JP2007192076A - Turbo vacuum pump - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a turbo vacuum pump having pump discharge performance capable of dealing with increase in a flow rate and reduction of pressure by rotating a rotor at high speed without enlarging pump blade diameter, and favorable for use for a long period of time. <P>SOLUTION: The turbo vacuum pump 1 is provided with a suction part 23A sucking gas in an axial direction, an discharge part 50 having rotors 70, 24 and fixed blades 71, 28 alternately arranged thereon, and a rotary shaft 21 rotating the rotor. The rotor includes a one or more stages of turbine blades 70 discharging the sucked gas in the axial direction and one or more stage centrifugal blades 24 positioned in a slip stream side of the turbine blades and further discharging the discharged gas by centrifugal drag action. The centrifugal blade is fixed on the rotary shaft passing through the centrifugal blade and is arranged between the centrifugal blade and the rotary shaft, and is provided with a cylindrical ring 41 interference-fitted on the rotary shaft. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a turbo type vacuum pump suitable for use in exhausting a relatively large flow rate gas, and more particularly to a turbo type vacuum pump having a large exhaust speed in a region where a pump inlet pressure is 1 to 1000 Pa.

  FIG. 16 shows an example of a conventional turbo vacuum pump 1C. Currently, there is a turbo molecular pump as a vacuum pump that is widely used for semiconductor processes such as semiconductor manufacturing equipment.

  This turbo type vacuum pump 1C is composed of a blade exhaust part L1 and a groove exhaust part L2 by a rotor (rotating part) R and a stator (fixed part) S inside a cylindrical pump casing 101 arranged vertically above and below. An exhaust portion L is configured. A lower portion of the pump casing 101 is covered with a pump base 102, and an exhaust port 120 communicating with the exhaust side of the groove exhaust portion L2 is formed in the pump base 102. A flange (not shown in FIG. 16) is provided on the upper portion of the pump casing 101 having the intake port 101a for connection to a device or piping for exhausting gas. The stator S is mainly composed of a fixed cylindrical portion 103 erected at the center of the pump base portion 102 and fixed side portions of the blade exhaust portion L1 and the groove exhaust portion L2.

  The rotor R includes a rotating shaft 104 inserted into the fixed cylindrical portion 103 and a rotating cylindrical portion 105 attached thereto. The fixed cylindrical portion 103 is accommodated in the hollow portion 105 a of the rotating cylindrical portion 105. A driving motor 106 is provided between the rotating shaft 104 and the fixed cylindrical portion 103, and an upper radial bearing 107 and a lower radial bearing 108 are provided above and below the driving motor 106. An axial bearing 111 having a target disk 109 at the lower end of the rotating shaft 104 and upper and lower electromagnets 110 a and 110 b on the stator S side is disposed below the rotating shaft 104. With such a configuration, the rotor R rotates at high speed while receiving active control of five axes.

  Rotor blades 112 are integrally provided on the upper and lower outer circumferences of the rotating cylindrical portion 105 to form an impeller, and fixed vanes 113 alternately disposed with the rotor blades 112 are provided on the inner surface of the pump casing. Constitutes a blade exhaust portion L1 that performs exhaust by the interaction between the rotating blade 112 and the stationary blade 113 that is stationary when rotating at high speed. The fixed wing 113 is fixed with its peripheral edge pressed from above and below by a fixed wing spacer 114.

  Further, a groove exhaust portion L2 is provided below the blade exhaust portion L1. That is, a thread groove spacer 119 surrounding the outer periphery of the rotor R is disposed in the stator S, and a thread groove 119a is formed in the thread groove spacer 119. The groove exhaust portion L2 exhausts by the drag action of the screw groove 119a facing the rotor R that rotates at high speed (for example, Patent Document 1).

  Thus, the wide-area turbo vacuum pump 1C that can handle a wide flow rate range is configured by having the thread groove exhaust section L2 on the downstream side of the blade exhaust section L1. In this example, the thread groove of the thread groove exhaust portion L2 is formed on the stator S side, but the thread groove is also formed on the rotor R side.

  As described above, a composite turbo-type vacuum pump that combines a rotor blade with a turbine blade that efficiently exhausts gas in the molecular flow region and a rotor that has a thread groove that efficiently exhausts gas in the intermediate flow region. The composite type is suitable for applications in which a relatively large amount of gas flows.

