Ball bearing construction with tilt compensation

Abstract

A dental handpiece and method of turbine support therein, comprises a turbine wheel having a rotor shaft rotatably supported by a housing via ball bearings, and configured to rotationally drive a clamping chuck, at least one of the ball bearings including, an inner ring arranged on the rotor shaft, the inner ring having a raceway, an outer ring, having an outer raceway movably supported by the housing, and balls arranged between and guided by the inner and outer raceways. The raceway of the inner ring, in cross section, has a concave shape with a curvature corresponding to a radius of the balls. Optionally, the raceway of the outer ring in cross section, has, in a central area, a first curvature with a first radius and, adjoining the central area, a second curvature with a second radius, the first radius being smaller than the second. Optionally, the raceway of the outer ring has a substantially parabolic profile.

Claims

1. A ball bearing, comprising: an inner ring, the inner ring having an inner raceway; an outer ring, the outer ring having an outer raceway; and a plurality of balls arranged between and guided by the inner and outer raceways, wherein: the inner raceway of the inner ring, in cross section, has a concave profile with a curvature having a radius that substantially corresponds to a radius of the balls, the outer raceway of the outer ring, in cross section, has a central area having a first curvature having a first radius and an area adjoining the central area having a second curvature having a second radius, and the second radius is from about 50% to about 100% larger than the first radius.

2. The ball bearing according to claim 1, wherein a tangent to the first curvature at a boundary point between the first curvature and the second curvature is congruent with or has a slope mutually identical to a tangent to the second curvature at the boundary point.

3. The ball bearing according to claim 2, wherein the first curvature extends over an angular range of from about 8° to about 15°.

4. The ball bearing according to claim 1, wherein the ball bearing is for a floating installation.

5. The ball bearing according to claim 1, wherein the first curvature extends over an angular range of from about 8° to about 15°.

Description

DETAILED DESCRIPTION OF THE INVENTION

(1) The invention will now be explained in more detail on the basis of an embodiment with reference to the attached drawings, in which:

(2) FIG. 1 shows a cross section through a dental handpiece with a dental tool inserted therein.

(3) FIGS. 2a to 2c show cross sections through ball bearings according to the prior art in the untilted state, wherein FIG. 2a illustrates a radial deep groove ball bearing, while FIG. 2b illustrates an angular ball bearing, in which the inner ring is formed with only one shoulder, and FIG. 2c illustrates a further angular ball bearing, in which the outer ring is formed with only one shoulder.

(4) FIG. 3 shows a cross section through a ball bearing according to the prior art when the inner ring is tilted.

(5) FIG. 4 shows a cross section through a ball bearing according to the invention when the inner ring is tilted.

(6) FIG. 5 shows a cross section through the outer ring according to the invention of a ball bearing installed in the dental piece of FIG. 1.

(7) As is shown in FIG. 1, a dental handpiece has a handle 12 and a housing 10, at a distal end of the handle 12, for receiving a chuck 14 and a turbine wheel 16. A rotor shaft 18 is connected both to the turbine wheel 16 and to the chuck 14, so that a rotation of the turbine wheel 16 is transmitted to the chuck 14 via the rotor shaft 18. In other words, the turbine wheel 16, the rotor shaft 18 and the chuck 14 rotate together.

(8) A dental tool 20, for example a drill, can be inserted into the chuck 14, so that the dental tool 20 can rotate together with the chuck 14, when the turbine wheel 16 is set in rotation by compressed gas, such as compressed air.

(9) The rotor shaft 18 is rotationally supported in the housing 10 of the dental handpiece via two ball bearings 30. Each of the ball bearings 30 has a plurality of balls 31 that are guided by means of a ball bearing cage 33. The rotor shaft 18 is inserted into the inner rings 32 of the ball bearings 30, for example by means of a press fit or by means of a sliding fit and simultaneous adhesive bonding of the inner rings 32 to the rotor shaft 18. In contrast to this, the outer rings 34 of the ball bearings 30 are elastically supported in the housing 10 by the elastic elements 40, such as O-rings, being arranged in corresponding recesses of the housing 10.

(10) In this way, the outer rings 34 are not rigidly fixed in the housing 10, but can move in the housing 10 to a certain extent and thus also tilt. In addition, the outer rings 34 are each axially resiliently biased via a wave spring washer 50 and can also compensate for or perform axial movements due to the elasticity of the wave spring washer 50, which can reinforce the tilting of the outer rings 34. Due to the installation situation of the rotor shaft 18 in the inner rings 32, they can tilt as well.

