CA1186358A - Bearing segment for a hydrodynamic radial friction bearing - Google Patents
Bearing segment for a hydrodynamic radial friction bearingInfo
- Publication number
- CA1186358A CA1186358A CA000399441A CA399441A CA1186358A CA 1186358 A CA1186358 A CA 1186358A CA 000399441 A CA000399441 A CA 000399441A CA 399441 A CA399441 A CA 399441A CA 1186358 A CA1186358 A CA 1186358A
- Authority
- CA
- Canada
- Prior art keywords
- bearing
- segment
- lubricant
- supply groove
- grooves
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired
Links
- 239000000314 lubricant Substances 0.000 claims abstract description 19
- 238000009826 distribution Methods 0.000 claims abstract description 13
- 238000005461 lubrication Methods 0.000 claims description 7
- 229920000136 polysorbate Polymers 0.000 claims description 2
- 239000003921 oil Substances 0.000 description 16
- 238000010276 construction Methods 0.000 description 6
- 230000008901 benefit Effects 0.000 description 4
- 230000002706 hydrostatic effect Effects 0.000 description 4
- 239000010687 lubricating oil Substances 0.000 description 4
- 230000000694 effects Effects 0.000 description 2
- 238000007689 inspection Methods 0.000 description 2
- 230000003068 static effect Effects 0.000 description 2
- 208000031501 Emergencies Diseases 0.000 description 1
- 230000004075 alteration Effects 0.000 description 1
- 230000008859 change Effects 0.000 description 1
- 229910052729 chemical element Inorganic materials 0.000 description 1
- 238000001816 cooling Methods 0.000 description 1
- 230000006872 improvement Effects 0.000 description 1
- 230000001050 lubricating effect Effects 0.000 description 1
- 238000003754 machining Methods 0.000 description 1
- 239000002184 metal Substances 0.000 description 1
- 230000010355 oscillation Effects 0.000 description 1
- 230000002093 peripheral effect Effects 0.000 description 1
- 230000009467 reduction Effects 0.000 description 1
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
- F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
- F16C17/00—Sliding-contact bearings for exclusively rotary movement
- F16C17/02—Sliding-contact bearings for exclusively rotary movement for radial load only
- F16C17/03—Sliding-contact bearings for exclusively rotary movement for radial load only with tiltably-supported segments, e.g. Michell bearings
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
- F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
- F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
- F16C33/02—Parts of sliding-contact bearings
- F16C33/04—Brasses; Bushes; Linings
- F16C33/06—Sliding surface mainly made of metal
- F16C33/10—Construction relative to lubrication
Landscapes
- Engineering & Computer Science (AREA)
- General Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- Physics & Mathematics (AREA)
- Fluid Mechanics (AREA)
- Sliding-Contact Bearings (AREA)
- Magnetic Bearings And Hydrostatic Bearings (AREA)
Abstract
ABSTRACT OF THE DISCLOSURE
A bearing segment for a hydrodynamic radial friction bearing adapted to support a rotary member for rotation about an axis has an arcuate bearing surface ad-jacent one end of which is a lubricant supply groove ex-tending axially of the segment. The bearing segment also has a pair of lubricant distribution grooves in its bearing surface, one of which extends circumferentially of the seg-ment adjacent one lateral edge thereof and the other of which extends circumferentially of the segment adjacent its opposite lateral edge, the distribution grooves being in communication at corresponding ends thereof with the supply groove via passages through which lubricant may pass from the supply groove into both of the distribution grooves.
A bearing segment for a hydrodynamic radial friction bearing adapted to support a rotary member for rotation about an axis has an arcuate bearing surface ad-jacent one end of which is a lubricant supply groove ex-tending axially of the segment. The bearing segment also has a pair of lubricant distribution grooves in its bearing surface, one of which extends circumferentially of the seg-ment adjacent one lateral edge thereof and the other of which extends circumferentially of the segment adjacent its opposite lateral edge, the distribution grooves being in communication at corresponding ends thereof with the supply groove via passages through which lubricant may pass from the supply groove into both of the distribution grooves.
