JPH0276923A - Static pressure gas bearing - Google Patents
Static pressure gas bearingInfo
- Publication number
- JPH0276923A JPH0276923A JP157288A JP157288A JPH0276923A JP H0276923 A JPH0276923 A JP H0276923A JP 157288 A JP157288 A JP 157288A JP 157288 A JP157288 A JP 157288A JP H0276923 A JPH0276923 A JP H0276923A
- Authority
- JP
- Japan
- Prior art keywords
- bearing
- air supply
- groove
- shaft
- pressure
- 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.)
- Granted
Links
- 230000003068 static effect Effects 0.000 title claims description 11
- 238000013016 damping Methods 0.000 abstract description 12
- 230000000694 effects Effects 0.000 abstract description 4
- 230000001050 lubricating effect Effects 0.000 abstract description 2
- 230000002706 hydrostatic effect Effects 0.000 description 19
- 230000008859 change Effects 0.000 description 11
- 238000009826 distribution Methods 0.000 description 9
- 230000009471 action Effects 0.000 description 6
- 238000004364 calculation method Methods 0.000 description 4
- 239000000314 lubricant Substances 0.000 description 2
- 238000004891 communication Methods 0.000 description 1
- 230000000052 comparative effect Effects 0.000 description 1
- 238000005516 engineering process Methods 0.000 description 1
- 239000012530 fluid Substances 0.000 description 1
- 239000003292 glue Substances 0.000 description 1
- 238000005461 lubrication Methods 0.000 description 1
- 238000004519 manufacturing process Methods 0.000 description 1
- 238000000034 method Methods 0.000 description 1
- 239000011148 porous material Substances 0.000 description 1
- 230000008569 process Effects 0.000 description 1
- 230000004044 response Effects 0.000 description 1
Landscapes
- Magnetic Bearings And Hydrostatic Bearings (AREA)
Abstract
Description
【発明の詳細な説明】
〔産業上の利用分野〕
この発明は、給気孔から軸と軸受面間に形成された軸受
すきまに圧縮気体を供給して軸を支持するようにした静
圧気体軸受に関するものである。[Detailed Description of the Invention] [Field of Industrial Application] This invention provides a hydrostatic gas bearing that supports a shaft by supplying compressed gas from an air supply hole to a bearing gap formed between a shaft and a bearing surface. It is related to.
静圧気体軸受は、軸受すきまに形成される圧縮気体膜に
よって軸を支持するため、回転精度が高いこと、軸受部
の摩耗損失が小さいこと、さらに、スティックスリップ
を生じないこと等の利点を有する。このため、精密加工
機、精密測定器等の回転軸の軸受として広く用いられて
いる。Hydrostatic gas bearings support the shaft with a compressed gas film formed in the bearing clearance, so they have advantages such as high rotational accuracy, low wear loss on the bearing, and no stick-slip. . For this reason, they are widely used as bearings for rotating shafts in precision processing machines, precision measuring instruments, etc.
しかし、気体による流体潤滑であるため、他の軸受に比
較して軸受剛性が小さいという欠点がある。However, since it uses fluid lubrication using gas, it has the disadvantage that the bearing rigidity is lower than that of other bearings.
したがって、軸受剛性の向上は、静圧気体軸受の重要な
課題の一つになっている。Therefore, improving bearing rigidity is one of the important issues for hydrostatic gas bearings.
その課題解決のため、特公昭５８−１４９２７号公軸に
示された静圧気体軸受においては、第１２図に示すよう
に、軸２１を支持する軸受部材２２の軸受面２３に、一
対の周方向溝２４．２４を軸方向に所要の間隔をおいて
設け、各周方向溝２４のまわりには、周方向に所要の間
隔をおいて給気孔２５を形成し、各給気孔２５の先端の
ノズル２６を同方向溝２４に連通させた構成としである
。In order to solve this problem, in the hydrostatic gas bearing shown in the public shaft of Japanese Patent Publication No. 58-14927, as shown in FIG. Directional grooves 24, 24 are provided at required intervals in the axial direction, air supply holes 25 are formed around each circumferential groove 24 at required intervals in the circumferential direction, and the tip of each air supply hole 25 is The nozzle 26 is configured to communicate with the groove 24 in the same direction.