(11) Thus, the entire rotary assembly consisting of the two ball bearings 30, the rotor shaft 18, of the turbine wheel 16 and the chuck 14 in supported elastically in the housing 10 via the elastic element 40 and the wave spring washer 50. This elastic support of the rotary assembly can damp vibrations in an excellent manner, so that impairments due to vibrations can be reduced to a minimum both for the dentist and for the patient.

(12) Although FIG. 1 shows the support of the outer rings 34 via the two elastic elements 40 in conjunction with the two wave spring washers 50, also two elastic elements 40 can be arranged on the outer rings 34 in conjunction with only one wave spring washer 50. Alternatively, only the two elastic elements 40 can be arranged and the wave spring washers 50 can be dispensed with. If one or both wave spring washers are dispensed with, the elastic element(s) 40 at the same time serve(s) as a radial damping element and as an axial biasing member by arranging the elastic element(s) 40 in such a way that they can also receive axial forces. In this case, the outer ring(s) 34 is/are provided with a recess in the form of a step at the outer circumference.

(13) FIG. 2a shows the cross section through a portion of a ball bearing in the form of a radial deep groove ball bearing with untilted inner and outer rings. In ball bearings or deep groove ball bearings, the raceways L of the inner and outer rings have, seen in cross-section, an identical or uniform radius of curvature, seen over the entire cross section of the raceways L, which is slightly larger than that of the ball. In addition, prior to their installation, such ball bearings have a so-called radial clearance, i.e., seen in cross section, a small distance between the outer circumference of the balls with respect to the raceways L of the inner and outer rings in order to not clamp the balls between the inner ring and the outer ring. After the installation of the ball bearings, therefore, once again seen in cross section, the inner and outer rings must be displaced relative to one another in the horizontal direction (the so-called setting of a ball bearing) in order to enable the balls to rest against the raceways L of the inner and outer rings. The points at which the balls then rest are referred to as contact points.

(14) An axis H extending vertically when seen in cross section, as shown in FIG. 2a, is perpendicular to the axis of rotation (not shown) of the ball bearing before the ball bearing is positioned and runs through the center of the ball and, seen in cross section, the lowest or deepest points of the raceways L of the inner and outer rings.

(15) As a result of the required setting of the ball bearing and the associated mutual horizontal displacement of the outer ring relative to the inner ring, seen in cross section, the theoretical contact points of the ball originally, i.e. prior to the installation of the ball bearing, located at the lowest point of the raceways are displaced. Due to this displacement, the contact points thus migrate to an axis A, which is offset by an angle α relative to the axis H. The axis H remains vertically and running through the center of the ball after the setting. In FIG. 2a, the newly resulting contact points are designed with K.sub.A (contact point on the outer ring) and K.sub.I (contact point on the inner ring). In other words, when seen in the cross section illustrated, an offset of the theoretical contact points located at the lowest point of the raceways with respect to the contact points K.sub.A and K.sub.I now located on the axis A results, which can also be referred to as a pressure angle α between the axes H and A.

(16) With the described offset, i.e. even if the contact points K.sub.A and K.sub.I have the pressure angle α with respect to the vertical axis H, the balls of the ball bearing during operation also run on a circular path that does not excessively wear the ball bearing, which can also be referred to as a circular path, as long as the inner ring and the outer ring do not tilt relative to each or do not mutually tilt.

(17) FIG. 2b shows, analogously to FIG. 2a, an angular ball bearing, in which the outer ring is identical to the outer ring of the radial deep groove ball bearing of FIG. 2a and is therefore not explained in more detail here. In contrast to the radial deep groove ball bearing shown in FIG. 2a, the angular ball bearing of FIG. 2b, however, has an inner ring with only one shoulder, i.e. the concave raceway or groove is not symmetrical, but has a shoulder on only one side, while the opposite side is substantially flat, i.e. substantially horizontally or slightly conically in the section shown.

(18) Moreover, FIG. 2c shows an angular ball bearing comprising a symmetrical inner ring and an outer ring, which has only one shoulder.

(19) Due to the elastic support of the outer ring in the housing 10, a relative tilting of the outer ring relative to the inner ring may occur. This is to be explained in more detail on the basis of the following model calculation, wherein firstly the tilting of a ball bearing according to the prior art is examined (FIG. 3) and then compared to a tilting of a ball bearing according to the invention (FIG. 4).