Description
The invention relates to a bearing segment for a hydrodynamic radial friction bearing with an axial lubri-cant supply groove which is provided adjacent one end of an arcuate bearing surface and serves to distribute lubricating oil over the axial dimension of the beariny and to form a hydrodynamic lubrication wedge between the bearing surface and a rotary element.
Bearings which are driven both hydrostatically and hydrodynamically are known. Both bearing systems have certain advantages and disadvantages.
Thus, as a rule the lubrication and inspection of a hydrostatic bearing is very costly, and -the emergency operating characteristics are poor. On the other hand, a hydrostatic bearing can be operated wi-th high surface loads, an approximately uniEorm surface load being pro-duced over the whole bearing surface. The average specific surface load thus approximately corresponds to -the maximum pressure on the bearing surface.
By contrast, the lubrication and inspection of a hydrodynamic bearing are considerably easier. Its emer-gency operating characteristics, tha-t is to say its be-haviour in the event of a sudden Eailure of lubrica-tion, are a great deal better than those of the hydrosta-tic bear-ing because of its smooth continuous bearing surface.
On the other hand, because of the way in which pressure builds up in the hydrodynamic wedge, enormous pressures occur in the so-called pressure peak in hydro-dynamic bearings or their supporting bearing segments.
The average specific surface loading of a hydrodynamic bearing is well below this peak pressure. As a result, a hydrodynamic bearing is of larger construction than a hydrostatic bearing with the same loading.
An object of the present inven-tion, therefore, is to provide a bearing segment for a hydrodynamic radial friction bearing which is distinguished by a higher average specific surface loading in comparison with known bearing constructions and permi-ts a par-ticularly narrow constructio ~-:
of the bearing.
The present invention provides a bearing segment for a hydrodynamic radial f:riction hearing adapted to sup-port a rotary member for rota-tion about an axis, the segment having an arcuate bearing surface adjacent one end of which is a lubricant supply ~roove extending axially of the seg-men-t, and means for delivering a lubricant into the axially extending supply groove, wherein -the bearing segment has a pair of lubricant distribution grooves in its bearing surface, one of which extends circumferentially of the segment adjacent one lateral edge thereof and the o-ther of which extends circumferentially of the se~ment adjacent its opposite lateral edge, the distribution grooves being in communication at corresponding ends thereof with the supply groove via passages through which lubricant may pass from the supply groove into both of the distribution grooves.
The German publication "VDI-BERICHTE", No. 111, 1966, pages 15 to 19, deals with the construction of radial friction bearings with a view to proper lubrication, and in particular page 19, illustration 15 IV, shows an embodi-ment in which a plurality of axially extending transverse grooves can be provided over the whole periphery of a radial friction bearing and a-t their axial ends the said grooves change directly into peripheral annular grooves. This means therefore that - quite apart from the fact that this is not a description of a bearing segment Eor such a radial friction bearing - the -transverse grooves are connected to the annular grooves provided at the axial bearing ends in such a way that the oil overflows from the transverse grooves into the annular grooves without any alterations in pressure. As a result of this known cons-truction, more oil can flow through the bearing in order to achieve better cooling, but an improvement in the bearing capacity and thus the economy of a hydrodynamic radial friction bearing cannot be achieved in this way.
A prior art bearing and two embodiments o~ the i3~
present invention are described below with reference to the accompanying drawings, wherein:-Figure 1 shows a plan view of a segment of a prior art hydrodynamic radial friction bearing;
Figure 2 shows a section along the line II in Figure l;
Figure 3 shows the pressure dlstribution along the periphery of the bearing segment in the plane II-II
of Figure 1 (i.e~ in the centre of the bearing);
Figures 4 and 5 show cross-sections through this pressure peak (along the lines IV-IV and V-V in Figure 3);
Figure 6 being a plan view of a bearing segment;
Figure 7 being a section along the line VII-VII
of Figure 6;
Figures 8, 9 and 10 are sections through the pressure build-up along the lines VIII-VIII, IX-IX and X-X, respectively;
Figure 11 shows a plan view of a bearing segment according to another embodiment of the invention; and Figure 12 shows a view taken in cross-section along the line XII-XII of Figure 11.
