CA1325119C - Translation ring element for an eccentric gear - Google Patents

Abstract

ABSTRACT OF THE DISCLOSURE
A new movement translation element, a so-called transmitter or driving dog, translates the rotation of the centre of gravity of an eccentrically mounted eccentric gear wheel to a first shaft while the eccentric gear wheel simultaneously executes a planetary movement about a second shaft and is characterized by a rigid body mounted for pivoting about first pins which are disposed substantially at right angles to the first shaft. The rigid body has in its symmetry plane for the first pins and first shaft axially directed, tooth-like projections which fit with negligible play into two radially counter-directed apertures in the eccentric gear wheel which allows the projections to execute radial sliding reciprocatory motion in the apertures simultaneously as they execute a rolling movement therein, and the rigid body pivots round the first pins.

Description

132~1~9 ~is application is a division of A~plication Serial No. 611,211 filed Sep~3~er 13, 1989.

The present invention relates to an excentric gear o' the kind which allows high reduction, while the volume of the gear is s~all. A gear of this ~ind has an excentric spur gear, also known as a satellite wheel or planet wheel, which is freely journalled on an excentric stub shaft rotating together with the input shaft of the gear. The excentric gear wheel is in mesh with a stationary, internally toothed gear wheel and rolls round on the internal surface thereof. The excentric gear wheel has a circumference which is somewhat less than that of the stationary gear wheel, and as it rolls round on the inside of the stationary gear wheel it executes a slow rotation about its journal on the excentric stub shaft simultaneously as its centre of gravity rotates at a great rotational rate about the input shaft.

The excentric wheel has a number of teeth which is less than that of the stationary gear wheel- For one revolution of the input shaft the excentric wheel has rotated in the opposite direction an angle responsive to the relationship between the number of teeth on the respective gear wheel.

Gear ratio (reduction) i = _ z r z Zs = number of teeth on stationary gear wheel Zr = number of teeth on the excentric gear w~eel ' :~
It will be understood that to obtain a large gear ratio the difference in the number of teeth should be small. But already at a difference of about ~ teeth there is risk of so-called tooth interference. See ~igure 12.
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~, ~o obtain a large gear ratio, the numbe- o_ teet:n mu~_ be ~"

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great so that the module ~size of the tooth) will not be too small. Such gears are used as reduction gears in large ships engines. Scalir.g down these gears to sizes suitable as servo gears is practically impossible since the module would be impossibly small.

Harmonic Drive is a type of gear which has managed to avoid the problem of the teeth interfering with each other while maintaining a high gear ratio. In this gear the means corresponding to the moving gear wheel comprise a soecial thin wall ball race which is pressed on to an elliptical inner former, this ball race then receiving an externally toothed steel ring. This now oval-shaped ring meshes with an internally toothed rigid steel ring, which is also stationary.
Thus, in this Xnown apparatus two diametrically opposed parts of the internally toothed ring are in engagement with the elliptical, elastically shaped means. As a result of the elliptical shape, the unmeshing teeth are kept away from the regions where they would otherwise interfere with the teeth of the stationary ring- For this known structure to function well and have small backlash, it is required that very close tolerances on the comoonent parts are maintained, which results in that this gear becomes expensive to produce.
Furthermore, both ~he moment of inertia and the frictional moment for the input shaft are large, which is a disadvantage, e.g. in connection with servo systems.

Cyclo Drive is another gear structure avoiding the interference problem by not utilising teeth with involute cutting, but a kind of cycloid tooth. The stationary internally toothed ring has this cycloid form, and pins axially fastened to a disc on the output shaft roll off the teeth one arter the other. The radial forces in this structure ~, are very high, and very tight tolerances on the components are ~ reauired to ootain sm211 backlash. The greatest disadvantage `~ witr. ~his gear is that a superposed pulsation on the output 3 132~119 shaft is obtained, i-e- the gear does not give a -- e angular transmissior.. The Dojan is a variant o-^ this ty?e of gear.

Conventional spur gear boxes comprise a plurality o- reduction stages. The greater the total ge~r ratio desired, '-.e greater the number o. reduction stages. This gear box thus comprises a plurality of reduction stages connected in series. The total gear ratio is the product of the gear ratios in the participating reduction stages. In order that the gear box is not given a too large radial extension, gear ratios o' more than 6:1 are seldom used in each stage. In practice, about four reduction stages are required to obtain a gear ratio of 100:1. The disadvantage with this chain of reduc-ion stages is that only a few teeth are simultaneously in mesh in each reduction stage- The gearing is resilient and ~he total backlash is large. This gearing type is not suitable in servo systems, where it is important that the gearing has small backlash and great rigidity.