Japanese Patent Application Laid-Open No. 60-12595

  However, the conventional turbo vacuum pump has a characteristic that the exhaust speed decreases as the pump intake pressure increases on the high pressure side of 1 Pa or more. For this reason, in order to cope with a large flow rate and a low pressure, a large pump is necessary.

  Needless to say, the exhaust performance of the turbo-type vacuum pump is better when the rotating cylindrical portion is rotated as fast as possible. However, a general turbo molecular pump structure is formed in a shape in which a rotating cylindrical part constituting an impeller covers and surrounds an outer side of a fixed cylindrical part constituting a stator. Therefore, the rotation speed of the turbo vacuum pump is limited by the stress generated in the maximum inner diameter portion of the rotating cylindrical portion. In the conventional turbo type vacuum pump, since the rotational speed is limited, when a relatively high flow rate and a low pressure are required, a pump with a high exhaust speed, that is, a pump with a large turbine blade diameter is required. The pump will become large.

  In addition, since the rotating cylindrical portion is formed as described above, the rotating cylindrical portion needs to have an integral structure. If a portion of the rotating cylindrical portion is damaged, deformed, corroded, etc., the rotating cylindrical portion It was likely that the entire department had to be replaced, which was disadvantageous for long-term use.

  Therefore, in view of the above points, the present invention rotates a rotor blade having high exhaust efficiency at a higher speed in a pressure range of 1 to 1000 Pa, and without increasing the pump blade diameter, An object of the present invention is to provide a turbo vacuum pump that can increase the exhaust speed and is advantageous for long-term use.

  In order to achieve the above object, a turbo type vacuum pump 1 according to the invention according to claim 1 includes an intake portion 23A for sucking gas in the axial direction, rotating blades 70 and 24, a fixed blade 71, as shown in FIG. 28 and the rotating shaft 21 that rotates the rotor blades 70 and 24; the turbine blades 70 and 24 are one or more stages of turbines that exhaust the sucked gas in the axial direction. The blade 70 is located on the downstream side of the turbine blade 70 and includes one or more stages of centrifugal blades 24 for exhausting the exhausted gas by centrifugal drag action; And a circular pipe ring 41 which is fixed between the centrifugal blade 24 and the rotary shaft 21 and is tightly fitted to the rotary shaft 21.

  With this configuration, gas is sucked in from the intake portion in the axial direction, and exhausted by the interaction between the rotating blades and the stationary stationary blades when rotating at high speed. The sucked gas is exhausted in the axial direction by the turbine blade, and further exhausted by the centrifugal drag action by the centrifugal blade.

  With the above blade configuration, a turbo type vacuum pump is configured by combining a turbine blade having high exhaust efficiency on the relatively low pressure side and a centrifugal blade having high exhaust efficiency on the relatively high pressure side. Can be high. In addition, since the centrifugal blade exhausts gas in the radial direction, the flow path length can be increased without increasing the axial length. Therefore, since the length of the rotating shaft portion to which the turbine blade and the centrifugal blade are attached can be shortened, the natural frequency of the entire rotor is increased, and high-speed rotation is facilitated.

  Since the centrifugal blade is fixed to the rotating shaft passing through the centrifugal blade, the diameter of the boss portion of the centrifugal blade can be reduced. Moreover, since the flow of a radial direction can be produced and the flow path length can be lengthened, compression performance improves. In addition, since it is further provided with a circular ring that is placed between the centrifugal blade and the rotary shaft and is tightly fitted to the rotary shaft, the bending rigidity of the rotary shaft that is tightly fitted with the circular ring can be increased, and high-speed rotation is achieved. Is possible. As a result, even when a large flow rate gas is sucked in, the intake pressure can be reduced by the exhaust action of the turbine blades, and it can be compressed to a high pressure by the exhaust action of the centrifugal blades.

  A turbo type vacuum pump 1 according to a second aspect of the present invention is the turbo type vacuum pump according to the first aspect, wherein, for example, as shown in FIG. 1, the turbine blades 70 are fixed to the intake portion side end surface 15 of the rotating shaft 21. Is done.

  Turbine blades are fixed to the end surface of the rotary shaft on the intake side, so the diameter of the bosses of the turbine blades can be reduced, the centrifugal force acting on the bosses of the turbine blades can be reduced, and high-speed rotation is possible It becomes. As a result, even when a large flow rate gas is sucked in, the intake pressure can be reduced by the exhaust action of the turbine blades, and it can be compressed to a high pressure by the exhaust action of the centrifugal blades. Further, since the turbine blade and the centrifugal blade are separated from each other by this structure, when the rotor blade is partially damaged, deformed, corroded, etc., the rotor blade may be replaced. Since it is not necessary to replace the whole, a turbo type vacuum pump which is advantageous for long-term use can be obtained.