On the inlet side of the rotating element (not shown) in the bearing surface ? of this bearing segment 1 there is an axial lubricant supply groove 3 which is connected to a lubricating oil supply line 4 and serves to distribute the lubricating oil over the axial width of the bearing. This groove 3 serves in a known manner to supply the hydrodynamic wedge which forms between the bearing surface 2 and the rotating elemen-t during operation.
Depending upon the clearances, the mechanical stability of -the bearing segment 1 and the rotating element and the quantity of oil supplied, a more or less great sta-tic pressure builds up in the groove 3. From the ~uantity of oil present in the groove 3 the rotating ele-ment draws in sufficien-t into the gap to build the hydro-dynamic wedge. This "drawing in" is supported by the static 3~1~
pressure in the groove 3.
Some of the oil drawn in-to the gap be-tween -the bearing segment 1 and the rotating elemen-t now ~lows away from the bearing segment at the region of the longi-tudinal edges 2a, 2b of the bearing surface. This flow is un-desirable since it takes oil away from the centre of -the bearing which causes the pressure build-up to peak more, i.e. to be more uneven, and the height oE -the oil film to be reduced.
It is clear tha-t narrow bearings, i.e. bearings with a small axial dimension B, are particularly suscep-tible -to such a lateral flow.
Figures 6 to 10 show by contrast an embodiment of the invention.
As shown in Figure 6, the bearing segment 11 contains in its bearing surface 12 an axial lubrican-t supply groove 13 which is connected to a lubricating oil supply line 14~ The two ends of this groove 13 are connected via respective throttles or constrictions 15, 16 to cir-cumferentially e~tending grooves 17, 18 which extend axially of the bearing surface 12 close to the edges 12a, 12b.
Figures 8 to 10 show - in contrast -to Figures 3 to 5 - how the pressure build-up alters as a result of -the presence of the two circumferen-tial grooves 17, 18.
In the example shown in the drawing the oil pressure in the axial groove 3 is approximately 24 bars and in the circumferential grooves 17, 18 approxima-tely 20 bars.
The hydrodynamic lubrication wedge which forms in the gap between the bearing segment 11 and the rotating element during operation is consequently supported on its two flanks (i.e. in the region of the grooves 17, 18) against a higher pressure level than atmospheric pressure.
In this way flow past the edges 12a and 12b is consider-ably reduced and the volume of the pressure peak is in-creased approximately on the basis shown by the transverselines in Figures 8 to 10. This increase in -the volume oE the pressure peak corresponds to a propor-tional increase in the bearing capaci-ty of the bearing segment.
Thus it becomes possible either to increase the bearing capacity of the bearing ancl/or to make the beariny of narrower construction.
Referring now to Figures 11 and 12, which show a further embodiment of the invention, it will be seen that in this embodimen-t, in addition to the axial groove 33 provided on the inlet side of the rotating element and the two circumferential grooves 37, 38 connected via throttles 35, 36, the bearing segment 31 also contains in its bearing surface 32 an axial groove 39 provided on the outlet side of the rota-ting element and connected via throttles or constrictions 40, 41 to the longitudinal grooves 37, 38.
]5 The dimensions are advantageously such that -the throttles 35, 36 effect a pressure drop of 5 to 30~, pre-ferably 10 to 20%, in normal operation, whilst the -throttles 40, 41 effect a pressure drop of 70 to 95~, preferably 80 to 90~.
The construction according to Figures 11 and ]2 is particularly advantageous when the rotating element (for example a tube mill) opera-tes in only one direc-tion of rotation but oscillates in the other direc-tion of rota-tion when stopped. For this oscillation -the groove 39 represents the oil reservoir from which the hydrodynamic lubrica-ting wedge is supplied. The connec-tion of this groove 39 to the grooves 37, 38 via the thro-ttles 40, 41 ensures that a small oil reservoir is formed in the groove 39 in each spatial position of the bearing sec~men-t 31 so that when the direction of rotation is changed dry running of the bearing is prevented for a limi-ted time.
