JPS6235947Y2 - - Google Patents
Info
- Publication number
- JPS6235947Y2 JPS6235947Y2 JP1979091516U JP9151679U JPS6235947Y2 JP S6235947 Y2 JPS6235947 Y2 JP S6235947Y2 JP 1979091516 U JP1979091516 U JP 1979091516U JP 9151679 U JP9151679 U JP 9151679U JP S6235947 Y2 JPS6235947 Y2 JP S6235947Y2
- Authority
- JP
- Japan
- Prior art keywords
- bearing
- air supply
- radial
- exhaust
- radial bearing
- 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
- 238000007599 discharging Methods 0.000 claims description 2
- 230000003068 static effect Effects 0.000 claims description 2
- 230000000694 effects Effects 0.000 description 10
- 230000002706 hydrostatic effect Effects 0.000 description 8
- 230000007423 decrease Effects 0.000 description 5
- 238000010586 diagram Methods 0.000 description 2
- 230000003247 decreasing effect Effects 0.000 description 1
- 238000005516 engineering process Methods 0.000 description 1
- 238000005461 lubrication Methods 0.000 description 1
- 238000003754 machining Methods 0.000 description 1
- 238000000034 method Methods 0.000 description 1
- 230000002093 peripheral effect Effects 0.000 description 1
- 230000007103 stamina Effects 0.000 description 1
Landscapes
- Magnetic Bearings And Hydrostatic Bearings (AREA)
Description
【考案の詳細な説明】
〔産業上の利用分野〕
本考案は、外部から加圧気体を微小な給気孔を
通して軸受隙間内に供給し、回転軸をラジアル方
向及びスラスト方向に支える静圧気体軸受に関す
る。[Detailed description of the invention] [Field of industrial application] The present invention is a hydrostatic gas bearing that supports a rotating shaft in the radial and thrust directions by supplying pressurized gas from the outside into the bearing gap through minute air supply holes. Regarding.
回転軸をラジアル方向及びスラスト方向の両方
において支える静圧気体軸受としては、従来第1
図に示すようなものがある。この軸受は、回転軸
01を包む軸受本体02のラジアル軸受面02a
にその中央部周囲に沿つて1列あるいは2列に多
数の給気孔03を設けると共に左右のスラスト軸
受面02bに多数の給気孔04を環状に配列して
なり、給気孔03,04のオリフイス絞りあるい
は自成絞り効果を利用して回転軸01を支えるも
のである。
As a static pressure gas bearing that supports the rotating shaft in both the radial direction and the thrust direction, the first
There is something like the one shown in the figure. This bearing has a radial bearing surface 02a of a bearing body 02 that surrounds a rotating shaft 01.
A large number of air supply holes 03 are provided in one or two rows along the periphery of the central portion, and a large number of air supply holes 04 are arranged in a ring shape on the left and right thrust bearing surfaces 02b, and the orifice throttle of the air supply holes 03, 04 Alternatively, the rotating shaft 01 is supported using a self-generating aperture effect.
従来のこの種の軸受の他のものとしては、特公
昭45−21404号公報及び特公昭46−12682号公報等
に開示されているような、軸の回転方向と直角方
向に多数の微細溝を有する表面絞り軸受がある。 Other conventional bearings of this type include those that have a large number of fine grooves in the direction perpendicular to the rotational direction of the shaft, as disclosed in Japanese Patent Publication No. 45-21404 and Japanese Patent Publication No. 46-12682, etc. There are surface drawing bearings with.
前者の軸受においては、給気孔03,04の径
は加工上0.1mmより小さくすることができないた
め、負荷容量が最大となる最適軸受隙間は30μm
以上となつてしまう。そのため、軸受に負荷を加
えた時の軸受の移動量が大きく、その負荷と軸受
の移動量との比で定義される軸受のステイフネス
(負荷/移動量)が小さく、また多数の給気孔を
必要とするため給気流量が多いという欠点を有し
ていた。
In the former type of bearing, the diameter of the air supply holes 03 and 04 cannot be made smaller than 0.1 mm due to processing reasons, so the optimal bearing clearance for maximum load capacity is 30 μm.