  In order to achieve the above object, a turbo vacuum pump 1 according to a third aspect of the present invention includes an intake portion 23A for sucking gas in the axial direction and rotating blades 70 and 24, as shown in FIGS. An exhaust section 50 in which blades 71 and 28 are alternately arranged, and a rotary shaft 21 that rotates the rotary blades 70 and 24; one stage in which the rotary blades 70 and 24 exhaust the sucked gas in the axial direction. The turbine blade 70 is located on the downstream side of the turbine blade 70 and includes one or more centrifugal blades 24 for exhausting the exhausted gas by centrifugal drag action; The axial distance LX between the turbine blade 70 and the final stage of the turbine blade 70 is 12% or more of the outer diameter Dt (FIG. 8) of the final stage of the turbine blade 70.

  With this configuration, gas is sucked in from the intake portion in the axial direction, and exhausted by the interaction between the rotating blades and the stationary stationary blades when rotating at high speed. The sucked gas is exhausted in the axial direction by the turbine blade, and further exhausted by the centrifugal drag action by the centrifugal blade.

  With the above blade configuration, a turbo type vacuum pump is configured by combining a turbine blade having high exhaust efficiency on the relatively low pressure side and a centrifugal blade having high exhaust efficiency on the relatively high pressure side. Can be high. In addition, since the centrifugal blade exhausts gas in the radial direction, the flow path length can be increased without increasing the axial length. Therefore, since the length of the rotating shaft portion to which the turbine blade and the centrifugal blade are attached can be shortened, the natural frequency of the entire rotor is increased, and high-speed rotation is facilitated.

  Since the axial distance between the first stage of the centrifugal blade and the final stage of the turbine blade is 12% or more of the outer diameter of the final stage of the turbine blade 70, it flows in the axial direction from the final stage of the turbine blade. The gas flow direction smoothly changes to the intake section of the first stage of the centrifugal blade in the space (axial distance) between the first stage of the centrifugal blade and the final stage of the turbine blade, and the gas flows to the first stage of the centrifugal blade. Inhaled smoothly into the eyes. Since the flow changes smoothly, there is little pressure loss in this space, and the performance of the turbo vacuum pump can be improved.

  In order to achieve the above object, a turbo type vacuum pump 1 according to a fourth aspect of the present invention includes an intake portion 23A for sucking gas in the axial direction and rotating blades 70 and 24, as shown in FIGS. An exhaust section 50 in which blades 71 and 28 are alternately arranged, and a rotary shaft 21 that rotates the rotary blades 70 and 24; one stage in which the rotary blades 70 and 24 exhaust the sucked gas in the axial direction. The turbine blade 70 is located on the downstream side of the turbine blade 70 and includes one or more centrifugal blades 24 for exhausting the exhausted gas by centrifugal drag action; Further comprising a partition wall 43 having an opening 43A immediately upstream thereof; a first stage of the centrifugal blade 24 is disposed so as to suck the gas from the opening 43A; a shaft between the partition wall 43 and the final stage of the turbine blade 70 The direction distance Ly is Is about 12% or more of the outside diameter of the last stage of the turbine blades 70.

  Since the axial distance between the bulkhead and the final stage of the turbine blade is about 12% or more of the outer diameter of the final stage of the turbine blade, the flow direction of the axially flowing gas exiting the final stage of the turbine blade However, in the space between the bulkhead and the final stage of the turbine blade (axial distance), the gas changes smoothly to the opening of the bulkhead and the first suction section of the centrifugal blade. Sucked into. Since the flow changes smoothly, there is little pressure loss in this space, and the performance of the turbo vacuum pump can be improved.

  A turbo type vacuum pump 1 according to a fifth aspect of the present invention is the turbo type vacuum pump according to the third or fourth aspect, wherein the centrifugal blade 24 penetrates the centrifugal blade 24 as shown in FIG. The turbine blade 70 is fixed to the intake portion side end face 15 of the rotation shaft 21.