一般に、静圧気体軸受では、第１３図に示すように、軸
に荷重Ｗが加わると、軸２１が偏心し、それに伴って、
給気孔２５の絞り作用により、軸受面２３の圧力分布が
実線イから点線口のように変化し、軸受面２３の荷重方
向と圧荷重方向で差圧が生じる。これが軸受圧力となり
、荷重Ｗとつり合う。Generally, in a static pressure gas bearing, as shown in FIG. 13, when a load W is applied to the shaft, the shaft 21 becomes eccentric, and accordingly,
Due to the throttling action of the air supply hole 25, the pressure distribution on the bearing surface 23 changes from the solid line A to the dotted line A, and a pressure difference is generated between the load direction and the pressure load direction of the bearing surface 23. This becomes the bearing pressure and balances the load W.
軸受の剛性は、軸受圧力／軸の偏心量で定義される。The rigidity of a bearing is defined as bearing pressure/shaft eccentricity.
ところで、第１２図に示す静圧気体軸受においては、軸
２１の偏心によって、第１３図の点線口のように軸受面
２３に差圧が生じると、荷重方向の圧力が周方向溝２４
を流れて圧荷重方向に移動する。By the way, in the static pressure gas bearing shown in FIG. 12, when a differential pressure is generated on the bearing surface 23 as shown by the dotted line in FIG.
flows in the direction of the pressure load.
このため、差圧が小さくなり、軸受剛性が低いという問
題がある。Therefore, there is a problem that the differential pressure is small and the bearing rigidity is low.
そこで、周方向への圧力の流れを減少させるため、特開
昭５６−１３４６２３号公報に示された静圧気体軸受に
おいては、第１４図に示すように、軸受部材３２の軸受
面３３に、軸方向溝３４とその両端に周方向溝３５を連
通させたＨ形給気溝３６を周方向に等間隔に設けると共
に、給気溝列を軸方向に２列設け、各給気溝３６におけ
る軸方向溝３４の中央に、給気孔３７の先端のノズル３
８を連通させた構成としである。Therefore, in order to reduce the flow of pressure in the circumferential direction, in the hydrostatic gas bearing disclosed in Japanese Patent Application Laid-Open No. 56-134623, as shown in FIG. H-shaped air supply grooves 36 in which the axial groove 34 and the circumferential groove 35 communicate with each other are provided at equal intervals in the circumferential direction, and two rows of air supply grooves are provided in the axial direction. The nozzle 3 at the tip of the air supply hole 37 is located in the center of the axial groove 34.
8 is connected to each other.
ところで、上記の静圧気体軸受においては、給気溝３６
の容積が比較的大きくなるので、給気孔３７の絞り作用
による給気孔出口部の圧力変動が、軸３１の単位偏心量
に対して最大になるように軸受すきま３９を設定すると
、減衰係数が小さくなり、自励振動が発生する危険性が
高くなる。一方、減衰係数を大きくするため、軸受すき
ま３９を小さくすると、給気孔３７の絞り作用が低下し
、軸３１が偏心したときの給気孔出口部の圧力変動が小
さくなり、剛性が低下する。By the way, in the above static pressure gas bearing, the air supply groove 36
Since the volume of the air supply hole 37 is relatively large, if the bearing clearance 39 is set so that the pressure fluctuation at the outlet of the air supply hole due to the throttling action of the air supply hole 37 is maximized relative to the unit eccentricity of the shaft 31, the damping coefficient will be small. This increases the risk of self-excited vibration occurring. On the other hand, if the bearing clearance 39 is made smaller in order to increase the damping coefficient, the throttling effect of the air supply hole 37 will be reduced, the pressure fluctuation at the air supply hole outlet will be reduced when the shaft 31 is eccentric, and the rigidity will be reduced.
このとき、この小さい軸受すきま３９に対して給気溝３
６の深さを適切に決めれば、軸受端側の周方向溝３５部
では、軸３１の偏心に対し表面絞り作用により、有効な
圧力変動を生じるようにすることができるが、給気孔３
７列より内側の周方向溝３５にはさまれた軸受中央部で
は、軸３１の偏心に対し、圧力は、給気孔３７出ロ部と
ほぼ等しく保持されるため、圧力変動も給気孔３７出ロ
部と同様に圧力変動は少ない、したがって、この軸受ば
、軸受全体の面積を十分有効に利用することができない
。At this time, the air supply groove 3
If the depth of the air supply hole 3 is determined appropriately, effective pressure fluctuations can be generated in the circumferential groove 35 on the bearing end side due to the surface throttling action against the eccentricity of the shaft 31.