So-called differential gearing is often used for achieving large reductions- Such gearing usually comprises a planet gear in which the sun wheel is driven at a given rotational rate and the outer, internal gear is d.iven in the opposite direction at approximately the same rate. There th~s occurs a revolutionary rate difference which can be taken ou- from the planet wheel directly on the output shaft. The great disadvantage with this gearing is that it has a very low efficiency. If the rotational rate di~ference is very small;
the unuseful power can attain almost at 100~. Since i~ is also desired here that the gearing shall have small backlash, tight manufacturing tolerances are required.

The o-esent invention has the object o achievir.g ar excentric ~eari..g of the kind mentioned in the in.roduc ion, with a one-tooth difference in the number o' teeth. The gea: ng shall have great reductior., low fric~ional moment, low ~oment of 4 132~119 inertia for the input sha~-, dynamically balances operation, nesligible backlash and hich rigidity.

The inventiOn also discloses a method of gra~hically con igurating the teetn o. the s-ationary and excentric gear wheels in an excentric gear in accordance with the invention.
By this graphic confi5uration of the gear wheels, it will be possible to select the difference in the number of teeth of the two gear wheels to be the least possible, i.e. 1, whereby the reduction will be the greatest possible, simultar.eously as the number of teeth in mesh will be maximum and the gear thus rigid.

The distinguishing features for the invention ar- apparent from the accompanying claims. The invention enables obtaining very small backlash with normal workshop tolerances between the two gear wheels of the excenter gear. For the one-tooth difference in the number of teeth, the teeth must have shifted profiles, as well as being stubbed and possibly having the pressure angle corrected: This correction must be made taking into account that the length of engagement between the teeth ln mesh shall not be too small. A short meshing leng-h reduces the power possible to .~ansmit by the gear. At the same time, the clearance between the the tooth faces must not be too tight. These two factors must be weighed against each other. A
condition for practically being able to use the corrected gear wheel teeth is in other words that the excentric means eliminates certain manufacturing tolerances. Tooth interference will other-~ise be obtained.
~he invention permits manufacturing of a reduc_ion gear with small oacklash, but not a gear with large backlash.

~iff_~ ent embocimen s c- the invention will now be desc~ibed ln more detall i.. conr.ection with the accompanying drawings where, 132~119 Figure l is a~. ex?loded view o- a first embodiment of an excentric gear .- accordance with the presen. invention, Figure 2 is a _-ont view of a fi-st excentric means, Figure 3 is a front view of the excentric gear wheel, Figure 4 is a front view of a second excentric means, Figure 5 is a front view of a driving dog, Figure 6 is a side view of the driving dog i!lustrated in Figure 5, Figure 7 is a side view of the driving dog in Figure 5, along the line VII-VII in Figure 5, Figure 8 is a front view of the output shaft, Figure 9 is a longitudinal cross section of the gear output ;;"~4: shaft in an assembled condition, Figure 10 is a longitudinal cross section of the gear input part in an asse.~bled condition, Figure 11 is a longitudinal cross section of the assembled -~ input and output parts of Figures 9 and lO, ~., Figure 12 depicts meshing of teeth in an excentric gear with uncorrected teeth and a tooth difference of 8 teeth, .
Fic~re 13 depic~s meshing of teeth in an excentric gear with unc~rrected teeth and a tooth difference of l tooth, Figure 14 de~ic~, meshing of teeth in an excentr-c gear, with 6 132~119 a tooth difference of 1 tooth, the teeth having been corrected by profile shlfting, stubbin- and modifica-ion o- pressure angle, Figure 15 is an enlarged depic_ion of tooth meshing with engagement length, .

Figure 16 is an enlarged depiction of a region illustrating a pair of teeth with clearance dimensions.

The excentric gear lncludes a cylindrical housing 1 with a stationary gear wheel 2 having internal teeth, the wheel being made integrally with the housing 1. A plurality o. screws 2 are intended for use in fitting the gear to the endwall of a motor. The excentric gear further includes: a first excen~ric means 4 with a first counterweight S, a second excentric means 6 with a counterweight 7, a first ball bearing 8, an excentric gear wheel 9, also termed satellite wheel, an outpu~ shart 10, two pivot pins 11, 12, a bush 13, a driving dog 14, a second ball bearing lS, a bearing housing 16, a third ball bea_ing 17, a washer 18, a shims washer 19 and a circli? 20.

