in unloaded condition) are in contact with each other. In addition, a certain correction of either of the press angles of the gear rings may be relevant to obtain the desired optimum attenuation between the gear edges. Then, one or both of the gear rings are stumped so that the teeth can pass each other as they go out of engagement. The size of the stubble should be such that the manufacturing tolerances cannot cause the tooth peaks to collide. The stubble causes the abutment length in the radial joint to decrease and thus the surface pressure in the gear grip is increased. A compromise must therefore be made between the size of the stubble and the length of the sting. The result

501 463 is shown in Figure 3. Z_r denotes the number of teeth (89) on the eccentric gear ring 10 B, Z_s the number of teeth (90) on the gear ring 7B, &lt;|)_r the diameter of the circle gear peaks on the eccentric gear 10B describes and φ₈ the diameter of the curved tooth peaks on the gear ring 7B describes. The profile offset does not change the pitch circle designated by 30B for the gear ring 10B, but means that during the manufacture the tool is displaced towards the center of the blank for the gear ring 10B by the profile offset value. The size of the profile offset, expressed in millimeters, is -xm, where the minus sign indicates that the tool is offset towards the center of the blank. Stubble means that the height of the tooth peaks is reduced by changing the diameter of the tooth ring. It can be seen from Figure 3 that the dividing circles 30A, 30B for the two gear rings intersect at approximately 10.30 and in an area between about 10.30 and 01.30 the dividing circle 30B lies outside the dividing circle 30A.

Detailed description of the invention

Figure 4 shows schematically a longitudinal section of the gear rings 7, 10 according to figure 1, the angles being slightly exaggerated for the sake of clarity. The axis of symmetry of the eccentric gear ring 25 forms an angle and with the axis of symmetry of the cylindrical internal gear 24. The axial section A - A of the tapered gear ring at its major end is in engagement with the dividing diameter 26 of the cylindrical internal gear.

An optimum tooth alignment is performed in this axial section A - A by the method described above and in US Patent 5,030,184. In this way, a first value of the eccentricity is obtained._A. For the axial section B - B, then the geometric eccentricity e_B determined to:

e_B = e_A + b · tan «where oc is the desired cone angle and b is the tooth width.

The optimal ball solution for the axial section B - B is now developed as was done in cross section A - A. However, in the section B - B, the given fixed eccentricity is now assumed

501 463 e_B (and any pressure angle correction) and make profile shifts and stubs on the gear rings to obtain optimal abutment.

Figures 5 and 6 show a gear solution where the gear ratio i = 10,167, the module m = 0.7 and the difference in gear number is = 67 61 = 6. Active gear width (= the distance between the axial cross sections A - A and B - B) is 8 mm.

At the optimum solution of the large-end (axial cross-section AA), it is found for unloaded gear that the two pairs of gears that are in engagement with each other form the angle y_A = 10.75 ° towards the eccentric direction. The eccentricity is here e_A = 2,305 mm.

At the optimum solution of the small end (axial cross-section BB), it is found for unloaded gear that the two tooth pairs that engage each other form the angle = 16.16 ° towards the eccentric direction. The eccentricity is here e_B = 2.51 mm.

From these two Figures 5 and 6, it can be found that in an axial cross section that moves from section A - A to section B - B, the angular positions of the contact points relative to the eccentric direction will be continuously changed from 10.75 ° to 16.16 °. It can also be noted that the size of the angular change in this example corresponds approximately to a pitch (5.37 °). For larger gears, ie for smaller differences in gear ratios, this change in angle will be greater. Thus, with this tooth optimization method, a helical tooth bearing has been obtained, which is very favorable for the tooth strength and gives the gear a quiet go. This helical contact line can be likened to the helical contact line obtained at an oblique cut tooth. A straight conical tooth, with the described method, obtains a helical contact line which provides a smooth operation and permits the transmission of high torque. Thus, this tooth is not to be made with the cone angle 2oc (and is the angle that the axis of rotation of the tooth should have relative to the axis of rotation of the internal tooth ring) but must be made with the cone angle 2β obtained by the expression:

501 463 β = arctane (g / b) where g is the difference in profile offset (-x_B&gt; m) - (-x_A-m) in the cross sections A - A and B - B and b is the perpendicular distance between said cross sections. The blank of the conical gear ring must be turned conically with the angle 2v obtained by the expression:

v = arctane ((S_B-S_A) / b) where S_A and S_B is the stub in the respective axial sections AA and BB.

In summary, in order to obtain optimum gear positioning in the gear, the tapered gear ring, whose axis of rotation forms the angle cc with the axis of rotation of the cylindrical internal gear, shall be manufactured with the gear angle 2β and the top angle 2v.

The described embodiment of the invention relates to a tapered eccentric gear for a rivulet assembly where the eccentric shaft is stationary and the internal gear ring rotates.

As mentioned earlier, the invention can also be used in applications where the internal gear ring is stationary and the eccentric gear ring is freely stored on the input, rapidly rotating shaft. With a suitable material combination, e.g. In this embodiment, a reduction in speed can be obtained which is characterized by a very soft and silent operation due to the screw abutment.

Figure 7 shows a longitudinal section of an eccentric gear intended for a medical application where quiet walking is a primary requirement. On a fast-rotating input shaft 40, one end of a conical gear 42 is freely eccentrically stored by a ball bearing 41. The gear 42 rolls off on a stationary internal gear 42. The other end of the gear 42 is via an angular engagement coupling (arc tooth coupling) with the output shaft of the gear 45. In this application is geared

501 463 tion 50: 1. Note that the output shaft has reduced reverse rotation towards the input shaft.

In order to eliminate such gaps in the gearing caused by manufacturing tolerances or wear, the internal gear ring 56 (see Fig. 8) may be conical and abutment optimized in the previously described manner against the eccentric gear 52, the gear gear 57 of which may be cylindrical or tapered in the opposite direction, i.e. is directed to the left in Figure 8. By means of a washer spring 58 acting between the input shaft 50 and the ball bearing 51 an automatic elimination of gap is obtained. The disadvantage of this gear design is that the internal gear is relatively expensive to manufacture. In series production, however, it is possible to produce such gears, e.g. by powder sintering technique or plastic casting.

In the described application as gear in a tear roller assembly, the internal cogs are also used as splines, because in the reciprocating motion, the eccentric cog will move axially along the internal cog, which then serves as the splines. In other applications such as reduction gear only, where no axial movement between the gears does not occur, the internal gear ring may be conical and the outer eccentric gear ring may be cylindrical. Alternatively, both gear rings may be tapered.

The eccentric gear described can also be used backwards, ie to increase the speed.

The gear type can also be used in differential design, i.e., driving both said input shaft and either the internal gear ring or the formerly called outgoing shaft in such a way that outgoing rpm on a third shaft becomes the sum of the input rpm of the two other shafts.

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