The centrifugal impeller is because it is fixed to the rotary shaft passing through the centrifugal blade, it is possible to reduce the diameter of the boss portion of the centrifugal impellers. Moreover, since the flow of a radial direction can be produced and the flow path length can be lengthened, compression performance improves.
Turbine blades are fixed to the end surface of the rotary shaft on the intake side, so the diameter of the bosses of the turbine blades can be reduced, the centrifugal force acting on the bosses of the turbine blades can be reduced, and high-speed rotation is possible It becomes. As a result, even when a large flow rate gas is sucked in, the intake pressure can be reduced by the exhaust action of the turbine blades, and it can be compressed to a high pressure by the exhaust action of the centrifugal blades. Further, since the turbine blade and the centrifugal blade are separated from each other by this structure, when the rotor blade is partially damaged, deformed, corroded, etc., the rotor blade may be replaced. Since it is not necessary to replace the whole, a turbo type vacuum pump which is advantageous for long-term use can be obtained.

  The turbo type vacuum pump 1 according to the invention of claim 6 is the turbo type vacuum pump according to claim 5, which is disposed between the centrifugal blade 24 and the rotating shaft 21 as shown in FIG. Further, a circular ring 41 fitted with an interference fit is provided.

  It is arranged between the centrifugal blade and the rotating shaft, and is further equipped with a circular tube ring that is tightly fitted to the rotating shaft, so that the bending rigidity of the rotating shaft that is tightly fitted with the circular ring can be increased, and high-speed rotation is possible. It becomes. As a result, even when a large flow rate gas is sucked in, the intake pressure can be reduced by the exhaust action of the turbine blades, and it can be compressed to a high pressure by the exhaust action of the centrifugal blades.

  The turbo type vacuum pump 1 according to the invention of claim 7 is the turbo type vacuum pump according to any one of claim 1, claim 2 or claim 6, for example, as shown in FIG. The ratio of the outer diameter of the rotating shaft 21 to the outer diameter of 41 is 75% or more.

  If comprised in this way, the bending rigidity of the rotating shaft by which the circular ring was tightly fitted can be improved more effectively, and high-speed rotation will be attained.

According to the present invention, a turbo type vacuum pump is configured by combining a turbine blade having a high exhaust efficiency on a relatively low pressure side and a centrifugal blade having a high exhaust efficiency on a relatively high pressure side. Can increase the exhaust efficiency. The centrifugal impeller is to exhaust gas in the radial direction, without increasing the axial length, it can lengthen the flow path length. Therefore, since the length of the rotating shaft portion to which the turbine blade and the centrifugal blade are attached can be shortened, the natural frequency of the entire rotor is increased, high-speed rotation is facilitated, and exhaust efficiency is high in a pressure range of 1 to 1000 Pa. It is possible to provide a turbo-type vacuum pump that can rotate a rotor blade at a higher speed, increase the flow rate and pressure, that is, increase the exhaust speed without increasing the pump blade diameter, and is advantageous for long-term use. .
Further, the centrifugal vanes and fixed to the rotary shaft passing through the centrifugal blade, it is possible to reduce the diameter of the boss portion of the centrifugal impellers. Moreover, since the flow of a radial direction can be produced and the flow path length can be lengthened, compression performance improves. In addition, when the turbine blade is fixed to the end surface of the rotating shaft on the intake section side, the diameter of the boss section of the turbine blade can be reduced, the centrifugal force acting on the boss section of the turbine blade can be reduced, and high speed rotation is possible. It becomes. As a result, the high flow rate gas can be low pressure by an exhaust effect of the turbine blade, it becomes possible to compress to a high pressure by the exhaust action of centrifugal impellers. Furthermore, since the turbine blade and the centrifugal blade are separated from each other due to this structure, when the rotor blade is partially damaged, deformed, corroded, etc., the rotor blade can be replaced. Therefore, it is possible to provide a turbo vacuum pump that is advantageous for long-term use.

  Embodiments of the present invention will be described below with reference to the drawings. In addition, in each figure, the same code | symbol is attached | subjected to the mutually same or equivalent member, and the overlapping description is abbreviate | omitted.

  FIG. 1 is a front sectional view showing a configuration of a turbo vacuum pump 1 according to a first embodiment of the present invention. Hereinafter, a description will be given with reference to the drawings. The turbo vacuum pump 1 (hereinafter referred to as “pump 1” as appropriate) is a vertical type, and includes an exhaust unit 50, a motion control unit 51, a rotary shaft 21, an exhaust unit 50, a motion control unit 51, and a rotary shaft 21. And a casing 53 for housing. The rotating shaft 21 is disposed vertically above and below the exhaust part side 21A on the exhaust part 50 side, the motion control part side part 21B on the motion control part 51 side, the exhaust part side part 21A, and the motion control part side part 21B. And a disk-shaped large-diameter portion 54 therebetween.