Finally a reservoir 42 is connected -to the groove 33 -to maintain a certain oil pressure in ~he event of fail-ure of the main oil supply.
The throttles provided in the bearinc3 segment as connections between the axial groove and the two cir-cumferential grooves constitute saEety means which prevent 3 ~
uncontrolled flowing off of the lubricatlng oil from t~e two circumferential grooves in the event of unfavourable clearances. I-t should be borne in mind that particularly when the rotating element is of large dimensions (such as in the mounting of large tube mills), ideal gap sizes can never be xelied upon and machining clearances, mech-anical deformations and deformations caused by temperature must be taken into account.
The feed pressure in the axial lubricant supply groove on -the inlet side of the rotating elemen-t should be between 3 and 30 bars for large bearing segments such as are used for the mounting of tube mills, drums and the like. The dimensions of the throttles leading to the cir-cumferential grooves are such that -the pressure loss occur-ring there is preferably between 10 and 20~.
In addition to the advantages already referredto there is a further advantage, namely tha-t the auxiliary starting means conventionally used in hydrodynamic bearings can be dispensed with (in large hydrodynamic bearings such auxiliary starting means conventionally consist of a central bore in -the bearing segmen-t by means of which the metal contact between the journal and the bearing is counterac-ted by high pressure oil before ro-tation).
A further advantage is tha-t because of the con-siderable reduction in -the lateral flow more oil can be passed through the bearing so that the temperature level in the bearing is lowered. ~s a result the oil remains cooler and thus its viscoslty and bearing capacity is higher, which improves the general bearing capacity of the bearing.
Bearings which are driven both hydrostatically and hydrodynamically are known. Both bearing systems have certain advantages and disadvantages.
Thus, as a rule the lubrication and inspection of a hydrostatic bearing is very costly, and -the emergency operating characteristics are poor. On the other hand, a hydrostatic bearing can be operated wi-th high surface loads, an approximately uniEorm surface load being pro-duced over the whole bearing surface. The average specific surface load thus approximately corresponds to -the maximum pressure on the bearing surface.
By contrast, the lubrication and inspection of a hydrodynamic bearing are considerably easier. Its emer-gency operating characteristics, tha-t is to say its be-haviour in the event of a sudden Eailure of lubrica-tion, are a great deal better than those of the hydrosta-tic bear-ing because of its smooth continuous bearing surface.
On the other hand, because of the way in which pressure builds up in the hydrodynamic wedge, enormous pressures occur in the so-called pressure peak in hydro-dynamic bearings or their supporting bearing segments.
The average specific surface loading of a hydrodynamic bearing is well below this peak pressure. As a result, a hydrodynamic bearing is of larger construction than a hydrostatic bearing with the same loading.
An object of the present inven-tion, therefore, is to provide a bearing segment for a hydrodynamic radial friction bearing which is distinguished by a higher average specific surface loading in comparison with known bearing constructions and permi-ts a par-ticularly narrow constructio ~-:
of the bearing.
The present invention provides a bearing segment for a hydrodynamic radial f:riction hearing adapted to sup-port a rotary member for rota-tion about an axis, the segment having an arcuate bearing surface adjacent one end of which is a lubricant supply ~roove extending axially of the seg-men-t, and means for delivering a lubricant into the axially extending supply groove, wherein -the bearing segment has a pair of lubricant distribution grooves in its bearing surface, one of which extends circumferentially of the segment adjacent one lateral edge thereof and the o-ther of which extends circumferentially of the se~ment adjacent its opposite lateral edge, the distribution grooves being in communication at corresponding ends thereof with the supply groove via passages through which lubricant may pass from the supply groove into both of the distribution grooves.
The German publication "VDI-BERICHTE", No. 111, 1966, pages 15 to 19, deals with the construction of radial friction bearings with a view to proper lubrication, and in particular page 19, illustration 15 IV, shows an embodi-ment in which a plurality of axially extending transverse grooves can be provided over the whole periphery of a radial friction bearing and a-t their axial ends the said grooves change directly into peripheral annular grooves. This means therefore that - quite apart from the fact that this is not a description of a bearing segment Eor such a radial friction bearing - the -transverse grooves are connected to the annular grooves provided at the axial bearing ends in such a way that the oil overflows from the transverse grooves into the annular grooves without any alterations in pressure. As a result of this known cons-truction, more oil can flow through the bearing in order to achieve better cooling, but an improvement in the bearing capacity and thus the economy of a hydrodynamic radial friction bearing cannot be achieved in this way.