That's all there is to it. Therefore, the amount of movement of the bearing when a load is applied to the bearing is large, the stiffness (load/travel) of the bearing, defined as the ratio of the load to the amount of movement of the bearing, is small, and a large number of air supply holes are required. Therefore, it had the disadvantage that the air supply flow rate was large.
次に、後者の軸受においては、5μm前後の微
小な軸受隙間で高い負荷容量が得られるため、軸
受のステイフネスが大きく、給気流量が少なくて
済むという長所を有しているが、回転方向と直角
方向の多数の微細溝を加工するのが難しく、軸受
がかなり高価になつてしまうという欠点があつ
た。 Next, the latter type of bearing has the advantage that a high load capacity can be obtained with a minute bearing gap of around 5 μm, so the bearing stiffness is large and the supply air flow rate is small. The drawback was that it was difficult to machine many fine grooves in the right angle direction, and the bearings were quite expensive.
本考案は、回転軸をラジアル方向及びスラスト
方向の両方において支える従来の静圧気体軸受に
おける上述のような欠点を解消することを目的と
してなされたもので、その構成は、単数又は複数
のラジアル軸受面からなるラジアル軸受部とその
両側に形成されたスラスト軸受部とから構成さ
れ、ラジアル軸受部とスラスト軸受部との境界に
軸受周囲に通じる排気孔に接続されて周囲圧とな
つている排気通路を有するとともに複数のラジア
ル軸受面の間に軸受周囲に通じる排気孔に接続さ
れて周囲圧となつている排気溝を有し、軸をラジ
アル方向及びスラスト方向に支持する静圧気体軸
受において、このラジアル軸受面の両端面近傍及
び前記スラスト軸受部の両端部近傍に軸受すき間
の2ないし5倍の深さの環状溝を形成すると共
に、気体を吐出する微小な給気孔を前記環状溝内
にそれぞれ3個以上開口したものである。
The present invention was made with the aim of eliminating the above-mentioned drawbacks of conventional hydrostatic gas bearings that support the rotating shaft in both the radial and thrust directions. An exhaust passage is composed of a radial bearing section consisting of a surface and a thrust bearing section formed on both sides of the radial bearing section, and is connected to an exhaust hole leading around the bearing at the boundary between the radial bearing section and the thrust bearing section, and is at ambient pressure. In a hydrostatic gas bearing that supports a shaft in the radial direction and the thrust direction, the bearing has an exhaust groove connected to an exhaust hole that communicates around the bearing between a plurality of radial bearing surfaces and is at ambient pressure. An annular groove having a depth of 2 to 5 times the bearing clearance is formed near both end surfaces of the radial bearing surface and near both ends of the thrust bearing part, and a minute air supply hole for discharging gas is formed in each of the annular grooves. It has three or more openings.
本考案では、加圧空気が給気圧力Psから給気
孔の絞り抵抗R1を通り、環状溝を通つて円周方
向に拡つた後に軸受すき間内の流れ抵抗R2を通
つて排気溝、排気通路等の大気圧Paへと低下す
る。このため、軸受すき間の変化により、流れ抵
抗R2が場所によつて変化し、狭まり軸受すき間
領域の軸受内部圧力が増加し、拡がり軸受すき間
領域の軸受内部圧力が減少して回転軸を支持して
いる。特に環状溝の効果により円周方向に緩慢に
変化する圧力分布を形成できるため、円周方向に
3個と少ない給気孔でも加工不可能な程の微小な
孔径の給気孔を多数あけた場合と同一の効果があ
り、回転軸の任意方向の偏心を復元する圧力を発
生できる。すなわち、両側の排気溝の近傍に給気
孔列を設けても環状溝によつて円周方向の圧力変
化の連続性が保持され、円周方向の軸受面の利用
効率が低下せず、かつ2つの環状溝間の広い領域
で大きな復元力を発生できる。
In the present invention, pressurized air passes from the supply pressure Ps through the restriction resistance R 1 of the air supply hole, expands in the circumferential direction through the annular groove, and then passes through the flow resistance R 2 in the bearing gap to the exhaust groove. The pressure drops to atmospheric pressure Pa in passageways, etc. Therefore, due to changes in the bearing clearance, the flow resistance R 2 changes depending on the location, increasing the bearing internal pressure in the narrowing bearing clearance area, and decreasing the bearing internal pressure in the widening bearing clearance area to support the rotating shaft. ing. In particular, the effect of the annular groove allows a pressure distribution that changes slowly in the circumferential direction to be formed, so even if there are as few as three air supply holes in the circumferential direction, it is possible to create a large number of air supply holes with a diameter so small that it is impossible to process them. It has the same effect and can generate pressure to restore the eccentricity of the rotating shaft in any direction. In other words, even if air supply hole rows are provided near the exhaust grooves on both sides, the continuity of the pressure change in the circumferential direction is maintained by the annular groove, and the utilization efficiency of the bearing surface in the circumferential direction does not decrease. A large restoring force can be generated in the wide area between the two annular grooves.