  The casing 53 is disposed between the upper housing 23 (pump stator) 23, the lower housing 37 disposed below the upper housing 23 in the vertical direction (the axial direction of the pump 1), and the upper housing 23 and the lower housing 37. A sub-casing 40 is provided. The upper housing 23 has an intake nozzle 23A as an intake portion formed at the uppermost portion, and the sub casing 40 has an exhaust nozzle 23B as an exhaust portion formed on a side surface. The upper housing 23 accommodates the exhaust part 50 and the exhaust part side part 21 </ b> A on the exhaust part 50 side of the rotary shaft 21. An intake opening 55A is formed in the intake nozzle 23A, and an exhaust opening 55B is formed in the exhaust nozzle 23B. The intake nozzle 23A sucks gas as a fluid (for example, corrosive process gas or gas containing a reaction product) vertically downward from the intake opening 55A, and the exhaust nozzle 23B is sucked from the exhaust opening 55B. Exhaust gas horizontally.

  The exhaust part 50 includes a plurality of stages (five stages) of fixed blades 71 and 28, a plurality of stages (three stages) of turbine blades 73 having turbine blades 70 as rotating blades, and a plurality of stages (three stages). And a centrifugal blade (centrifugal drag blade) 24 as a rotating blade. The fixed vane 71 has three stages and is arranged on the immediately downstream side of the turbine blade 70, and the fixed vane 28 has two stages and is arranged on the immediately downstream side of the first and second stage centrifugal vanes 24. A centrifugal partition wall 43 serving as a partition wall is disposed on the upstream side of the first stage centrifugal blade 24, and the gas exiting the turbine blade 70 passes through the opening 43 A of the centrifugal partition wall 43 from the opening 43 A to the first stage centrifugal blade 24. Inhaled.

  The fixed vane 71 arranged on the flow side immediately after the final stage of the turbine blade 70 has an exhaust side 79 (FIG. 3) formed in a plane on the exhaust side, and the centrifugal partition wall 43 is an exhaust side 97 formed in a plane. (FIG. 3) is provided on the intake side, and a substantially hollow cylindrical space is formed between the exhaust side surface 79 and the exhaust side surface 97. The outer diameter of this space is formed approximately equal to the outer diameter of the final stage of the turbine blade 70.

  The centrifugal blade 24 is disposed away from the turbine blade 70 by an axial distance Lx. That is, between the exhaust side surface 79 of the stationary blade 71 arranged on the flow side immediately after the final stage of the turbine blade 70 and the front end surface 26A (FIG. 6A) on the intake side, which will be described later, of the first stage of the centrifugal blade 24. The axial distance is Lx. Further, the axial distance between the exhaust side surface 79 of the stationary blade 71 at the final stage of the turbine blade 70 and the exhaust side surface 97 of the centrifugal partition wall 43 is Ly.

  The exhaust part 50 includes a turbine blade part 73 having three stages of turbine blades 70. A hollow portion 12 is formed in the boss portion 74 of the turbine blade portion 73, and a through hole 58 is formed in the bottom portion 12 </ b> B of the hollow portion 12. The inner diameter of the hollow portion 12 is formed larger than the inner diameter of the through hole 58. The inner diameter of the through hole 58 is smaller than the outer diameter of the rotating shaft 21. A stepped portion 14 that protrudes from the end surface 11B is formed on the lower end surface (the end surface on the side opposite to the intake side) 11B of the turbine blade portion 73. The through hole 58 also penetrates the stepped portion 14.

  A concave portion 13 is formed in the intake portion side end surface 15 at the top of the rotating shaft 21, and a screw hole 18 is formed in the bottom portion of the concave portion 13. A turbine blade portion 73 is fixedly attached to the intake portion side end surface 15 with a hexagonal bolt 78 as a screw member, and the stepped portion 14 of the turbine blade portion 73 is engaged with the concave portion 13 of the rotating shaft 21. . The structure in which the stepped portion 14 is engaged with the concave portion 13 facilitates concentric alignment of the turbine blade portion 73 with respect to the rotating shaft 21, and allows the turbine blade portion 73 to be attached without causing an inclination with the center axis aligned. It is possible to prevent an increase in imbalance during high-speed rotation and to obtain stability during high-speed rotation. The hexagonal bolt 78 passes through the through hole 58 and is inserted into the screw hole 18. The inner diameter of the hollow portion 12 is formed to be slightly larger than the outer diameter of the head of the hexagon bolt 78, and is set to a value suitable for insertion and screwing of the hexagon bolt 78.