A prior art bearing and two embodiments o~ the i3~
present invention are described below with reference to the accompanying drawings, wherein:-Figure 1 shows a plan view of a segment of a prior art hydrodynamic radial friction bearing;
Figure 2 shows a section along the line II in Figure l;
Figure 3 shows the pressure dlstribution along the periphery of the bearing segment in the plane II-II
of Figure 1 (i.e~ in the centre of the bearing);
Figures 4 and 5 show cross-sections through this pressure peak (along the lines IV-IV and V-V in Figure 3);
Figure 6 being a plan view of a bearing segment;
Figure 7 being a section along the line VII-VII
of Figure 6;
Figures 8, 9 and 10 are sections through the pressure build-up along the lines VIII-VIII, IX-IX and X-X, respectively;
Figure 11 shows a plan view of a bearing segment according to another embodiment of the invention; and Figure 12 shows a view taken in cross-section along the line XII-XII of Figure 11.
On the inlet side of the rotating element (not shown) in the bearing surface ? of this bearing segment 1 there is an axial lubricant supply groove 3 which is connected to a lubricating oil supply line 4 and serves to distribute the lubricating oil over the axial width of the bearing. This groove 3 serves in a known manner to supply the hydrodynamic wedge which forms between the bearing surface 2 and the rotating elemen-t during operation.
Depending upon the clearances, the mechanical stability of -the bearing segment 1 and the rotating element and the quantity of oil supplied, a more or less great sta-tic pressure builds up in the groove 3. From the ~uantity of oil present in the groove 3 the rotating ele-ment draws in sufficien-t into the gap to build the hydro-dynamic wedge. This "drawing in" is supported by the static 3~1~
pressure in the groove 3.
Some of the oil drawn in-to the gap be-tween -the bearing segment 1 and the rotating elemen-t now ~lows away from the bearing segment at the region of the longi-tudinal edges 2a, 2b of the bearing surface. This flow is un-desirable since it takes oil away from the centre of -the bearing which causes the pressure build-up to peak more, i.e. to be more uneven, and the height oE -the oil film to be reduced.
It is clear tha-t narrow bearings, i.e. bearings with a small axial dimension B, are particularly suscep-tible -to such a lateral flow.
Figures 6 to 10 show by contrast an embodiment of the invention.
As shown in Figure 6, the bearing segment 11 contains in its bearing surface 12 an axial lubrican-t supply groove 13 which is connected to a lubricating oil supply line 14~ The two ends of this groove 13 are connected via respective throttles or constrictions 15, 16 to cir-cumferentially e~tending grooves 17, 18 which extend axially of the bearing surface 12 close to the edges 12a, 12b.
Figures 8 to 10 show - in contrast -to Figures 3 to 5 - how the pressure build-up alters as a result of -the presence of the two circumferen-tial grooves 17, 18.
In the example shown in the drawing the oil pressure in the axial groove 3 is approximately 24 bars and in the circumferential grooves 17, 18 approxima-tely 20 bars.
The hydrodynamic lubrication wedge which forms in the gap between the bearing segment 11 and the rotating element during operation is consequently supported on its two flanks (i.e. in the region of the grooves 17, 18) against a higher pressure level than atmospheric pressure.
In this way flow past the edges 12a and 12b is consider-ably reduced and the volume of the pressure peak is in-creased approximately on the basis shown by the transverselines in Figures 8 to 10. This increase in -the volume oE the pressure peak corresponds to a propor-tional increase in the bearing capaci-ty of the bearing segment.
Thus it becomes possible either to increase the bearing capacity of the bearing ancl/or to make the beariny of narrower construction.