It will be seen from Figure 2 that the first excer.tric means 4 has a counterweight 5 in the form of a tongue projecting radizlly liXe a flange, and situated at one end of the sleeve 4 forming the mounting of the excentric means. The second excentric means 6 illustrated in Figure 4 has a similar configuration as the means 4 and is therefore not described in detail. The sleeve of the second excentric means 6 has less excentricity than that of the first excentric means 4. Figure 3 is a front view of the excentric gear wheel 9 with its teeth 21. This wheel has two diametrically opposing openings or "tooth gaps" 22, 23 in its side sur'ace facing towards the driving means- The whe_l is aiso Drovided on this side wit.h an outs'anding annular ring 2~ accommodatins the f rst ~all ~eari3s 8.

' 1325119 It will be seen from Fi?ures 5-7 that the driving dog includes an annular ring part 2~, on one side o- which projects two diametrically opposed lugs or teeth 26, 27. These luss are intended to slide reciprocatingly in the elongate openings 22, 23 on the excentric gear wheel. To ensure freedom from play in this motion, each lug is provided with a sli~ 23, clearly shown in Figure 7. There are two radial through bores 29, 30 in the annular ring, these bores being diametrically opposed and with the same angular position as the pins 26, 27 on the annular ring 25. These bores are intended to receive the oivot pins 11 and 12. The driving dog 10 also has bevelled parts 31 arld 32 forming either side surface of the annular ring.

The output shaft 14 associated with the driving dog is provided with a flange 3~ of a radial extension allowing it to be accommodated within the annular ring part 25 of the driving dog. This flange is provided with two diagonally disposed bores 34, 35 for the pivot pins 11 and 12, which are a press fit in the bores- The output shaft is thus pivotably mounted on the pivot pins 11, 12. As will be seen from Figures 1 and 8, the output shaft has a central bore 36 for accommodating the bush 13 at its end provided with the flange 33. An annular sroove 37 for the circlip 20 is arranged on the out?ut shaft a distance away from the flange 33. The output part of the gear is assembled in the following way: The pivot p- ns 11, 12 are first thrust into the bores 29, 30 on the driving dog and into the bores 34, 35 on the output shaft flange. The bush 13, servi~g as a be3ring for the not yet described input shaft to the excentric gear is pushed into the central bore 36 on the output shaft. The second ball bearing 15 is then mounted on the output shaft and subseauently thrust into the bearing housing 16. The third ball beari..g 17 is then moun-ed on the outpu_ shaft from the outside o_ the bearing housing 16. The washer 18 and the shims washer 19 are then l~laced on the sha'~, the whole assemblv being loc~ced in placs bv the ci-clip .

.~ 20. The completed assembly illustrated in Figure g is thus prestressed and thus constitutes the output part o- the sear.
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The input part of the gear is assembled in the following way:
The housing 1 is screwed by the screws 3 onto the end wall of a motor 39 with an out~ut shaft 40, illustrated schematlcally in Figure 10, this shaft thus constituting the input shaft of the gear. The ~irst ball bearing 8 is then glued into the excentric gear wheel 9. A suitable glue is Araldite or Locktite. The exterior cylirdrical part of the sleeve in the second excentric means 6 is then coated with glue and the means thrust into the inner ring of the ball bearing 8. Glue is then applied to the exterior cylindrical surface of the sleeve pertaining to the first excentric means 4 and this excentric means is thrust from the other side of the excentric gear wheel into the second excentric means 6, these two means 4 and 6 then assuming a mutual angular position such that the combined excentricity is minimum. Glue is then applied to the internal cylindrical surface of the sleeve of the excentric means 4 and the whole unit 4, 6, 8, 9 is thrust onto the output shaft 40 of the motor. Glue is then removed from the part of the sha-t 41 projectlng out over the upper surface of the first excen'ric means.4 and its counterweight 5. The excentric means 4 and 6 are then turned relative each other, causing the excentricity to increase, until there is no backlash between the teeth on the excentric wheel 2 and those of the fixed sear wheel 2, with which the excentric wheel is in mesh. The glue is then allowed to harden. By turning the excentric means relative each other in this way, all backlash is reduced to a minimum, and thereby manufacturing tolerances for the excentric and fixed gear wheels as well as the excentric means are eliminated. The only throw which now exist is that in the ball bea~ing 8, and since this normally is not greater than about 5 ~m it can be said that all backlash in the gear has beer. eliminated.
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9 132~119 Figure 10 illustrates the ur.it ~hich is obtained when the input part o' the gear is mounted in the descri~ed .~anr.e~. The input and out~ut parts are then fi-ted together into the unit illustrated in Figure ll. The bearing housing 16 is fixed to the housing 1 by the radial pins ~11, set screws or the liXe.

