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FIELD OF THE INVENTION AND RELATED ART STATEMENT
1. FIELD OF THE INVENTION
The present invention relates to a bearing unit, and more particularly to a
hydrodynamic lubrication bearing unit using a hydrodynamic lubrication
film between a cylindrical shaft and a cylindrical bearing.
2. DESCRIPTION OF THE RELATED ART
Heretofore, many types of plain bearings were known. In such bearings, a
spiral-groove bearing unit, which is a kind of bearing unit utilizing
hydrodynamic lubrication, has lately attracted considerable attention as
being an excellent bearing unit to get high performance. FIG. 12 shows the
cross-sectional view of the conventional spiral-groove bearing unit, and
FIGS. 13 and 14 show pressure distributions of the bearing unit shown in
FIG. 12. In a hydrodynamic lubrication film 45, spiral grooves 44 which
are formed on a cylindrical surface 43 of a shaft 41 (or on a cylindrical
surface 42 of a bearing 40) generate pumping pressure vectors 50 of an
equal intensity shown by arrows in FIG. 13 by relative rotation between
both cylindrical surfaces 42 and 43. Ends of the vectors 50 form a
circular configuration 51, that is, uniform pressure is impressed to the
film 45. These pumping pressure vectors 50 serve to avoid unstable holding
of a kinetic pressure bearing at high-speed rotation, and thereby the
shaft 41 rotates in the bearing 40 without any eccentricity and touching
on the cylindrical surface 42. Therefore, the spiral-groove bearing unit
has very little noise and vibration, a precise dynamic holding of the
shaft 41 and a high stability of bearing within a small eccentricity which
is suitable for high speed rotation.
On the other hand, as shown in FIG. 16, a static pressure bearing unit
having circumferential grooves 46 on the rotating shaft 41 (or on the
cylindrical surface 42 of the bearing 40) has been used. High pressure
fluid shown by an arrow 52 is applied to the circumferential grooves 46
through an inner passage 41a and an orifice 47 which connects the inner
passage 41a with the circumferential grooves 46. When the shaft 41 has an
eccentricity against the bearing 40, a clearance between both cylindrical
surfaces 42 and 43 is changed, and thereby a part of large fluid friction
and a part of small fluid friction are formed in the hydrodynamic
lubrication film 45. As a result, pressure distribution in a radial
direction around the cylindrical surface 43 varies, and thereby the shaft
41 receives restoring force. Thus, touching between the cylindrical
surfaces 42 and 43 is prevented.
Hereupon, load capacity of the spiral-groove bearing unit is obtained by
wedge force 48 generated by the eccentricity of the shaft 41 as shown in
FIG. 14. FIG. 15 shows a relation between the restoring force and the
eccentricity of the shaft 41. In the figure, the restoring force by the
pumping pressure is generated at small values of the eccentricity but such
restoring force is very small. The pumping pressure which is uniformly
generated around the shaft 41 does not serve to get the load capacity in
radial direction but serves to stabilize the rotation of the shaft 41. The
restoring force by the wedge force acceleratively becomes large responding
to increase of the eccentricity, whereas the restoring force is small at
small values of the eccentricity. Thereby, the load capacity is small when
an allowable eccentricity is small. When it is necessary to get sufficient
load capacity for a bearing of revolving mirror of a laser printer, by
applying an air kinetic pressure bearing using air of low viscosity, the
clearance between the cylindrical surfaces 42 and 43 have to be made 2-3
.mu.m even at the rotation speed of 10-20 thousand r.p.m. Therefore,
high-precision is required for finishing of the cylindrical surfaces 42
and 43, and mass-production of such a high precision bearing is difficult.
In the static pressure bearing unit having the circumferential grooves, as
shown in FIG. 17, the clearance between the cylindrical surfaces 42 and 43
is changed by the eccentricity of the shaft 41, and thereby the fluid
friction in the hydrodynamic lubrication film 45 is changed. Thereby, the
pressure distribution in the radial direction is greatly changed
responding to the clearance. The restoring force generated by unevenness
of the pressure distribution and the wedge force added thereto give the
bearing unit a high load capacity, but a bulky and heavy pump apparatus is
required for supplying pressure to the static pressure bearing unit.
Therefore, it is difficult to apply this bearing unit to the bearing unit
for commodity goods, for instance a VTR cylinder, a laser printer for
office use or a hard disk for office use etc..
OBJECT AND SUMMARY OF THE INVENTION
The object of the present invention is to provide an improved bearing unit
which has a large load capacity without precision finishing and is stable
at high speed rotation, simple and compact.