また、静圧軸受において、給気孔数n、給気孔
径d、給気孔から排気溝あるいは排気通路に接す
る軸受端までの距離l1とすると負荷容量を最大と
する最適軸受すき間はn・d・l1の積に比例する
ことが知られており、本考案では前述のような環
状溝の効果によつて給気孔数n、給気孔径d、距
離l1のすべてを小さくできることから最適軸受す
き間を最小にでき、加圧気体の消費量をきわめて
少なくできる。 Furthermore, in a hydrostatic bearing, if the number of air supply holes is n, the diameter of the air supply hole is d, and the distance from the air supply hole to the end of the bearing in contact with the exhaust groove or exhaust passage is l1 , then the optimal bearing clearance that maximizes the load capacity is n・d・It is known that the bearing clearance is proportional to the product of l1 , and in this invention, the number of air supply holes n, the air supply hole diameter d, and the distance l1 can all be reduced due to the effect of the annular groove as described above. can be minimized, and consumption of pressurized gas can be extremely reduced.
以下、本考案の一実施例を図面に基づき詳細に
説明する。
Hereinafter, one embodiment of the present invention will be described in detail based on the drawings.
第2図において回転軸1に設けた小径部1aと
該小径部1a表面に直角なその両端の側壁部1b
とからなる溝内に軸受本体2は収められる。前記
小径部1aを囲む軸受本体2のラジアル軸受面2
aと前記側壁部1bと対面する左右二つのスラス
ト軸受面2bとの交わり部(境界部)には面取り
が施され、面取り部には、回転軸1に形成されか
つ回転軸1の周囲に通じる排気孔3が接続され
て、周囲圧(大気圧)になつている排気通路4が
構成されている。軸受本体2のラジアル軸受面2
aのある面の中央部には環状排気溝5が設けられ
て、ラジアル軸受面2aは二つに分割されてい
る。前記環状排気溝5には、軸受本体2の周囲に
通じさせて軸受本体2に設けた排気孔6が接続さ
れて、前記環状排気溝5を周囲圧にしている。分
割された二つのラジアル軸受面2aにおける前記
排気通路4の近傍及び前記環状排気溝5の近傍に
はラジアル環状溝7が形成されている。これらの
ラジアル環状溝7内には、軸受本体2に形成され
た気体通路8に通じる微小な径の給気孔9が第3
図に示すように等間隔に三つ開口している。ま
た、前記二つのスラスト軸受面2bにおいてはそ
の外周面近傍及び前記排気通路4近傍にスラスト
環状溝10が形成され、それぞれの環状溝10内
には第4図に示すように等間隔に三つの給気孔1
1が開口されている。図面中、12は前記気体通
路8に気体を供給するための口金である。 In FIG. 2, a small diameter portion 1a provided on the rotating shaft 1 and side wall portions 1b at both ends thereof perpendicular to the surface of the small diameter portion 1a.
The bearing body 2 is housed in the groove formed by the above. Radial bearing surface 2 of the bearing body 2 surrounding the small diameter portion 1a
The intersection (boundary) between the left and right thrust bearing surfaces 2b facing the side wall portion 1b is chamfered, and the chamfered portion is formed on the rotating shaft 1 and communicates with the surroundings of the rotating shaft 1. The exhaust hole 3 is connected to constitute an exhaust passage 4 which is at ambient pressure (atmospheric pressure). Radial bearing surface 2 of bearing body 2
An annular exhaust groove 5 is provided in the center of the surface a, and the radial bearing surface 2a is divided into two. An exhaust hole 6 provided in the bearing body 2 and communicating around the bearing body 2 is connected to the annular exhaust groove 5, so that the annular exhaust groove 5 is brought to ambient pressure. A radial annular groove 7 is formed near the exhaust passage 4 and near the annular exhaust groove 5 on the two divided radial bearing surfaces 2a. Inside these radial annular grooves 7, there are third air supply holes 9 with minute diameters that communicate with the gas passages 8 formed in the bearing body 2.