Referring now to Figures 11 and 12, which show a further embodiment of the invention, it will be seen that in this embodimen-t, in addition to the axial groove 33 provided on the inlet side of the rotating element and the two circumferential grooves 37, 38 connected via throttles 35, 36, the bearing segment 31 also contains in its bearing surface 32 an axial groove 39 provided on the outlet side of the rota-ting element and connected via throttles or constrictions 40, 41 to the longitudinal grooves 37, 38.
]5 The dimensions are advantageously such that -the throttles 35, 36 effect a pressure drop of 5 to 30~, pre-ferably 10 to 20%, in normal operation, whilst the -throttles 40, 41 effect a pressure drop of 70 to 95~, preferably 80 to 90~.
The construction according to Figures 11 and ]2 is particularly advantageous when the rotating element (for example a tube mill) opera-tes in only one direc-tion of rotation but oscillates in the other direc-tion of rota-tion when stopped. For this oscillation -the groove 39 represents the oil reservoir from which the hydrodynamic lubrica-ting wedge is supplied. The connec-tion of this groove 39 to the grooves 37, 38 via the thro-ttles 40, 41 ensures that a small oil reservoir is formed in the groove 39 in each spatial position of the bearing sec~men-t 31 so that when the direction of rotation is changed dry running of the bearing is prevented for a limi-ted time.
Finally a reservoir 42 is connected -to the groove 33 -to maintain a certain oil pressure in ~he event of fail-ure of the main oil supply.
The throttles provided in the bearinc3 segment as connections between the axial groove and the two cir-cumferential grooves constitute saEety means which prevent 3 ~
uncontrolled flowing off of the lubricatlng oil from t~e two circumferential grooves in the event of unfavourable clearances. I-t should be borne in mind that particularly when the rotating element is of large dimensions (such as in the mounting of large tube mills), ideal gap sizes can never be xelied upon and machining clearances, mech-anical deformations and deformations caused by temperature must be taken into account.
The feed pressure in the axial lubricant supply groove on -the inlet side of the rotating elemen-t should be between 3 and 30 bars for large bearing segments such as are used for the mounting of tube mills, drums and the like. The dimensions of the throttles leading to the cir-cumferential grooves are such that -the pressure loss occur-ring there is preferably between 10 and 20~.
In addition to the advantages already referredto there is a further advantage, namely tha-t the auxiliary starting means conventionally used in hydrodynamic bearings can be dispensed with (in large hydrodynamic bearings such auxiliary starting means conventionally consist of a central bore in -the bearing segmen-t by means of which the metal contact between the journal and the bearing is counterac-ted by high pressure oil before ro-tation).
A further advantage is tha-t because of the con-siderable reduction in -the lateral flow more oil can be passed through the bearing so that the temperature level in the bearing is lowered. ~s a result the oil remains cooler and thus its viscoslty and bearing capacity is higher, which improves the general bearing capacity of the bearing.
Claims (14)
1. A bearing segment for a hydrodynamic radial friction bearing adapted to support a rotary member for rotation about an axis, said segment having an arcuate bearing surface adjacent one end of which is a lubricant supply groove extending axially of said segment, and means for delivering a lubricant into said axially extending supply groove, wherein said bearing segment has a pair of lubricant distribution grooves in its bearing surface, one of which extends circumferentially of said segment adjacent one lateral edge thereof and the other of which extends circumferentially of said segment adjacent its opposite lateral edge, said distribution grooves being in communication at corresponding ends thereof with said supply groove via passages through which lubricant may pass from said supply groove into both of said distribu-tion grooves.
2. A bearing segment according to claim 1 wherein said segment has a further axially extending groove in said bearing surface adjacent its opposite end and parallel to said supply groove.
3. A bearing segment according to claim 2 wherein said further axially extending groove is in communication with said distribution grooves by passages extending be-tween said further groove and said distribution grooves and through which lubricant may pass.
4. A bearing segment according to claim 3 wherein the passages between said further groove and said distri-bution grooves are throttle passages.
5. A bearing segment according to claim 4 wherein said throttle passages are of such dimensions that in normal operation a pressure drop of between about 70% and 95%
occurs therein.
occurs therein.