In the central part of the bearing sandwiched between the circumferential grooves 35 on the inner side of the seventh row, the pressure is maintained approximately equal to that at the outer part of the air supply hole 37 in response to eccentricity of the shaft 31, so pressure fluctuations also occur at the outer part of the air supply hole 37. As with the lower part, pressure fluctuations are small, so with this bearing, the area of the entire bearing cannot be used effectively enough.
第１５図は、第１４図に示す静圧気体軸受の軸受すきま
を小さくした場合の周方向の圧力分布を示し、第１６図
に示す展開図の軸受幅の中央部１ｆ（Ｚ＝Ｏ）、給気孔
３７の位置＜Ｚ−Ｚｓ ） 、周方向溝３５の溝幅中央
位置（Ｚ＝Ｚｇ）の３ケ所について示したものである。FIG. 15 shows the pressure distribution in the circumferential direction when the bearing clearance of the static pressure gas bearing shown in FIG. Three locations are shown: the position of the air supply hole 37<Z-Zs), and the center position of the groove width of the circumferential groove 35 (Z=Zg).
また、実線は、負荷のない場合、点線は、負荷が加えら
れ、軸３１が偏心した場合に対応する。これにより、Ｚ
−０（軸受中央部）での圧力変化が少ないことが分る。Further, the solid line corresponds to the case where no load is applied, and the dotted line corresponds to the case where a load is applied and the shaft 31 is eccentric. This allows Z
It can be seen that there is little pressure change at -0 (center of the bearing).
この発明は、上記の欠点を解決し、静圧気体軸受の軸受
剛性および減衰係数を高めることを技術的ｉ１題として
いる。The technical objective of the present invention is to solve the above-mentioned drawbacks and increase the bearing rigidity and damping coefficient of a hydrostatic gas bearing.
上記の課題を解決するために、この発明の静圧気体軸受
においては、軸の外径面とこれを支持する軸受部材の軸
受面におけるいずれか一方の面に、周方向に等間隔に並
ぶ複数の給気溝を、上記軸受面両端に近接して２列設け
、各給気溝は、軸受面端に近接して設けられた周方向溝
と、その周方向溝から軸受面中央部に向く短かい長さの
軸方向溝とから形成され、各給気溝の軸方向溝に給気孔
を連通させた構成としである。In order to solve the above problems, in the hydrostatic gas bearing of the present invention, a plurality of Two rows of air supply grooves are provided close to both ends of the bearing surface, and each air supply groove includes a circumferential groove provided close to the end of the bearing surface and a direction from the circumferential groove toward the center of the bearing surface. The air supply hole is formed of short axial grooves, and the air supply hole is communicated with the axial groove of each air supply groove.
上記のように構成すれば、周方向に並ぶ複数の給気溝ば
互に独立しているため、荷重方向の圧力が周方向溝を流
れて圧荷重側に移動し、圧力差が小さくなるということ
がない、また、給気溝列間における軸受中央部に広いラ
ンド部を設けることができ、小さい軸受すきまに対して
給気溝の深さを適切に決めれば、軸の偏心時、周方向溝
の圧力が、表面絞り作用によって大きく変化したのと同
様に、比較的大きな圧力変化をランド部にも生じさせる
ことができる。したがって、大きく圧力変化する部分の
、軸受面における割合を従来の空気軸受に比較し増加さ
せうるため、軸受剛性を高めることができる。With the above configuration, the multiple air supply grooves lined up in the circumferential direction are independent of each other, so the pressure in the load direction flows through the circumferential grooves and moves to the pressure load side, reducing the pressure difference. In addition, a wide land can be provided at the center of the bearing between the rows of air supply grooves, and if the depth of the air supply groove is determined appropriately for a small bearing clearance, when the shaft is eccentric, In the same way that the pressure in the groove changes significantly due to the surface throttling action, a relatively large pressure change can also be caused in the land portion. Therefore, the proportion of the bearing surface where the pressure changes greatly can be increased compared to conventional air bearings, so the bearing rigidity can be increased.
さらに、軸受すきまが小さい条件で軸受剛性を向上させ
て使用することができるため、及び給気溝の容積が小さ
いため、減衰係数も向上し、安定な軸受を形成すること
ができる。Furthermore, since the bearing rigidity can be improved and used under conditions where the bearing clearance is small and the volume of the air supply groove is small, the damping coefficient is also improved and a stable bearing can be formed.
以下、この発明の実施例を第１図乃至第１１図に基づい
て説明する。Embodiments of the present invention will be described below with reference to FIGS. 1 to 11.