It should be noted that the excentric gear wheel is freely mounted on the motor shaft. If the gear ratio of the gear is 89, this means that when the input shaft has rotated 89 revolutions the excentric wheel has rotated one revolution in the opposite direction. Since the excentric whee7 rotates synchronously with the output shaft, the polar moment of the inertia of the outpput shaft as reduced to the input shaft (=
the output shaft of the motor), is negligible, since the gear ratio participates with a factor to the power of two.
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The moment of the inertia J of the system can be regarded as comprised of the moment of inertia on the out~ut shaft, driving dog and excentric wheel, all three being reduced to the input shaft by dividing by the square of the gear ratio (i). This resulting moment of inertia is negligible. The total moment of inertia of the system is therefore solely the mass of the excentric wheel multiplied by the square o' the excentricity. Furthermore, there is the moment of inertla of the counterweights 5, 7 which should be dominating. The rotation radius of the balancing mass should be kept low so that its effect on the moment of inertia of the ouput shaft will be small. For a typical excentric gear wheel, the moment of inertia J = 0,046 x lO 8 kgm2 and the moment of inertia of the counterweights J = 0,8 x lO 8 kgm2, the total moment of inertia, J = 0,9 x lO 8 kgm . This total moment of inertia is thus the one for the input shaft and this value should be compa_ed wit~ the input moment o' the iner.ia for the Harmonic gear which is stated to be 15 x 10-8 Xgm2 for the correspondina sear size. In spite of the excentric gear wheel havir.g low rotational rate, i's centre o- gravity rotates with lo 132~119 the high rotational rate of the inpu~ shaft 40.

Figu-e 12 illustrates certain conditions in a conven-ional excentric gear- The excentric gear wheel is freelv mounted on the excentric s'ub shaft of the input shaf- with the excentricity e. The pitch diameters are here 27 and 24, 6, the tooth module 0,3 and the number of teeth 90 and 82. The dif'erence in the number of teeht is thus 8. This gives a rather small gear ratio i = -10,2;. It can however be seen that here already there is danger of cog inte~erence in an earea the extremity of which is denoted by an arrow.

Figure 13 illustrates an uncorrected excentric gear of the same type as in Figure 12, but with a tooth difference of 1 tooth. The gear ratio is 89:1. It will be seen that there is heavy cog interference, which is apparent fro~ the area between 6 and 9 o clock.

Figure 14 shows the same gear as in Figure 13, but corrected.
The tooth difference is one tooth. The gear ratio is 89:1.
.~odi ication by profile shifting xm = -0,18 mm, stubbing and prsssure angle change to 22 degrees. The tooth face clearance is here 0.06 mm- Theoretically, only one tooth par in each loading direc~ion is simultaneously in mesh. In practice, the teeth are slightly resilient during loading and one can therefore recXon with about 8 teeth being simultaneously in - -mesh in the load direction, which gives very good load distri~ution.

Figure 14 shows the result of the method proposed in accordance with the invention for graphical reproduction of teeth, for a fixed internally toothed gear wheel and an excer.~~ic gear wheel in the excen~-ic gear acco_d~ng to the invenlion. The _ixed wheel is denoted by 2 and the excent~ic wheel by 9. The ~itch circle of the excentric wheel is denoted -v ;~ and t~.at o- the fixed wheel by 56.