  The first-stage centrifugal blade 24 is disposed at a position away from the intake portion side end face 15 of the rotating shaft 21. In the figure, the number of hexagon bolts 78 is one, but a plurality of hexagon bolts 78 may be equally spaced from the axis.

  A circular tubular ring 41 is attached to the exhaust part side 21 </ b> A on the exhaust part 50 side of the rotating shaft 21 by shrink fitting. A fitting hole 25 is formed at the center of the centrifugal blade 24. The rotary shaft 21 to which the circular ring 41 is attached by shrink fitting passes through the fitting hole 25, and the centrifugal blades 24 are fixedly attached to the rotary shaft 21 by fitting, and are stacked one after another. The circular pipe ring 41 is located between the centrifugal blade 24 and the rotary shaft 21 in the radial direction of the rotary shaft 21. The circular ring 41 covers a portion where the three-stage centrifugal blade 24 is attached in the axial direction of the rotary shaft 21, and further, a portion extending from the portion where the centrifugal blade 24 is not attached to the intake portion side end face 15. Cover. A shaft sleeve 42 is attached to the outer side of the circular pipe ring 41 in the radial direction at a portion protruding from the centrifugal blade 24 of the circular pipe ring 41.

  2 (a) is a perspective view of the circular ring 41, FIG. 2 (b) is a partial perspective view of the rotary shaft 21 (showing the exhaust portion side 21A on the exhaust portion 50 side), and FIG. FIG. 3 is a partial perspective view of the rotating shaft 21 of FIG. 2B in which a circular pipe ring 41 is shrink-fitted. FIG. 2D is a partial perspective view of the rotating shaft 221 on which the circular tube ring is not shrink-fitted, and illustrates a portion corresponding to the rotating shaft 21 of FIG. The circular pipe ring 41 has an outer diameter D1 and an inner diameter D2. The outer diameter of the rotating shaft 21 is D3, and the outer diameter of the rotating shaft 221 is D1. Since D3> D2, the circular ring 41 is tightly fitted to the rotary shaft 21. The outer diameter of the circular pipe ring 41 that is shrink-fitted on the rotary shaft 21 is D1. The rotating shaft 21, the circular pipe ring 41, and the rotating shaft 221 have the same lengths as illustrated. Although not shown, the only difference between the rotary shaft 21 and the rotary shaft 221 is the presence or absence of a circular ring, and the same turbine blade and centrifugal blade are attached.

  Again, a description will be given with reference to FIG. As described above, the circular pipe ring 41 is shrink-fitted on the rotary shaft 21 of the pump 1 of the present embodiment at the portion where the centrifugal blade 24 is fixedly penetrated. By shrink fitting the circular pipe ring 41 to the rotary shaft 21, internal stress acts on the rotary shaft 21 and the circular pipe ring 41, and the bending rigidity (hereinafter referred to as rigidity) of the entire rotary shaft 21 including the circular pipe ring 41 is increased. The natural frequency increases. If this configuration of the present embodiment is adopted, the natural frequency of the entire shaft including the rotating shaft 21 having the outer diameter D3 (<D1) into which the circular ring 41 having the outer diameter D1 is shrink-fitted is Is higher than the natural frequency of the rotating shaft of the outer diameter D1 that is not shrink-fitted. (1) The outer diameter of the rotating shaft (circular ring and rotating shaft) at the position where the centrifugal blade 24 is fixed is D1. Yes, since it is the same as the outer diameter D1 of the rotating shaft 121 (FIG. 2) without the circular ring, the stress generated in the centrifugal blade 24 is the same under the same rotational speed. (2) Since the overall rigidity of the rotating shaft including the circular ring 41 is increased, it is possible to extend the rotating shaft, and between the turbine blade 70 at the final stage and the centrifugal blade 24 at the first stage. A sufficient axial dimension can be taken, and the exhaust performance of the turbine blade 70 can be improved.