第１図に示すように、軸１の外径面と、その軸１を支持
する軸受部材２の軸受面３間には、軸受すきま４が設け
られている。As shown in FIG. 1, a bearing clearance 4 is provided between the outer diameter surface of the shaft 1 and the bearing surface 3 of the bearing member 2 that supports the shaft 1.
上記軸受面３には、周方向に等間隔に並ぶ複数の給気溝
５が、軸受面中央を対称線とする対称位置で、軸受面端
に近接して設けられている。この給気溝５は、短かい軸
方向溝６の軸受外側端に同じ深さの周方向溝７を連通さ
せたＴ形をなし、その軸方向溝６の軸受内側端部に、給
気孔８の先端のノズル９が連通している。In the bearing surface 3, a plurality of air supply grooves 5 arranged at equal intervals in the circumferential direction are provided close to the end of the bearing surface at symmetrical positions with the center of the bearing surface as a line of symmetry. The air supply groove 5 has a T-shape in which the outer end of the short axial groove 6 communicates with the circumferential groove 7 of the same depth, and the inner end of the axial groove 6 has an air supply hole 8. The nozzle 9 at the tip is in communication.
いま、給気孔８に圧縮気体を供給すると、その圧縮気体
は、ノズル９から軸受すきま４に流れ、この軸受すきま
４において気体潤滑膜が形成される。その気体潤滑膜に
より、軸１は、軸受面３に対して非接触の状態に保持さ
れる。Now, when compressed gas is supplied to the air supply hole 8, the compressed gas flows from the nozzle 9 to the bearing gap 4, and a gas lubricant film is formed in the bearing gap 4. The gas lubricating film keeps the shaft 1 out of contact with the bearing surface 3.
第２図は、上記気体潤滑膜の周方向圧力分布を、第３図
に示す展開図の軸受幅の中央部！ （Ｚ−〇）、給気孔
８の位置（Ｚ＝１ ） 、周方向溝７の溝幅中央部？１
ｆ （ｚ＝ｚ、）の３ケ所について示したものであり、
実線は負荷のない場合、点線は負荷が加えられ、軸１が
偏心した場合に対応する。Figure 2 shows the circumferential pressure distribution of the gas lubricant film at the center of the bearing width in the developed view shown in Figure 3! (Z-〇), the position of the air supply hole 8 (Z=1), the center part of the groove width of the circumferential groove 7? 1
It shows three locations of f (z=z,),
The solid line corresponds to the case without a load, and the dotted line corresponds to the case when a load is applied and the shaft 1 is eccentric.
この第２図は、第１図に示す静圧気体軸受の軸１の外径
りを６０龍、軸受面３の幅寸法Ｗを４０１１給気孔８の
ノズル９の直径ｄを０．３ｍｍ、給気孔８の個数６ケ×
２列、給気溝５の溝幅ａを０．７１ｍ、給気溝５の溝深
さｂを２０ＰＭ、軸受すきま４を５−とした場合の計算
結果である。2, the outer diameter of the shaft 1 of the hydrostatic gas bearing shown in FIG. Number of pores 8: 6 x
These are calculation results when the groove width a of the air supply groove 5 is 0.71 m, the groove depth b of the air supply groove 5 is 20 PM, and the bearing clearance 4 is 5-.
この第２図から、Ｚ＝０ （軸受中央部）の位置の圧力
変化が、前述の第１５図で示すＺ＝０（軸受中央部）に
おける圧力変化に対して増加していることが分る。From this Figure 2, it can be seen that the pressure change at the position Z = 0 (bearing center) increases compared to the pressure change at Z = 0 (bearing center) shown in Figure 15 above. .
第４図ａ、第５図ａ、第６図ａは、第１図に示す実施例
の静圧気体軸受（本発明品）について、無負荷時の圧力
、軸１の偏心１μあたりの圧力変化（軸受の剛性）、お
よび軸１の偏心速度Ｉｎ／Ｓあたりの圧力変化（軸受の
減衰係数）の軸方向の分布を、第７図ａに示す展開図の
給気孔８の位置（θ＝０）、周方向溝７の先端部（θ＝
０１）、隣接する周方向溝７．７間の中央位置（θ−θ
、）の３ケ所について示す、各部の寸法は、前記と同じ
である。Figure 4a, Figure 5a, and Figure 6a show the pressure at no load and the pressure change per 1μ eccentricity of shaft 1 for the static pressure gas bearing (product of the present invention) of the example shown in Figure 1. (rigidity of the bearing) and pressure change (damping coefficient of the bearing) per eccentric speed In/S of the shaft 1 are calculated at the position of the air supply hole 8 (θ=0 ), the tip of the circumferential groove 7 (θ=
01), the center position between adjacent circumferential grooves 7.7 (θ-θ
, ) The dimensions of each part shown in the three locations are the same as above.