As shown in the figure, there are three openings at equal intervals. Further, in the two thrust bearing surfaces 2b, thrust annular grooves 10 are formed in the vicinity of the outer peripheral surface thereof and in the vicinity of the exhaust passage 4, and within each annular groove 10, as shown in FIG. Air supply hole 1
1 is open. In the drawing, 12 is a base for supplying gas to the gas passage 8.
口金12に供給された気体は気体通路8を通
り、各給気孔9,11より吐出されて、ラジアル
軸受面2aと回転軸1の小径部1aとの間及びス
ラスト軸受面2bと回転軸1の側壁部1bとの間
にそれぞれ流入し、軸を支える負荷を発生する。 The gas supplied to the mouthpiece 12 passes through the gas passage 8 and is discharged from each air supply hole 9, 11, and is discharged between the radial bearing surface 2a and the small diameter portion 1a of the rotating shaft 1, and between the thrust bearing surface 2b and the rotating shaft 1. They each flow between the side wall portion 1b and generate a load that supports the shaft.
本考案の第1の特徴は各軸受面2a,2bの両
端近傍に設けられた環状溝7,10が給気孔9,
11からの加圧気体の大部分を円周方向に導くに
十分なだけ深く、且つ軸受隙間が十分大きい領域
では給気孔からの気体の大半が即座に軸受端まで
流出する程度に浅いことである。第2の特徴はラ
ジアル環状溝7及びスラスト環状溝10内に設け
られた給気孔9,11の内径は加工上可能な限り
小さく、例えば0.1mm程度で、また円周方向の孔
の数も3〜6個程度(本実施例では120度間隔に
3個)としたことである。すなわち、環状溝7,
10がなければ給気孔を多数(例えば10個以上)
開口する必要があるが、環状溝の効果によつて加
圧気体の圧力低下を抑えつつ円周方向に分散させ
ることのできるので、給気孔数は少なくとも加工
上不可能な微小な孔径の給気孔を円周方向に多数
あけたものと同じ効果を得ることができる。な
お、環状溝内に2つの給気孔では給気孔方向に偏
心したときには復元力が働くが、給気孔方向と直
角な方向などの任意の方向では復元力は弱くな
り、回転軸の偏心を元にもどせなくなる。そこ
で、回転軸が任意の方向に偏心しても一様な復元
力を持たせるには、少なくとも環状溝内に120℃
間隔で3つの給気孔を開口させる必要がある。 The first feature of the present invention is that the annular grooves 7, 10 provided near both ends of each bearing surface 2a, 2b are connected to the air supply hole 9,
It is deep enough to guide most of the pressurized gas from No. 11 in the circumferential direction, and shallow enough that in areas where the bearing gap is large enough, most of the gas from the air supply hole immediately flows out to the end of the bearing. . The second feature is that the inner diameters of the air supply holes 9 and 11 provided in the radial annular groove 7 and the thrust annular groove 10 are as small as possible due to machining, for example, about 0.1 mm, and the number of holes in the circumferential direction is 3. ~6 pieces (in this example, 3 pieces at 120 degree intervals). That is, the annular groove 7,
If there is no 10, install many air supply holes (for example, 10 or more)
However, due to the effect of the annular groove, the pressurized gas can be dispersed in the circumferential direction while suppressing the pressure drop. It is possible to obtain the same effect as having a large number of holes in the circumferential direction. Note that with two air supply holes in an annular groove, a restoring force acts when the air supply hole is eccentric in the direction of the air supply hole, but the restoring force becomes weak in any direction such as a direction perpendicular to the direction of the air supply hole, and the restoring force acts on the eccentricity of the rotation axis. I can't go back. Therefore, in order to have a uniform restoring force even if the rotating shaft is eccentric in any direction, it is necessary to
It is necessary to open three air supply holes at intervals.