If the outer diameter D3 of the rotating shaft 21 and the inner diameter D2 of the circular pipe ring 41 are made too small relative to the outer diameter D1 of the circular pipe ring 41 to be shrink-fitted, the circular pipe ring 41 becomes more than an improvement in internal stress due to shrink-fitting. The action (mass, moment of inertia) as the load mass acting on the rotating shaft 21 becomes larger. As a result, the rigidity of the entire rotating shaft including the circular ring 41 and the rotating shaft 21 cannot be improved. .
When the ratio (D3 / D1) is increased, the outer diameter D3 of the rotary shaft 21 itself and the 41 inner diameter D2 of the circular ring can be increased, so that the action (mass, moment of inertia) as the load mass of the circular ring 41 is small. made, but the thickness of the circular tube ring 41 shrink-fitting becomes thinner. If the wall thickness is too thin, the inner diameter stress during rotation of the tube ring 41 exceeds the allowable stress, and the tube ring 41 may be damaged.
The ratio (D3 / D1) of the outer diameter dimensions of the circular pipe ring 41 and the rotary shaft 21 is determined to be an optimum value depending on the material, shrinkage allowance, and the like. According to the calculation result, the ratio (D3 / D1) of the outer diameter D1 of the circular pipe ring 41 to be shrink-fitted and the outer diameter D3 of the rotating shaft 21 is preferably set to 75% or more. The maximum value of the ratio (D3 / D1) takes into consideration the outer diameter dimensions of the circular tube ring 41 and the rotating shaft 21, the respective materials, shrinkage allowance, the number of rotations, etc., so that the circular tube ring 41 is not damaged. As appropriate.

  By dividing the rotating shaft into the circular tube ring 41 and the rotating shaft 21, the material of the circular tube ring 41 can be a high Young's modulus material different from that of the rotating shaft 21. The material of the rotating shaft 21 is generally martensitic stainless steel, and the Young's modulus is about 200 GPa. If a high Young's modulus steel compounded with titanium boride particles is used as the material of the circular ring 41 (Young's modulus: 250 GPa or more), the rigidity of the entire shaft can be improved and the natural frequency of the rotor can be improved.

  The lower housing 37 houses the motion control unit 51 and the motion control unit side portion 21B on the motion control unit 51 side of the rotating shaft 21. The motion control unit 51 includes an upper protective bearing 35, an upper radial magnetic bearing 31, a motor 32 that rotationally drives the rotary shaft 21, a lower radial magnetic bearing 33, a lower protective bearing 36, and an axial magnetic bearing 34. It is configured to include in this order from the top to the bottom in the vertical direction. The upper radial magnetic bearing 31 and the lower radial magnetic bearing 33 support the rotary shaft 21 in a freely rotatable manner. The axial magnetic bearing 34 supports a force due to the weight of the rotating body in the downward direction in the figure and a thrust force in the vertical direction in the figure.

  Each of the magnetic bearings 31, 33, 34 is an active magnetic bearing. When an abnormality occurs in any of the magnetic bearings 31, 33, 34, the upper protective bearing 35 supports the rotary shaft 21 in the radial direction of the rotary shaft 21 instead of the upper radial magnetic bearing 31, and the lower protective bearing 36 Instead of the lower radial magnetic bearing 33 and the axial magnetic bearing 34, the rotary shaft 21 is supported in the radial direction and the axial direction of the rotary shaft 21.

The outer diameter dimensions of the turbine blade 70 and the centrifugal blade 24 will be described with reference to FIG.
The outer diameter of the second stage and third-stage turbine blades 70 are equal, smaller than the outer diameter of the first stage of the turbine blades 70. Further, all the turbine blades 70 are located closer to the intake portion side (intake nozzle 23A side) than the intake portion side end face 15 of the rotary shaft 21, and the lowermost stage (stage closest to the exhaust nozzle 23B) of the turbine blades 70, that is, three stages. The turbine blade 70 of the eye is located closer to the intake portion than the intake portion side end face 15 of the rotating shaft 21. If comprised in this way, the diameter of the boss | hub part of the turbine blade 70 of all the stages can be made small, the centrifugal force which acts on the boss | hub part of the turbine blade 70 of all the stages can be reduced, and high speed rotation will be more effective. Can be done automatically. The position of the turbine blade 70 located at the lowermost stage is closer to the axial intake section side than the end face on the intake section side is when the anti-intake side surface of the lowermost turbine blade 70 is axially the same position as the end face on the intake section side Including the original. The turbine blade 70 located at the lowest stage is located farthest from the intake portion.
It can be said that the outer diameter Dtmin of the third stage turbine blade 70 shown in the figure is the minimum outer diameter of the outer diameters of the turbine blade 70 whose axial position is closer to the intake portion side than the intake portion side end face 15. Outer diameter generally the turbine blade, the outer diameter of the bottom is equal to the minimum outer diameter of the outer diameter of the turbine blades.