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11 132~119 ~s in?ut data fo~ the g-aphical r-production there is used the desired gear ratio for the excentric cear, in this case ~9. As an input value for the calculations, the apDroximate pitch diameter 56 (denoted by Ds in ~isure 14) is also given, and in this case i_ has been selectec to 27 mm. As the last input value there is also given the desired value o' the tooth module (reciprocal of the diametral pitch) in this case m =
0,3. A first approximate value of the pitch diamter of the fixed wheel is then calculated from these three input values.
These values are also used for calculating a firs~ basic excentricity eO which is equal to half the pitch diamter Ds for the fixed wheel reduced by half the pitch diameter ~r for the excentric wheel. The number of teeth (Zr and Zs) is calculated for each wheel. If the number of teeth is not an integer, the pitch diameter So is changed and the process repeated. When the number of teeth is an integer the teeth of the the two gear wheels are drawn (Figure 13) suitably using CAD, to a considerably enla5red scale, as illustrated in Figure l; and 16 and the areas with potential cog inteference are inspected- The areas where the teeth are in mesh are satisfactory, but where they start to mesh it must be arranged that the fzces of the respec-ive teeth do not collide, and that the top lands of the teeth do not collide. To correct such undesirzble collisions, a profile shif' is arranged, such that the basic excentricity is changed to a new value eO + x.m where m is the module and x is a profile shift factor. The top lands of the teeth are stubbed to prevent them from colliding.
The teeth of thetwo gear wheels are then drawn with the teeth corrected in the way mentioned by profile shi'ting and stubbing. The result is illustrated in Figure 14. I~ will be understood that profile shi'~ing does not affect the pitch c rcles. The teeth thus drawn are inspected, and a least cistance between the tooth lands is decided 'or the teeth which are not in mesh. If this distan-e is not suf^ cent for acceptance wit^. ~ecar~ to r.or.-correctable manu ac~u ing 12 132~119 tolerances, the teeth are corrected once again. The r~ference numeral 57 denotes in Fisure 14 the place where ~he dis.ance between the two top lands is minimum, namely 0,06 mm, al.hough sufficently large to avoid collision. As will be seen from .-igure 14, abo~t 8 teeth are in mesh with each other between abou. 10,30 and 11,30 (clockface)- There is also cor.esponding meshing on the other half between 12.30 and 1.30, i.e.
symmetrically about the vertial axis in Figure 4. I~ is thus clear that a large number of teeth are simultaneously in mesh, which inc~eases the rigidity of the gear. It snould be noted that meshing ta'ces place in these two areas and that there is clearance in the area around 12 o clock. This is favourable from the point of view of backlash. Finally, in figure 14 Zr denotes the number of teeth 89 on the e.xcentric gear wheel, Zs the number of teeth 90 on the fixed 5ear wheel, Zr the diameter of the circle the tooth top lands on the satellite gear wheel described and Zs the diameter of the curve the tooth top lands on the fixed wheel described. As will be seen from Figure 14, the pitch diameters 55 and 66 cut each other at approximately 11-30, so that the pitch diameter 5; for the excentric wneel is outside the pitch diameter 57 for the flxed wheel in the area betweer. approximaterly 10.30 and 1.30. Since the teeth have been stubbed, there are no collisions with the bottom lands within this area.
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Figure 15 illustrates a tooth pair in mesh. The length variation of the active face width can be determined here. 8y urther enlargement and measurement it can be determined from tooth pair to tooth par how well the tooth faces nestle against each other, and if it is necessary to make pressure angle corrections.

~he tooth to~ lanc clea~ance dimension at 57 ir. Figure 14, shown at an enlarged scale as f in Figure lo, is very impor~ant. ~ dimension wnich is too small gives risk for tooth inter~erence anc ~uts ;arge demancs or. the tole-ances o' the .

132~119 1~

components included. Too large a dimension reduces t..e meshing leng.h of the teeth.

The graphical method for producing the teeth o the two gear wheels just described is extremely convenier.t, com?ared with the very extensive calculation work which would be requi-ed in the case where each tooth were to be calculated by itself, and furthermore there would be a poor general impression o' the tooth situation- Starting with the graphically produced gear wheels, the gear wheels are then produced using conventional tooth cutting technique.

The embodiments of the invention described above can be modified in many ways and varied within the scope of the inventive concept- Instead of gluing the excentric means to each other, the counterweights can be prodivded with loc~ing elements e.g. pins on one counterweight and grooves or openings on the other one for fixing the relative angular turn between the sleeves of the excentric means. Either of the excentric means 4, 6 can be provided with a balancing weight for statically balancing the imbalance moment o. the excentric elements.

Claims

The embodiments of the invention in which an exclu-sive property or privilege is claimed, are defined as follows:
1. Translation element, a so-called driving dog or transmitter, for translating the rotation of the centre of gravity of an eccentrically mounted eccentric gear wheel to a first shaft while the eccentric gear wheel simultaneously executes a planetary movement about a second shaft.
characterized by a rigid body mounted for pivoting about first pivot pins which are disposed substantially at right angles to said first shaft, said rigid body having in its symmetry plane for the first pins and first shaft axially directed, tooth-like projections which fit with negligible play into two radially counter-directed apertures in the eccentric gear wheel for allowing said projections to execute radial sliding reciprocatory motion in the apertures simultaneously as they execute a rolling movement therein, and the rigid body pivots round the first pins.
2. Translation element as claimed in claim 1 characterized by slits formed in the projections for prestressing the projections in the apertures.