このような構造を有する軸受の回転軸1が第7
図に示すように偏心している場合(第7図は偏心
を強調している。)に加圧気体を供給すると各軸
受内の幅方向の圧力分布は軸受すき間が狭いとこ
ろと、軸受すき間の広いところとでおのおの第5
図a中のA及びA′のような台形となる。すなわ
ち、環状溝7間の距離が短いため、環状溝7間の
領域において円周方向の流路抵抗に比し、軸受幅
方向のそれが小さいので、軸受幅方向の圧力変化
は少なく、軸受隙間の増減に応じて台形圧力分布
はそれぞれA′及びAのように変化し、これら
A′及びAの圧力差によつて回転軸1の偏心を復
元するように働く。 The rotating shaft 1 of the bearing having such a structure is the seventh
When pressurized gas is supplied when the bearing is eccentric as shown in the figure (the eccentricity is emphasized in Figure 7), the pressure distribution in the width direction within each bearing will be found in areas where the bearing gap is narrow and where the bearing gap is wide. By the way, the fifth
It becomes a trapezoid like A and A' in figure a. In other words, since the distance between the annular grooves 7 is short, the flow resistance in the bearing width direction is smaller than the flow path resistance in the circumferential direction in the area between the annular grooves 7, so the pressure change in the bearing width direction is small, and the bearing clearance is small. The trapezoidal pressure distribution changes as A' and A, respectively, as
The pressure difference between A' and A works to restore the eccentricity of the rotating shaft 1.
なお、軸受幅が軸受径より大きい場合には、軸
受内の幅方向の流路抵抗が円周方向に対して無視
できなくなるため第5図b中BとB′とで示すよう
に中央部の圧力差が少なくなり、軸受の負荷容量
低下をもたらす。 Note that when the bearing width is larger than the bearing diameter, the flow path resistance in the width direction within the bearing cannot be ignored in the circumferential direction. The pressure difference decreases, resulting in a decrease in the load capacity of the bearing.
そこで、第2図に示すようにラジアル軸受面の
ある面の中央に周囲圧となつている環状排気溝5
と排気孔6を設けてラジアル軸受面を2つに分割
し、第5図c中CをC′で示すように軸受部すべ
てにわたつて圧力差が大きく発生するようにし
て、軸受の負荷容量を向上している。 Therefore, as shown in Fig. 2, an annular exhaust groove 5 with ambient pressure at the center of the radial bearing surface.
The radial bearing surface is divided into two parts by providing an exhaust hole 6 and an exhaust hole 6, and by creating a large pressure difference across all bearing parts as shown by C and C' in Figure 5, the load capacity of the bearing can be increased. has been improved.
次に、ラジアル軸受面2aにおける給気孔9列
に沿つてこの円周方向の圧力分布特性を第6図に
示す。第6図では、第7図に示すように3つの給
気孔9がある場合に回転軸1が下方に偏心した時
の最大すきま位置からの角度φと圧力の関係をラ
ジアル環状溝の深さをパラメータとして表してい
る。円周方向圧力分布特性は主として給気孔9を
開口してあるラジアル環状溝7の深さに依存し、
軸受すき間の2〜5倍程度の深さを持つた環状溝
を形成すると、第6図中の曲線Cで示すように、
回転軸1の変位に応じて生ずる、狭まりすき間領
域の圧力(第6図中で角度φが90゜から180゜の
領域を積分した圧力)と広がりすき間領域のの圧
力(第6図中で角度φが0゜から90゜の領域を積
分した圧力)の差が最大となる。 Next, FIG. 6 shows pressure distribution characteristics in the circumferential direction along the nine rows of air supply holes on the radial bearing surface 2a. Figure 6 shows the relationship between the angle φ from the maximum clearance position and the pressure when the rotating shaft 1 is eccentric downward when there are three air supply holes 9 as shown in Figure 7, and the depth of the radial annular groove. It is expressed as a parameter. The circumferential pressure distribution characteristics mainly depend on the depth of the radial annular groove 7 opening the air supply hole 9,
When an annular groove with a depth of about 2 to 5 times the bearing clearance is formed, as shown by curve C in Fig. 6,
The pressure in the narrowing gap area (the pressure integrated over the area where the angle φ ranges from 90° to 180° in Figure 6) and the pressure in the widening gap area (the angle in Figure 6 The difference in pressure (integrated pressure over the area where φ is 0° to 90°) is maximum.