(20) A relative tilting between the inner and outer rings of a ball bearing is shown in FIG. 3, on the basis of a tilting of the inner ring with respect to the outer ring. Although FIG. 3 shows the tilting on the basis of a radial deep groove ball bearing, a corresponding angular ball bearing may have the same tilt. The designations of FIGS. 2a to 2c are adopted, since otherwise the cross section shown does not differ from the cross section of FIGS. 2a to 2c. Such a relative tilting, however, results in an elliptical raceway of the balls that is excessively wearing the ball bearing during operation, which can also be called an elliptical path, regardless of which pressure angle is present. The balls therefore no longer run on the circular path during operation, but on an elliptical path, which cause corresponding undesired accelerations and decelerations of the balls during operation of the ball bearing.

(21) Due to the tilting, the contact points K.sub.I and K.sub.A are additionally displaced towards an angle β, which, when seen from the axis H, is greater than the pressure angle α. The angle β is approximately the same on the inner ring and on the outer ring. This larger angle β will be referred to as the contact angle below.

(22) The greater the contact angle β with respect to the horizontal axis H, the more the shape of the elliptical path changes, since when the contact angle increases, referred to as β in FIG. 3, the foci of the resulting ellipse diverge further apart, the ellipse thus assumes a narrower or elongated shape.

(23) As in the ball bearings according to FIGS. 2a to 2c, in the embodiment according to FIG. 3, the contact point K.sub.A on the outer ring 34 and the contact point K.sub.I on the inner ring 32 are displaced away from the axis H to an axis B. In the case of the tilting of the inner ring relative to the outer ring shown in FIG. 3, there results a contact angle β by which the contact points K.sub.A on the outer ring and K.sub.I on the inner ring move away from the axis H in the raceways L, which due to the tilting, as a rule, is larger than the pressure angle α of FIG. 2 when the inner ring and the outer ring are non-tilted with respect to one another. Consequently, in the event of a tilting of the inner ring to the outer ring, not only the described wear-generating elliptical path, but also a larger contact angle β of the contact points K.sub.A and K.sub.I results.

(24) A relative tilting between the inner and outer rings of a ball bearing, as shown in FIG. 3, is also shown in FIG. 4, again by way of example only on the basis of a tilting of the inner ring to the outer ring. FIG. 4 corresponds to the embodiment of FIG. 3, wherein in FIG. 4 the outer ring is now configured according to the invention (as will be explained in more detail with reference to the following FIG. 5). The inner ring remains unchanged and is thus identical to the inner ring shown in FIGS. 2 and 3.

(25) Therefore, the contact points are K.sub.I and K.sub.A again, which due to the change in the raceway L of the outer ring 34 according to the invention now move less far away from the axis H than is the case in FIG. 3. It is understood that a comparison between the embodiments of FIGS. 2 to 4 is only possible if the ball bearings shown there have identical dimensions.

(26) It becomes clear from FIG. 4 that the contact angle with respect to the axis H, here designated with γ, and thus the offset of the contact points K.sub.I and K.sub.A, which are now located on the axis C, is considerably smaller than in the prior art according to FIG. 3. The angle γ is approximately the same on the inner ring and on the outer ring. Due to the considerably smaller contact angle γ relative to that according to the prior art, the shape of the ellipse established during operation of the ball bearing is substantially reduced with regard to the rolling of the balls 31, because the distance between the foci of the ellipse is considerably reduced at a smaller contact angle γ, so that the occurring accelerations and decelerations of the balls 31 within the ball bearing 30 are considerably reduced. This results in lower vibrations of the ball bearing 30 and, in addition, the service life of the ball bearing 30 is considerably increased even in the case of insufficient lubrication.

(27) Test bench trials were carried out with a radial groove ball bearing according to the prior art, as shown in FIG. 3, and with a radial deep groove ball bearing according to the invention, as shown in FIG. 4. Here, the results of the model calculation explained above were confirmed. It has been shown in particular that in the event of a relative tilting of the outer ring relative to the inner ring, the ball bearing according to the invention has a considerably longer service life.

(28) The inventive design of the raceway L of the outer ring 34 of the ball bearing 30 will now be explained in more detail with reference to FIG. 5. At this point, it should be noted once again that the raceway L of the inner ring (not shown) is not changed, i.e. it is formed with a constant curvature or constant circular arc as usual and known, when seen in cross section.

(29) In contrast to the previously known embodiments of the raceways in the outer rings of a radial deep groove ball bearing or angular ball bearing, in the raceway L of the outer ring 34 according to the invention, a first radius or radius of curvature r.sub.0, which, as usual and known, is slightly larger than the radius of the balls 31, only extends over an angle or angular range δ0 (which can also be referred to as the first raceway angle δ0), which is approximately 8° to approximately 15°, i.e. not over the entire cross section of the raceway L, i.e. not as far as the end thereof, as is illustrated in FIG. 5. When seen in cross section, the angle δ0 extends from the lowest point of the raceway L to a boundary point G. Due to the symmetry of the outer ring 34 shown in the cross section in FIG. 5, only the half extending to the left from the axis of symmetry S in FIG. 5 is to be explained, because the half extending to the right from the axis of symmetry in FIG. 5 has an identical design. In the case of the angular ball bearing, the raceway L, seen in cross section from the axis of symmetry S, in principle only extends to one side, as is known, while the raceway L on the opposite side of the axis of symmetry S ends just behind the lowest point, i.e., seen in cross-section, shortly behind the axis of symmetry S.