In order to achieve the above-mentioned object, a bearing unit in
accordance with the present invention comprises:
a bearing;
a shaft which rotates relative to the bearing and has a plurality of
shallower groove thereon for pumping, at least one deeper circumferential
groove thereon and an inner passage therein for connecting a part on the
shallower groove to the circumferential groove; and
a fluid held in a gap formed between the bearing and the shaft, for forming
a hydrodynamic lubrication film between the bearing and the shaft by
relative rotation of the bearing and the shaft.
The above-mentioned bearing unit has a large load capacity, a stable
rotation in high speed without any unstable rotation like a whirl of the
shaft, and the configuration is simple. Further, since it is not required
to minimize a clearance between the shaft and the bearing, this bearing
unit is suitable for mass-production and enables a low rotation torque,
hence a low power loss, even when an oil is used as a fluid.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a cross-sectional view showing an embodiment of a bearing unit of
the present invention, showing outside face of a shaft, wherein symmetric
spiral shallow grooves are drawn by straight lines for the sake of
simplicity of drawing.
FIG. 2 is an internal cross-sectional view along an axis of the shaft in
FIG. 1.
FIG. 3 is a partial cross-sectional view showing pressure distribution
around the shaft 2 of FIGS. 1 and 2.
FIG. 4 is a cross-sectional view showing another embodiment of a bearing
unit of the present invention, showing an outside face of a shaft, wherein
spiral shallow grooves are drawn by straight lines for the sake of
simplicities of drawing.
FIG. 5 is an internal cross-sectional view along an axis of the shaft in
FIG. 4.
FIG. 6 is a graph showing a relation between pressure and position in a
bearing unit.
FIG. 7(a) is a cross-sectional view showing still another embodiment of a
bearing unit of the present invention, showing an outside face of a shaft,
wherein spiral shallow grooves are drawn by straight lines for the sake of
simplicities of drawing.
FIG. 7(b) is a cross-sectional view taken on line A--A of FIG. 7(a).
FIGS. 8, 9 and 10 are cross-sectional views showing still other embodiments
of the present invention.
FIG. 11 is an internal cross-sectional view along an axis of a shaft in
FIG. 10.
FIG. 12 shows a cross-sectional view along the axis of the conventional
bearing unit, showing an outside face of the shaft, wherein symmetric
spiral grooves are drawn by straight lines for the sake of simplicity of
drawing.
FIG. 13 and FIG. 14 are partial cross-sectional views showing the pressure
distributions around the shaft 41 of FIG. 12.
FIG. 15 is a graph showing the relation between restoring force and
eccentricity.
FIG. 16 is the cross-sectional view of other conventional bearing unit.
FIG. 17 is the partial cross-sectional view showing the pressure
distribution around the shaft 41 of FIG. 16.
DESCRIPTION OF THE PREFERRED EMBODIMENT
Hereafter, the preferred embodiment of the present invention is described
with reference to the accompanying drawings. FIG. 1 shows a
cross-sectional view of an embodiment of the invention, and FIG. 2 shows
an internal cross-sectional view along an axis of a rotation shaft 2 in
FIG. 1. The rotation shaft 2 penetrates a cylindrical hollow of a bearing
1 and is rotatably held therein. Between a cylindrical surface 3 of the
bearing 1 and a cylindrical surface 4 of the shaft 2 which face each
other, a hydrodynamic lubrication film 5 is formed by a lubricating fluid
like air or an oil. On the cylindrical surface 4, of the shaft 2, plural
spiral grooves 6a and 6b, which are symmetric with respect to a plane
perpendicular to axis of the shaft 2 at the center of the cylindrical
surface 4, are formed with a predetermined angle of inclination. These
spiral grooves 6a and 6b have a cross-section, for instance, of
rectangular shape, with a depth of from several to dozens of .mu.m and a
width of from dozens to several hundred .mu.m. Circumferential grooves 7a
and 7 b are formed on the shaft 2 inside both ends of the cylindrical
surface 3 with a predetermined interval. For instance, these
circumferential grooves 7a and 7b have a rectangular cross-section with a
width from several hundred .mu.m to 1 mm and a depth of dozens of .mu.m.
Radial passages 9 and 8 are formed in the radial direction from the center
of the shaft 2 to the circumferential grooves 7a, 7b and the center of the
cylindrical surface 4, respectively. And an axial passage 10 is formed in
the center of the shaft 2 in the axial direction thereof in order to
connect the passages 9, 8 and 9 with each other. By the radial passages 8
and 9 and the axial passage 10, a high pressure supplying passage 11,
which connects center parts of the spiral grooves 6a and 6b to the
circumferential grooves 7a and 7b, is formed. And ends of the radial
passages 9 opens in the grooves 7a and 7b, thereby forming an orifice
having a fluid friction. A hole 13, which is bored from an end of the
shaft 2 to a left end of the axial passage 10 in order to make the axial
passage 10 in the shaft, is closed by a screw plug 12. Sealing grooves 16a
and 16b are spirally formed on the shaft 2 in order to stop flowing-out of
lubricants to the outside.