もし、環状溝7の深さがこの最適値より小さけ
れば、加圧気体は給気孔9間に十分満たされぬま
ま軸受端から流出するため圧力分布は曲線Dで示
すようになり、逆に最適値より大きければ広がり
隙間領域内でも加圧気体が円周方向に充満し過
ぎ、圧力分布は曲線Eのようになる。したがつ
て、いずれの場合も圧力差は減少し軸受の負荷容
量が低下する。 If the depth of the annular groove 7 is smaller than this optimum value, the pressurized gas will flow out from the bearing end without being sufficiently filled between the air supply holes 9, so the pressure distribution will become as shown by curve D, and vice versa. If it is larger than this value, the pressurized gas will fill too much in the circumferential direction even within the expanded gap region, and the pressure distribution will become like curve E. Therefore, in either case, the pressure difference decreases and the load capacity of the bearing decreases.
第5図及び第6図に示した幅方向及び円周方向
の圧力分布特性は第2図の実施例に示すラジアル
軸受部、スラスト軸受部のいずれにも共通する特
性である。したがつて、本考案の軸受は給気孔数
が3〜6個ときわめて少ないにもかかわらず、環
状溝の効果により軸受内の圧力発生効率が良く、
従来の軸受に比べラジアル方向、スラスト方向及
び傾きに関する軸受負荷容量を50%以上高めるこ
とができるのである。 The pressure distribution characteristics in the width direction and circumferential direction shown in FIGS. 5 and 6 are common to both the radial bearing section and the thrust bearing section shown in the embodiment shown in FIG. 2. Therefore, although the bearing of the present invention has a very small number of air supply holes (3 to 6), the effect of the annular groove allows for high pressure generation efficiency within the bearing.
Compared to conventional bearings, the bearing load capacity in the radial direction, thrust direction, and inclination can be increased by more than 50%.
また、本考案の軸受は環状溝の効果により給気
孔列を軸受端近傍に配置できるとともに給気孔径
が0.1mm前後と小さくかつその数も3〜6個を少
なくてよく、負荷容量が最大となる最適軸受すき
間を5μm前後に小さくすることができる。すな
わち、気体潤滑理論にもとづく計算によれば、第
2図に示す構造の軸受でラジアル軸受径50mm、全
ラジアル軸受幅50mm、スラスト軸受外径100mm、
各環状溝当りの給気孔数3、給気孔径0.15mm、環
状溝幅1mm、環状溝と軸受端の距離2mmの場合に
は、スラスト軸受部の外周側の最適溝深さ約45μ
m、内周側の最適溝深さ約30μm、ラジアル軸受
部の最適溝深さ約25μmとなり、また負荷容量が
最大となる最適軸受すき間は約6μmとなる。こ
こで環状溝の最適深さは軸受すき間の2〜5倍で
ある。 In addition, the bearing of the present invention allows the air supply hole array to be placed near the bearing end due to the effect of the annular groove, and the diameter of the air supply hole is small, around 0.1 mm, and the number of air supply holes can be as small as 3 to 6, resulting in maximum load capacity. The optimum bearing clearance can be reduced to around 5 μm. In other words, according to calculations based on gas lubrication theory, a bearing with the structure shown in Figure 2 has a radial bearing diameter of 50 mm, a total radial bearing width of 50 mm, a thrust bearing outer diameter of 100 mm,
When the number of air supply holes per annular groove is 3, the air supply hole diameter is 0.15 mm, the annular groove width is 1 mm, and the distance between the annular groove and the bearing end is 2 mm, the optimal groove depth on the outer circumferential side of the thrust bearing part is approximately 45μ.
m, the optimum groove depth on the inner circumferential side is about 30 μm, the optimum groove depth on the radial bearing part is about 25 μm, and the optimum bearing clearance for maximum load capacity is about 6 μm. Here, the optimum depth of the annular groove is 2 to 5 times the bearing clearance.
このように本考案の軸受は給気孔径が小さく、
孔数が少なく、しかも軸受すきまが小さいので加
圧気体の消費量をきわめて少なくできる。 In this way, the bearing of this invention has a small air supply hole diameter,
Since the number of holes is small and the bearing clearance is small, the amount of pressurized gas consumed can be extremely reduced.