その比較として、第１４図に示す静圧気体軸受（比較品
）の軸方向分布を、第４図ｂ、第５図ｂ、および第６図
すに示す、第４図す乃至第６図すは、第７図すに示す展
開図の給気孔３７の位置（θ−〇）、周方向溝３５の先
端部（θ＝θ、）、隣接する周方向溝３５．３５間の中
央位置（θ＝０２）の３ケ所について示し、軸の外径、
軸受面の幅寸法、給気孔の直径、給気孔の個数、給気孔
の溝幅、溝深さおよび軸受すきまについては、本発明に
係る静圧気体軸受の各部の寸法と同じとしである。For comparison, the axial distribution of the static pressure gas bearing (comparison product) shown in Fig. 14 is shown in Figs. are the position of the air supply hole 37 (θ-〇), the tip of the circumferential groove 35 (θ=θ,), and the center position between the adjacent circumferential grooves 35 and 35 (θ =02), the outer diameter of the shaft,
The width dimension of the bearing surface, the diameter of the air supply hole, the number of air supply holes, the groove width and groove depth of the air supply hole, and the bearing clearance are the same as the dimensions of each part of the hydrostatic gas bearing according to the present invention.
第４図ａから分かるように、この発明に係る静圧気体軸
受では、軸受の給気孔８の位置より内側に溝が無いため
、周方向溝７より内側では、給気孔８の近傍をのぞき、
圧力は、周方向溝７部の圧力とほぼ等しい、このため、
負荷が加わり、軸１が偏心しても、周方向溝７の圧力が
、表面絞り作用によって大きく変化すると、第２図の点
線および第５図ａから分るように、周方向溝７より内側
の給気孔８の近傍をのぞく領域でも比較的大きな圧力変
化が生じることになり、第１４図の静圧気体軸受に比較
して、特に、軸受の中央部において大きな圧力変化が得
られる。つまり、より広い面積にわたって圧力が大きく
変化するために、大きな剛性が得られる。As can be seen from FIG. 4a, in the hydrostatic gas bearing according to the present invention, since there is no groove inside the position of the air supply hole 8 of the bearing, inside the circumferential groove 7, except for the vicinity of the air supply hole 8,
The pressure is approximately equal to the pressure in the circumferential groove 7, therefore,
Even if a load is applied and the shaft 1 is eccentric, the pressure in the circumferential groove 7 changes greatly due to the surface throttling action, and as can be seen from the dotted line in FIG. 2 and FIG. A relatively large pressure change occurs in the region other than the vicinity of the air supply hole 8, and compared to the static pressure gas bearing shown in FIG. 14, a particularly large pressure change is obtained in the central portion of the bearing. In other words, greater stiffness is achieved because the pressure varies more widely over a larger area.
また、２列の給気孔８の列の中間には溝部がない広いラ
ンド部１０を作ることができ、かつ使用しうる軸受すき
まが小さいため、第６図ａで示すように、減衰係数も大
幅に改善される。Furthermore, since a wide land portion 10 without a groove can be created between the two rows of air supply holes 8, and the usable bearing clearance is small, the damping coefficient can be significantly increased as shown in Figure 6a. will be improved.
第８図乃至第１Ｏ図は、第１図に示す静圧気体軸受（本
発明品）と、第１４図に示す静圧気体軸受（比較品）の
性能計算の結果を示す、ここで、静圧気体軸受の各部の
寸法は、先に述べたものと同様であり、第８図は軸受剛
性、第９図は減衰係数、第１０図は消費２ｉｉ量である
。各種の静圧気体軸受において、剛性が最大になるよう
に、それぞれの軸受すきまを定めた場合、この発明に係
る静圧気体軸受は、他の静圧気体軸受より高い剛性が得
られると共に、減衰係数も大幅に向上し、また、軸受ず
きまが小さいので、気体の消費流量も減少することが分
る。8 to 1O show the results of performance calculations of the static pressure gas bearing shown in FIG. 1 (product of the present invention) and the static pressure gas bearing (comparative product) shown in FIG. The dimensions of each part of the pressure gas bearing are the same as those described above, and FIG. 8 shows the bearing rigidity, FIG. 9 shows the damping coefficient, and FIG. 10 shows the consumption 2ii. In various types of hydrostatic gas bearings, if the bearing clearance is determined so that the rigidity is maximized, the hydrostatic gas bearing according to the present invention has higher rigidity than other hydrostatic gas bearings, and has less damping. It can be seen that the coefficient is significantly improved, and since the bearing clearance is small, the gas consumption flow rate is also reduced.