(30) From the boundary point G to up to the end of the raceway L, the first radius of curvature r.sub.0 transitions into a larger second radius or radius of curvature r.sub.1, to be precise up to the end of the raceway L. The radius of curvature r.sub.1 extends over a second angular range δ1 (which can be referred to as the second raceway angle δ1) and is approximately 50% to approximately 100% larger than the first radius or radius of curvature r.sub.0. Preferably, r.sub.1 is approximately 70% to approximately 80% larger than r.sub.0.

(31) A tangent applied to the first radius of curvature r.sub.0, at the boundary point G, is to be congruent with or is to have the same slope as a tangent applied to the second radius of curvature r.sub.1 likewise at the boundary point G, so that there is a smooth transition between the two different radii r.sub.0 and r.sub.1. Therefore, the center point M.sub.1 of the second radius of curvature r.sub.1 is offset relative to the center point M.sub.0 of the first radius of curvature r.sub.0 both in the vertical direction (in FIG. 5 in the direction of the axis of symmetry S) and in the horizontal direction (in FIG. 5 in the direction of the rotation axis (not shown) of the ball bearing 30), as is shown in FIG. 5.

(32) The contact point K.sub.A (see FIG. 4), which is not illustrated in FIG. 5, on the outer ring 34, in the case of a mutual tilting of the inner ring to the outer ring, is located in the area of the larger second radius of curvature r.sub.1 or in the area of the second curvature 36.

(33) Due to the design of the outer ring 34 of the ball bearing 30 (single-row radial deep groove ball bearing or angular ball bearing) according to the invention, at high rotational speeds such as occur in the dental handpiece, the vibrations generated by the ball bearing upon tilting of the inner ring relative to the outer ring are prevented, at least reduced to a very great extent. As a result, the ball bearing 30 can have a substantially higher service life even in the case of insufficient lubrication, as has been possible to date, so that failure probabilities are minimized. Possibly remaining vibrations can be damped via the floating or elastic installation mounting by means of the elastic elements 40 and the wave spring washers 50 (see FIG. 1) in order not to be transmitted to the handle 12 of the dental handpiece.

(34) By forming the outer ring 34 with the second curvature 36, which has the second radius r.sub.1, following the first curvature 35, which has the first radius r.sub.0, the contact points K.sub.A on the outer ring 34 and K.sub.I on the inner ring (not shown) are displaced less far away from the axis of symmetry S, so that the distance between the foci of the ellipse established by the tilting are reduced, which can therefore be approximated to a circular path easier compared to the prior art. Consequently, considerably lower accelerations and decelerations of the balls 31 occur during the operation of the ball bearing 30.

(35) It is to be noted again that the inner ring (not shown) has a constant curvature, seen in cross section, that does not change over the entire length of the raceway L, which in the present case corresponds to the first radius of curvature r.sub.0 of the outer ring 34, so that precise axial guidance of the balls 31 is ensured. Only the outer ring, seen in cross section, has at its raceway base the first curvature 35 with the smaller first radius r.sub.0 and subsequently, starting from the boundary point G, a second curvature 36 having a larger second radius r.sub.1.

(36) Although the exemplary embodiment describes the raceway L with the first curvature 35 and the second curvature 36, the same effects can be achieved with a raceway that, seen in cross section, is substantially parabolic.

LIST OF REFERENCE NUMERALS

(37) 10 Housing 12 Handle 14 Chuck 16 Turbine wheel 18 Rotor shaft 20 Dental tool 30 Ball bearing 31 Ball 32 Inner ring 33 Ball bearing cage 34 Outer ring 35 First curvature 36 Second curvature 40 Elastic element (O-ring) 50 Wave spring washer A Axis G Boundary point H Axis K.sub.A Contact point on the outer ring K.sub.I Contact point on the inner ring L Raceway r.sub.0 First radius r.sub.1 Second radius α Angle/pressure angle β Angle/contact angle γ Angle/contact angle δ0 First angular range/first raceway angle δ1 Second angular range/second raceway angle S Axis of symmetry M.sub.0 Center M.sub.1 Center