In the above-mentioned construction, when the shaft 2 is rotated in a
direction shown by an arrow "A", pumping pressure is generated on the
spiral grooves 6a and 6b, and thereby hydrodynamic pressure on the axial
center part 5a between the cylindrical surfaces 3 and 4 becomes high. By
this pumping pressure, the whole circumference of the shaft 2 is uniformly
pressurized. As a result, whirling of the shaft 2 is prevented and a
stable rotation is offered. A part of the pumping pressure is led to the
circumferential grooves 7a and 7b through the high pressure supplying
passage 11, and thereby a function of static pressure bearing is realized.
FIG. 3 shows a pressure distribution around the shaft 2. In FIG. 3, when
the shaft 2 has a load in the radial direction and thereby brings the
eccentricity, a clearance on one side between the cylindrical surfaces 3
and 4 is made small and another clearance on the opposite side is made
large. Fluid friction in the small clearance increases thereby raising the
hydrodynamic pressure thereof, and the fluid friction in the large
clearance decreases thereby decreasing the hydrodynamic pressure thereof.
As a result, a large restoring force is operated to the shaft 2 against
the eccentrical direction, and thereby an axis of the shaft 2 naturally
returns onto an axis of the bearing 1 so as to eliminate the eccentricity
thereof. Thus, a large load capacity is given to the shaft 2 together with
stable rotation which is an inherent advantage of the spiral-groove
bearing unit.
In the conventional bearing unit which has only spiral grooves 6a and 6b
using low viscosity fluid such as air etc. as a lubricant, the clearance
between the cylindrical surfaces 3 and 4 was required to be 2-3 .mu.m in
order to get sufficient load capacity. But, in this embodiment, the
additional provision of the circumferential grooves 7a and 7b for the
function of the static pressure bearing enables the bearing unit to have a
clearance more than two times as large as that of the conventional unit,
to generate enough load capacity. At that time, the restoring force is
encouraged by the wedge force generated by the eccentricity of the shaft
2.
Next, another embodiment of the present invention is described. FIG. 4
shows a cross-sectional view of this embodiment, and FIG. 5 shows an
internal cross-sectional view along an axis of a rotation shaft 22 in FIG.
4. A bearing 21 has a cylindrical surface 23 therein with one end thereof
open and the bottom 36 thereof closed. An end part of the rotation shaft
22 is rotatably held by the cylindrical surface 23 of the bearing 22.
Plural spiral grooves 26 are formed on a cylindrical surface 24 of the
shaft 22 which faces to the cylindrical surface 23, and a circumferential
groove 27 is formed on the shaft 22 at a left end of the spiral grooves
26. An axial passage 30 is formed on the axis of the shaft 22 from the
right-end surface of the shaft 22 to a position below the circumferential
groove 27. Radial passages 29 are formed in the radial direction of the
shaft 22 from the passage 30 to the circumferential groove 27. A high
pressure supplying passage 31 is formed by these axial passage 30 and
radial passages 29 wherein fluid friction R.sub.0 exists thereby to form
an orifice. Between the cylindrical surfaces 23 and 24, magnetic fluid is
lubricated thereby forming a hydrodynamic lubrication film 25. And,
magnetic fluid sealing 32 using a radially magnetized magnet is provided
around a mouth of the cylindrical surface 23.
FIG. 6 is a graph showing a relation between pressure and position in a
bearing unit in the above-mentioned construction when the shaft 22 is
rotated and the pumping pressure is generated by the spiral grooves 26.