さらに、この発明におけるランド部の影響を検討した結
果を第１７図に示す。Further, FIG. 17 shows the results of examining the influence of the land portion in this invention.
第１７図は第１図に示す静圧気体軸受において周方向溝
７の位置を軸受端近傍に固定し、軸方向溝６の長さを、
その内側端部にある給気孔８とともに変化させた場合の
剛性の計算結果である。FIG. 17 shows that in the hydrostatic gas bearing shown in FIG. 1, the position of the circumferential groove 7 is fixed near the bearing end, and the length of the axial groove 6 is
This is the calculation result of the rigidity when the rigidity is changed together with the air supply hole 8 located at the inner end.
軸受は軸外径６０曽厘、軸受面の幅４０鰭、オリフィス
孔径０，１＋ｎ、給気孔の個数６ケ×２列、給気溝の幅
０．７１ｊ、給気溝の深さ２０Ｉ１ｍ、周方向溝間の軸
方向間隔３０１ｕとした時の値であり、グラフ中のり、
は第１図における給気孔間の軸方向の距離を示す。第１
７図から明らかなように、軸方向溝６が短くなり給気孔
８が外側に移動するほど内側のランド部分が広がり、こ
の軸受の剛性は向上する。The bearing has a shaft outer diameter of 60 cm, a width of the bearing surface of 40 mm, an orifice hole diameter of 0,1+n, a number of air supply holes of 6 x 2 rows, a width of the air supply groove of 0.71J, a depth of the air supply groove of 20I1m, and a circumference of This is the value when the axial distance between the direction grooves is 301u, and the glue in the graph is
indicates the axial distance between the air supply holes in FIG. 1st
As is clear from FIG. 7, as the axial groove 6 becomes shorter and the air supply hole 8 moves outward, the inner land portion becomes wider and the rigidity of this bearing is improved.
ところで、軸受を実際に製作するにあたっては、給気孔
８を軸受外側端部に極端に近づけて加工するのは不可能
である。By the way, when actually manufacturing a bearing, it is impossible to process the air supply hole 8 extremely close to the outer end of the bearing.
また、この発明の軸受においては、周方向溝７も一１＝
記と同様、内側のランド部を広げて剛性を上げるという
理由から外側端部へ設ける方が望ましい。Further, in the bearing of the present invention, the circumferential groove 7 is also
As in the case described above, it is preferable to provide it at the outer end for the reason that the inner land portion is widened and the rigidity is increased.
以上より、設計に当っては、給気孔８が容易に加工しう
る範囲内で外側端部よりに給気孔８を設け、さらに、そ
の孔から外側端部よりに周方向溝７を設け、この周方向
溝７と給気孔８とを短い軸方向溝６でつないだ形に設計
すると良い。Based on the above, when designing the air supply hole 8, the air supply hole 8 is provided closer to the outer end within a range where the air supply hole 8 can be easily machined, and the circumferential groove 7 is further provided from the hole to the outer end. It is preferable to design the circumferential groove 7 and the air supply hole 8 to be connected by a short axial groove 6.
第】１図は、この発明に係る静圧軸受の他の実施例を示
す、この静圧気体軸受は、軸１に筒体１１を嵌着し、そ
の筒体１１の外径面に、周方向に等間隔に配！した複数
のＴ影絵気溝５′を軸方向に２列に設け、各給気溝５′
の軸方向６′に給気孔８′の先端のノズル９′を連通せ
しめ、軸１には、給気孔８′に対する圧縮空気の給気路
１２を設けた構成としたのである。FIG. 1 shows another embodiment of the hydrostatic bearing according to the present invention. In this hydrostatic gas bearing, a cylindrical body 11 is fitted onto a shaft 1, and the outer diameter surface of the cylindrical body 11 has a circumferential surface. Arranged at equal intervals in the direction! A plurality of T silhouette air grooves 5' are provided in two rows in the axial direction, and each air supply groove 5'
The nozzle 9' at the tip of the air supply hole 8' is communicated with the air supply hole 8' in the axial direction 6', and the shaft 1 is provided with an air supply passage 12 for supplying compressed air to the air supply hole 8'.