6. A bearing segment according to claim 4 wherein said throttle passages are of such dimensions that in normal operation a pressure drop of between about 80% and 90%
occurs therein.
occurs therein.
7. A bearing segment according to claim 4, 5 or 6 wherein said throttle passages are bores.
8. A bearing segment according to claim 1 wherein said passages are throttle passages.
9. A bearing segment according to claim 8 wherein said passages are grooves in said bearing surface.
10. A bearing segment according to claim 9 wherein said throttle passages are of such dimensions that in normal operation a pressure drop of between about 5% and 30% occurs therein.
11. A bearing segment according to claim 9 wherein said throttle passages are of such dimensions that in normal operation a pressure drop of between about 10% and 20%
occurs therein.
occurs therein.
12. A bearing support according to claim 1, 2 or 3 including a lubricant reservoir in communication with said supply groove.
13. A bearing segment according to claim 1, 2 or 3 wherein the bearing surface of said bearing segment extends circumferentially beyond said supply groove and axially beyond said distribution grooves.
14. A bearing segment for a radial friction bearing having an axially extending lubricant supply groove in its bearing surface adjacent one end thereof and from which a lubricant may be distributed to said bearing sur-faces of said segment to provide a lubrication wedge, and means for supplying a lubricant to said supply groove, wherein said segment has in its bearing surface a pair of spaced, parallel grooves extending circumferentially of said segment from adjacent opposite ends of said supply groove, each of the circumferentially extending grooves communicating with said supply groove at opposite ends of the latter via a throttle groove in the bearing surface of said segment.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| DE19813117746 DE3117746A1 (en) | 1981-05-05 | 1981-05-05 | HYDRODYNAMIC RADIAL SLIDING BEARING |
| DEP3117746.8 | 1981-05-05 |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1186358A true CA1186358A (en) | 1985-04-30 |
Family
ID=6131493
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000399441A Expired CA1186358A (en) | 1981-05-05 | 1982-03-25 | Bearing segment for a hydrodynamic radial friction bearing |
Country Status (7)
| Country | Link |
|---|---|
| EP (1) | EP0064598B1 (en) |
| AT (1) | ATE11328T1 (en) |
| BR (1) | BR8202572A (en) |
| CA (1) | CA1186358A (en) |
| DE (2) | DE3117746A1 (en) |
| ES (1) | ES277883Y (en) |
| ZA (1) | ZA822102B (en) |
Cited By (4)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| ITCO20090032A1 (en) * | 2009-09-22 | 2011-03-23 | Nuovo Pignone Spa | BEARING, OIL AND METHOD DISTRIBUTION MECHANISM |
| US8734019B2 (en) | 2009-09-22 | 2014-05-27 | Nuovo Pignone S.P.A. | Bearing device, retention mechanism and method for retaining at least one pad |
| US10408258B2 (en) | 2014-01-24 | 2019-09-10 | Man Energy Solutions Se | Tilting segment for a shaft bearing device, and shaft bearing device |
| US10710131B2 (en) | 2015-05-26 | 2020-07-14 | Sms Group Gmbh | Roll arrangement |
Families Citing this family (7)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| SE442328B (en) * | 1983-09-29 | 1985-12-16 | Jan R Schnittger | HYDRODYNAMIC STORAGE UNIT |
| US5290899A (en) * | 1988-09-22 | 1994-03-01 | Tosoh Corporation | Photosensitive material having a silicon-containing polymer |
| BR9100852A (en) * | 1991-02-25 | 1992-10-27 | Metal Leve Sa | SLIDING BEARING |