The graph shows that the farther the inner part of the bearing 21 (namely
the more rightward of FIG. 4 or FIG. 5) is, the more pressure P.sub.0 of
the magnetic fluid between the cylindrical surfaces 23 and 24 increases as
shown by a dotted line. And the pumping pressure P.sub.0 becomes maximum
at a position between an inner end of the surface 23 and a right end of
the shaft 22. By this pumping pressure, the whole circumference of the
shaft 22 is uniformly pressurized. As a result, an eccentric rotation of
the shaft 22 is prevented and a stable rotation is offered. A part of the
pumping pressure is led to the circumferential groove 27 through the high
pressure supplying passage 31, and thereby a static pressure bearing is
formed on the circumferential groove
Hereupon, relation between the pressure distribution and restoring force is
described. When relative amounts are defined as follows:
P; mean value of pressure on the circumferential groove 27,
.DELTA.P; pumping pressure,
P.sub.3 ; maximum pressure on inner ends of the cylindrical surfaces 23 and
24,
P.sub.0 ; maximum pressure on inner ends of the cylindrical surfaces 23 and
24 in case of no high pressure supplying passage 31,
Q; flow amount which passes through the high pressure supplying passage 31,
R.sub.0 ; fluid friction by the radial passage 29, and
R.sub.v ; mean value of fluid friction which varies responding to the
eccentricity of the shaft 22 between the cylindrical surfaces 23 and 24,
the following relation is held:
P.sub.3 =P+.DELTA.P=P+(P.sub.0 -Q(R.sub.0 +R.sub.v)).
Further, when other amounts are defined as follows:
P.sub.1 ; pressure on a position of minimum clearance in the
circumferential groove 27 under eccentrical condition of the shaft 22, and
P.sub.2 ; pressure on a position of maximum clearance in the
circumferential groove 27 under eccentrical condition of the shaft 22,
the pressure distribution is represented as in FIG. 6. An area of slanted
line which is surrounded by P.sub.1, P.sub.2 and P.sub.3 is operated as
the restoring force upon the shaft 2.
Besides the magnetic fluid sealing 32 provided in the above-mentioned
embodiment, spiral grooves 33 may be formed on the left side of the
circumferential groove 27 as shown in FIG. 7(a), in order to seal the
fluid by the pumping operation thereof.
FIG. 7(b) shows a cross-sectional view taken on line A--A in FIG. 7(a).
Thrust spiral grooves 37 for making a function as a thrust bearing are
formed on the end 35 of the shaft 22 as shown in the figure. The grooves
37 may be formed on the bottom 36. These grooves 37 generate another
pumping pressure, and thereby larger static pressure can be supplied to
this bearing unit.
In the above-mentioned three embodiments shown in FIGS. 1 through 7(b),
though the shaft 2 (or 22) is held by the bearing 1 (or 21), such reversed
construction that the shaft 2 (or 22) is fixed and the cylindrical bearing
1 (or 21) is rotatably held around the shaft 2 (or 22), is also
realizable. Further, the spiral grooves 6a and 6b (or 26) and the
circumferential groove 7a and 7b (or 27) may be provided either on the
shaft 2 (or 22) or on the cylindrical surface 3 (or 23). FIGS. 8, 9 and 10
are cross-sectional views showing still other embodiments of the present
invention, and FIG. 11 is an internal cross-sectional view along an axis
of the shaft in FIG. 10. In FIG. 8, both the spiral grooves 6a and 6b and
the circumferential grooves 7a and 7b are provided on the cylindrical
surface 3 of the bearing 1. In FIG. 9, the spiral grooves 6a and 6b are
provided on the shaft 2, and the circumferential grooves 7a and 7b are
provided on the cylindrical surface 3 of the bearing 1. In FIGS. 10 and
11, the spiral grooves 6a and 6b are provided on the cylindrical surface 3
of the bearing 1, and the circumferential grooves 7a and 7b are provided
on the shaft 2. Also, as for the embodiments shown in FIG. 4 and FIG.
7(a), similar variations based on whether the spiral grooves 26 and/or the
circumferential grooves 27 are provided on the shaft 22 or the bearing 21
can be realized.
Furthermore, for instance in FIG. 2, although at least three sets of radial
passages 9, 8 and 9 are necessary for ordinary use, only one or two sets
of radial passages 8 and/or 9 are applicable in case of single directional
load disposing the passage 8 or 9 to a position where the restoring force
is generated so as to offset the load. Also, since the radial passage 9
(or 29) is provided only for working as the fluid friction, the passage 9
(or 29) may be, not only orifice shaped, but also nozzle shaped which is
ordinarily used for the static pressure bearing.
Further, the circumferential groove 7a and 7b (or 27) may be formed on a
part of the circumference instead of making one round.
Further, grooves for generating pumping pressure are not limited to the
spiral configuration, but other configurations which give a similar
function are possible, as far as the lubricant is sent in one direction by
pressure induced by the relative rotation between the cylindrical surface
3 (or 23) of the bearing 1 (or 21) and the shaft 2 (or 22).
While specific embodiments of the invention have been illustrated and
described herein, it is realized that other modifications and changes will
occur to those skilled in the art. It is therefore to be understood that
the appended claims are intended to cover all modifications and changes
that fall within the true spirit and scope of the invention.
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