以上のように、この発明によれば、軸受外側端部におけ
る周方向溝の表面絞り作用による圧力変化と軸受中央部
におけるランド部の圧力変化とを有効に利用することが
できるので、軸受剛性を大きく向上させると共に、溝部
容積も小さいため減衰係数も大きく、安定な軸受を形成
することができる。As described above, according to the present invention, it is possible to effectively utilize the pressure change due to the surface squeezing action of the circumferential groove at the outer end of the bearing and the pressure change of the land portion at the center of the bearing, thereby increasing the bearing rigidity. This is greatly improved, and since the groove volume is small, the damping coefficient is also large, making it possible to form a stable bearing.
また、軸受すきまを小さくできるので、空気の消費流量
を減少させることもできる。Furthermore, since the bearing clearance can be reduced, the air consumption flow rate can also be reduced.
第１図は、この発明に係る静圧気体軸受の一実施例を示
す縦断正面図、第２図は上記軸受の開方向圧力分布を示
すグラフ、第３図は同上軸受面の展開図、第４図ａ、ｂ
、第５図ａ、ｂおよび第６図ａ、ｂは、本発明に係る静
圧気体軸受と第１４図に示す静圧気体軸受の軸方向圧力
分布、軸偏心量１ｎあたりの圧力変化の軸方向分布およ
び軸偏心速度１μ／Ｓあたりの圧力変化の軸方向分布を
示すグラフ、第７図ａ、ｂは、本発明の軸受と従来の軸
受の軸受面の展開図、第８図乃至第１０図は、本発明の
軸受と従来の軸受の、剛性、減衰係数および消費流量の
比較を示すグラフ、第１１図は、この発明に係る軸受の
他の実施例を示す縦断正面図、第１２図は、従来の軸受
を示す縦断正面図、第Ｉ３図は同上の縦断側面図、第１
４図は従来の軸受の縦断正面図、第１５図は第１４図に
示す軸受の周方向の圧力分布を示すグラフ、第１６図は
第１４図に示す軸受の軸受面の展開図、第１７図は、こ
の発明の軸受において、給気孔位置を変化させたときの
軸受剛性の計算結果である。
１・・・・・・軸、 ２・・・・・・軸受部
材、５・・・・・・給気溝、 ６・・・・・・軸
方向溝、７・・・・・・周方向溝、 ８・・・・・
・給気孔、９・・・・・・ノズル。FIG. 1 is a longitudinal sectional front view showing an embodiment of the hydrostatic gas bearing according to the present invention, FIG. 2 is a graph showing the pressure distribution in the opening direction of the bearing, FIG. 3 is a developed view of the bearing surface, and FIG. Figure 4 a, b
, Fig. 5 a, b, and Fig. 6 a, b show the axial pressure distribution and axis of pressure change per shaft eccentricity 1n of the hydrostatic gas bearing according to the present invention and the hydrostatic gas bearing shown in Fig. 14. Graphs showing the directional distribution and the axial distribution of pressure changes per shaft eccentric speed of 1 μ/S, FIGS. 11 is a graph showing a comparison of rigidity, damping coefficient, and flow rate consumption between the bearing of the present invention and a conventional bearing. FIG. 11 is a longitudinal sectional front view showing another embodiment of the bearing according to the present invention. FIG. is a longitudinal sectional front view showing a conventional bearing, FIG. I3 is a longitudinal sectional side view of the same, and
4 is a longitudinal sectional front view of a conventional bearing, FIG. 15 is a graph showing the pressure distribution in the circumferential direction of the bearing shown in FIG. 14, FIG. 16 is a developed view of the bearing surface of the bearing shown in FIG. 14, and FIG. The figure shows the calculation results of the bearing rigidity when the air supply hole position is changed in the bearing of the present invention. 1... Shaft, 2... Bearing member, 5... Air supply groove, 6... Axial groove, 7... Circumferential direction Groove, 8...
・Air supply hole, 9...Nozzle.