本考案の静圧気体軸受組立体は第2図に示した
実施例以外にも各種の形式が考えられる。例えば
ラジアル軸受部の長さが短かければ、中央の排気
溝5及びその両側の環状溝と給気孔列をなくし、
一個のラジアル軸受とすることも可能である。ま
た逆にラジアル軸受部を長くして、ラジアル方向
の負荷容量を大きくしたい場合にはラジアル軸受
部を2本あるいは3本の排気溝で3個あるいは4
個のラジアル軸受面に分割しそれぞれの軸受面に
ついて2列の環状溝と給気孔を設けることも可能
である。なお、第2図には、ラジアル軸受面2a
と該ラジアル軸受面2aが囲む小径部1aとから
なるラジアル軸受部及びスラスト軸受面2bとこ
れと対向する側壁部1bとからなるスラスト軸受
部において、ラジアル軸受面2a及びスラスト軸
受面2bにそれぞれ環状溝7,10、給気孔9,
11を設けたものを実施例として示したが、これ
らの環状溝、給気孔を回転軸1側の前記小径部1
a及び側壁部1bに設けることも可能である。 The hydrostatic gas bearing assembly of the present invention may have various types other than the embodiment shown in FIG. For example, if the length of the radial bearing part is short, the central exhaust groove 5 and the annular grooves and air supply hole rows on both sides of the central exhaust groove 5 are eliminated;
It is also possible to use a single radial bearing. On the other hand, if you want to increase the load capacity in the radial direction by making the radial bearing part longer, you can use 3 or 4 radial bearing parts with 2 or 3 exhaust grooves.
It is also possible to divide the bearing surface into separate radial bearing surfaces and provide two rows of annular grooves and air supply holes for each bearing surface. In addition, in FIG. 2, the radial bearing surface 2a
In the radial bearing section consisting of a small diameter section 1a surrounded by the radial bearing surface 2a, and the thrust bearing section consisting of a thrust bearing surface 2b and a side wall section 1b opposite thereto, the radial bearing surface 2a and the thrust bearing surface 2b each have an annular shape. Grooves 7, 10, air supply hole 9,
11 is shown as an example, these annular grooves and air supply holes are connected to the small diameter portion 1 on the rotating shaft 1 side.
It is also possible to provide it on the side wall portion a and the side wall portion 1b.
以上述べたように、本考案に係る静圧気体軸受
組立体によれば、微小な軸受すき間で高い負荷容
量を持たせることができるので、精密工作機械等
で必要なラジアル方向、スラスト方向及び傾きに
対するステイフネスが高まる。また、給気孔数は
少なくて済むので、給気流量は小さくなる。
As described above, according to the hydrostatic gas bearing assembly according to the present invention, it is possible to have a high load capacity with a small bearing gap, so that Increased stamina. Furthermore, since the number of air supply holes is small, the air supply flow rate is reduced.
第1図は従来の給気孔形静圧気体軸受組立体の
縦断面図、第2図は本考案に係る静圧気体軸受組
立体の一実施例の縦断面図、第3図は第2図に示
した実施例におけるラジアル軸受部の給気孔列に
沿つた断面図、第4図は第2図に示した実施例の
スラスト軸受面の側面図、第5図は実施例におけ
る軸受内の幅方向の圧力分布特性を示す説明図、
第6図は実施例における円周方向の圧力分布の環
状溝の深さによる違いを示すグラフ、第7図は実
施例における回転軸の偏心を強調して示した図で
ある。
1……回転軸、1a……小径部、1b……側壁
部、2……軸受本体、2a……ラジアル軸受面、
2b……スラスト軸受面、3,6……排気孔、4
……排気通路、5……環状排気溝、7……ラジア
ル環状溝、8……気体通路、9,11……給気
孔、10……スラスト環状溝。
FIG. 1 is a longitudinal sectional view of a conventional air supply hole type hydrostatic gas bearing assembly, FIG. 2 is a longitudinal sectional view of an embodiment of the hydrostatic gas bearing assembly according to the present invention, and FIG. FIG. 4 is a side view of the thrust bearing surface of the embodiment shown in FIG. 2, and FIG. 5 is a cross-sectional view along the air supply hole row of the radial bearing in the embodiment shown in FIG. An explanatory diagram showing the pressure distribution characteristics in the direction,
FIG. 6 is a graph showing the difference in pressure distribution in the circumferential direction depending on the depth of the annular groove in the example, and FIG. 7 is a diagram emphasizing the eccentricity of the rotation axis in the example. DESCRIPTION OF SYMBOLS 1... Rotating shaft, 1a... Small diameter part, 1b... Side wall part, 2... Bearing body, 2a... Radial bearing surface,
2b... Thrust bearing surface, 3, 6... Exhaust hole, 4
... Exhaust passage, 5 ... Annular exhaust groove, 7 ... Radial annular groove, 8 ... Gas passage, 9, 11 ... Air supply hole, 10 ... Thrust annular groove.