| DE502005009860D1 (en) * | 2005-05-11 | 2010-08-19 | Freudenberg Carl Kg | bearings |
| JP6184299B2 (en) | 2013-11-08 | 2017-08-23 | 三菱日立パワーシステムズ株式会社 | Tilting pad type thrust bearing and rotating machine equipped with the same |
| DE102013224117A1 (en) | 2013-11-26 | 2015-05-28 | Sms Siemag Ag | Roller arrangement for rolls in a rolling mill |
| EP4170191A1 (en) * | 2021-10-20 | 2023-04-26 | Flender-Graffenstaden S.A.S. | Hydrodynamic bearing, gearbox and use |
Family Cites Families (8)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| GB371479A (en) * | 1931-01-28 | 1932-04-28 | James Lawrence Myers | Grooved bearings |
| CH274953A (en) * | 1949-07-06 | 1951-04-30 | Sulzer Ag | Bearings. |
| US3352610A (en) * | 1965-02-04 | 1967-11-14 | Furukawa Yuichiro | Bearing for impeller shaft |
| SE350312B (en) * | 1970-09-23 | 1972-10-23 | Skf Svenska Kullagerfab Ab | |
| DE2100365B2 (en) * | 1971-01-07 | 1981-09-17 | Ekkehard Dr.-Ing. 7100 Heilbronn Grau | Plain bearings with lubrication groove system |
| DE2211414A1 (en) * | 1972-03-06 | 1973-09-13 | Siemens Ag | HYDRODYNAMIC TRACK BEARING WITH CENTRALLY SUPPORTED TILTING SEGMENTS FOR A SHAFT ROTATING IN TWO DIRECTIONS |
| FR2229287A5 (en) * | 1973-05-11 | 1974-12-06 | Worthington France | Rotary fluid press. generating device - has shaft rotating about axis eccentric w.r.t. axis of cylinder |
| DE2931383A1 (en) * | 1979-08-02 | 1981-02-19 | Polysius Ag | HYDRODYNAMIC BEARING |
-
1981
- 1981-05-05 DE DE19813117746 patent/DE3117746A1/en not_active Withdrawn
-
1982
- 1982-03-25 CA CA000399441A patent/CA1186358A/en not_active Expired
- 1982-03-29 ZA ZA822102A patent/ZA822102B/en unknown
- 1982-04-02 DE DE8282102838T patent/DE3261910D1/en not_active Expired
- 1982-04-02 AT AT82102838T patent/ATE11328T1/en not_active IP Right Cessation
- 1982-04-02 EP EP82102838A patent/EP0064598B1/en not_active Expired
- 1982-05-04 ES ES1982277883U patent/ES277883Y/en not_active Expired
- 1982-05-04 BR BR8202572A patent/BR8202572A/en unknown
Cited By (7)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| ITCO20090032A1 (en) * | 2009-09-22 | 2011-03-23 | Nuovo Pignone Spa | BEARING, OIL AND METHOD DISTRIBUTION MECHANISM |
| EP2302240A1 (en) * | 2009-09-22 | 2011-03-30 | Nuovo Pignone S.p.A. | Bearing device, oil distribution mechanism and oil distribution method |
| EP2543898A1 (en) * | 2009-09-22 | 2013-01-09 | Nuovo Pignone S.p.A. | Bearing pad, bearing device and oil distribution method |
| US8657501B2 (en) | 2009-09-22 | 2014-02-25 | Nuovo Pignone S.P.A. | Bearing device, oil distribution mechanism and method |
| US8734019B2 (en) | 2009-09-22 | 2014-05-27 | Nuovo Pignone S.P.A. | Bearing device, retention mechanism and method for retaining at least one pad |
| US10408258B2 (en) | 2014-01-24 | 2019-09-10 | Man Energy Solutions Se | Tilting segment for a shaft bearing device, and shaft bearing device |
| US10710131B2 (en) | 2015-05-26 | 2020-07-14 | Sms Group Gmbh | Roll arrangement |
Also Published As
| Publication number | Publication date |
|---|---|
| EP0064598B1 (en) | 1985-01-16 |
| EP0064598A1 (en) | 1982-11-17 |
| ATE11328T1 (en) | 1985-02-15 |
| ES277883U (en) | 1984-12-16 |
| ZA822102B (en) | 1983-03-30 |
| BR8202572A (en) | 1983-04-19 |
| DE3117746A1 (en) | 1982-12-09 |
| ES277883Y (en) | 1985-06-16 |
| DE3261910D1 (en) | 1985-02-28 |
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Legal Events
| Date | Code | Title | Description |
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| MKEX | Expiry |