Claims (1)
いずれか一方の面に、周方向に等間隔に並ぶ複数の給気
溝を、上記軸受面両端に近接して２列設け、各給気溝は
、軸受面端に近接して設けられた周方向溝と、その周方
向溝から軸受面中央部に向く短かい長さの軸方向溝とか
ら形成され、各給気溝の軸方向溝に給気孔を連通させた
静圧気体軸受。A plurality of air supply grooves arranged at equal intervals in the circumferential direction are provided in two rows close to both ends of the bearing surface on either the outer diameter surface of the shaft or the bearing surface of the bearing member that supports the shaft. The air supply groove is formed of a circumferential groove provided close to the end of the bearing surface and a short axial groove extending from the circumferential groove toward the center of the bearing surface. A static pressure gas bearing with an air supply hole communicating with the directional groove.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP63001572A JP2644247B2 (en) | 1988-01-07 | 1988-01-07 | Hydrostatic gas bearing |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP63001572A JP2644247B2 (en) | 1988-01-07 | 1988-01-07 | Hydrostatic gas bearing |
Related Child Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP22957388A Division JPH0280324A (en) | 1988-09-13 | 1988-09-13 | Method for synthesizing chevrel phase compound |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPH0276923A true JPH0276923A (en) | 1990-03-16 |
| JP2644247B2 JP2644247B2 (en) | 1997-08-25 |
Family
ID=11505236
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP63001572A Expired - Fee Related JP2644247B2 (en) | 1988-01-07 | 1988-01-07 | Hydrostatic gas bearing |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JP2644247B2 (en) |
Cited By (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US8308464B2 (en) | 2006-12-22 | 2012-11-13 | Kobe Steel, Ltd. | Bearing and liquid cooling type screw compressor |
| CN106644475A (en) * | 2016-10-12 | 2017-05-10 | 河南科技大学 | Static and dynamic characteristics and stability test method for micro gas bearing |
Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5910521U (en) * | 1982-07-13 | 1984-01-23 | 日本精工株式会社 | Orifice type hydrostatic bearing |
-
1988
- 1988-01-07 JP JP63001572A patent/JP2644247B2/en not_active Expired - Fee Related
Patent Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5910521U (en) * | 1982-07-13 | 1984-01-23 | 日本精工株式会社 | Orifice type hydrostatic bearing |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US8308464B2 (en) | 2006-12-22 | 2012-11-13 | Kobe Steel, Ltd. | Bearing and liquid cooling type screw compressor |
| CN106644475A (en) * | 2016-10-12 | 2017-05-10 | 河南科技大学 | Static and dynamic characteristics and stability test method for micro gas bearing |
| CN106644475B (en) * | 2016-10-12 | 2018-12-14 | 河南科技大学 | A kind of minitype gas bearing Static and dynamic performance and stability test method |
Also Published As
| Publication number | Publication date |
|---|---|
| JP2644247B2 (en) | 1997-08-25 |
Similar Documents
| Publication | Publication Date | Title |
|---|---|---|
| US8646979B2 (en) | Hybrid hydro (air) static multi-recess journal bearing | |
| US4448460A (en) | Hydrostatic bearing | |
| US7845855B2 (en) | Integral tilting pad bearing | |
| US4120544A (en) | Self-pressurizing radial friction bearing | |
| US2348928A (en) | Bearing construction | |
| US2683635A (en) | Air bearing | |
| JPH0765615B2 (en) | Journal bearing | |
| US3232680A (en) | Fluid bearing | |
| JPH10500469A (en) | Fluid thin film bearing | |
| US5433528A (en) | Two pad axially grooved hydrostatic bearing | |
| AU2017351649B2 (en) | Tilting-pad bearing and method of manufacturing thereof | |
| Rowe | Advances in hydrostatic and hybrid bearing technology | |
| US5983781A (en) | Sliding bearing with self-adjusted load bearing capacity | |
| US6296390B1 (en) | Single plate hydrodynamic bearing with extended single journal bearing | |
| US3508799A (en) | Gas bearings | |
| JPH0276923A (en) | Static pressure gas bearing | |
| JP7005810B1 (en) | Gas drawing structure and static pressure gas bearings that make up the static pressure gas bearing | |
| JPS6235947Y2 (en) | ||
| JP4373055B2 (en) | Thrust dynamic pressure gas bearing | |
| WO1997008470A1 (en) | Integrated shaft-self compensating hydrostatic bearings | |
| JPS5814927B2 (en) | gas bearing | |
| JPH11336753A (en) | Self-aligning roller bearing | |
| JPS63125820A (en) | Static pressure bearing | |
| CN114857273B (en) | End face seal assembly | |
| JPH0257723A (en) | Static pressure gas bearing |
Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| R250 | Receipt of annual fees |
Free format text: JAPANESE INTERMEDIATE CODE: R250 |
|
| LAPS | Cancellation because of no payment of annual fees |