Claims (1)
ル軸受部とその両側に形成されたスラスト軸受部
とから構成され、ラジアル軸受部とスラスト軸受
部との境界に軸受周囲に通じる排気孔に接続され
て周囲圧となつている排気通路を有するとともに
複数のラジアル軸受面の間に軸受周囲に通じる排
気孔に接続されて周囲圧となつている排気溝を有
し、軸をラジアル方向及びスラスト方向に支持す
る静圧気体軸受において、このラジアル軸受面の
両端面近傍及び前記スラスト軸受部の両端部近傍
に軸受すき間の2ないし5倍の深さの環状溝を形
成すると共に、気体を吐出する微小な給気孔を前
記環状溝内にそれぞれ3個以上開口してあること
を特徴とする静圧気体軸受組立体。 It consists of a radial bearing part consisting of one or more radial bearing surfaces and a thrust bearing part formed on both sides of the radial bearing part, and the boundary between the radial bearing part and the thrust bearing part is connected to an exhaust hole that communicates around the bearing to reduce the ambient pressure. It has an exhaust passage that is connected to the exhaust hole that communicates around the bearing between the plurality of radial bearing surfaces, and has an exhaust groove that is at ambient pressure. In the pressurized gas bearing, an annular groove with a depth of 2 to 5 times the bearing clearance is formed near both end faces of the radial bearing surface and near both ends of the thrust bearing part, and a minute air supply hole for discharging gas is formed. A static pressure gas bearing assembly characterized in that each of the annular grooves has three or more openings.
Priority Applications (4)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
JP1979091516U JPS6235947Y2 (en) | 1979-07-03 | 1979-07-03 | |
GB8000476A GB2046370B (en) | 1979-01-13 | 1980-01-08 | Gas bearing |
NL8000171A NL8000171A (en) | 1979-01-13 | 1980-01-11 | GAS BEARINGS. |
DE19803001061 DE3001061A1 (en) | 1979-01-13 | 1980-01-12 | GAS WAREHOUSE |
Applications Claiming Priority (1)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
JP1979091516U JPS6235947Y2 (en) | 1979-07-03 | 1979-07-03 |
Publications (2)
Publication Number | Publication Date |
---|---|
JPS568924U JPS568924U (en) | 1981-01-26 |
JPS6235947Y2 true JPS6235947Y2 (en) | 1987-09-12 |
Family
ID=29324576
Family Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
JP1979091516U Expired JPS6235947Y2 (en) | 1979-01-13 | 1979-07-03 |
Country Status (1)
Country | Link |
---|---|
JP (1) | JPS6235947Y2 (en) |
Families Citing this family (4)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
JP3517726B2 (en) * | 1996-08-12 | 2004-04-12 | 株式会社日平トヤマ | Spindle device |
DE102010015794B3 (en) * | 2010-04-20 | 2011-06-09 | Gebrüder Klöcker GmbH | Leno selvedge forming device, has bearing ring comprising circumferential distributed openings for compressed air supply, where circumferential distributed openings are arranged in region of bearing collar |
JP5915088B2 (en) * | 2011-10-31 | 2016-05-11 | オイレス工業株式会社 | Static pressure gas bearing and linear motion guide device using the static pressure gas bearing |
KR101165749B1 (en) | 2011-12-06 | 2012-07-18 | 한국기계연구원 | Double row gas bearing |
-
1979
- 1979-07-03 JP JP1979091516U patent/JPS6235947Y2/ja not_active Expired
Also Published As
Publication number | Publication date |
---|---|
JPS568924U (en) | 1981-01-26 |
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