  The first-stage to third-stage centrifugal blades 24 are formed to have the same outer diameter. The outer diameter Dgmax of the first stage (uppermost stage) centrifugal blade 24 shown in the figure is assumed to be the maximum outer diameter among the outer diameters of the centrifugal blade 24. That is, when the outer diameters of all the centrifugal blades are equal, the outer diameter is set as the maximum outer diameter. In general, the outer diameter of the multistage centrifugal blade is formed such that the outer diameter of the uppermost centrifugal blade has a maximum value.

  As shown in the figure, the minimum outer diameter Dtmin of the turbine blade 70 whose axial position is closer to the intake portion side than the intake portion side end face 15 is formed larger than the maximum outer diameter Dgmax of the centrifugal blade 24. If comprised in this way, the exhaust performance of the turbine blade which has the minimum outer diameter can be improved, and it can be set as the pump 1 which has high exhaust performance.

  The configuration of the turbine blade 73 (FIG. 1) will be described with reference to FIGS. FIG. 4A is a plan view of the turbine blade portion 73 as viewed from the intake nozzle 23A (FIG. 1) side. As for the turbine blade portion 73, only the first-stage turbine blade 70 is illustrated, and a hexagon bolt 78 (FIG. It is the figure which abbreviate | omitted 1). FIG. 4B is a diagram in which a view of the first stage turbine blade 70 seen radially toward the center is partially developed on a plane.

  The turbine blade portion 73 includes a boss portion 74 and a turbine blade 70, and the turbine blade 70 includes a plurality of plate-like blades 75 attached radially to the outer peripheral portion of the boss portion 74. The boss portion 74 is formed with a hollow portion 12 and a through hole 58. The blade 75 is attached with a twist angle twisted by β1 (for example, 10 to 40 degrees) from the central axis of the rotating shaft 21. The configuration of the second-stage and third-stage turbine blades 70 (not shown in FIGS. 4A and 4B) is the same as the configuration of the first-stage turbine blade 70, but the number of blades 75 and the number of blades 75 are the same. , The outer diameter of the portion of the boss portion 74 to which the blade 75 is attached, and the length of the blade 75 may be appropriately changed.

  With reference to FIGS. 5A, 5 </ b> B, and 5 </ b> C, the configuration of the first stage fixed blade 71 will be described. FIG. 5A is a plan view of the first stage fixed blade 71 as viewed from the intake nozzle 23A (FIG. 1) side. FIG. 5B is a diagram in which a view of the first stage fixed wing 71 seen radially toward the center is partially developed on a plane, and FIG. 5C is a cross-sectional view of FIG. It is X sectional drawing.

  The fixed wing 71 includes an annular ring part 76 and plate-like blades 77 attached radially to the outer periphery of the annular part 76. An inner peripheral portion of the annular portion 76 forms a shaft hole 60, and the rotation shaft 21 (FIG. 1) passes through the shaft hole 60. The blades 77 are attached with a twist angle twisted by β2 (for example, 10 to 40 degrees) from the central axis of the rotating shaft 21. The configuration of the second stage and third stage fixed wings 71 (not shown in FIGS. 5A, 5B, and 5C) is the same as the configuration of the first stage fixed wings 71. The number, the mounting angle β2 of the blade 77, the outer diameter of the annular portion 76, and the length of the blade 77 may be changed as appropriate.

  The configuration of the centrifugal blade 24 will be described with reference to FIGS. 6 (a) and 6 (b). 6A is a plan view of the first stage centrifugal blade 24 as viewed from the intake nozzle 23A (FIG. 1) side, and FIG. 6B is a front sectional view. The first-stage centrifugal blade 24 includes a substantially disk-shaped base portion 27 having a boss portion 61, and a spiral blade 26 fixed on a surface 27 </ b> A that is one surface of the base portion 27. The direction of rotation of the centrifugal blade 24 is the clockwise direction in FIG.

  The spiral blade 26 is composed of a plurality (six) of spiral blades as shown in FIG. The spiral blade 26 has a structure extending in the gas flow direction backward (opposite to the rotation direction) with respect to the rotation direction. The spiral blade 26 having the front end surface 26 </ b> A on the intake side extends from the outer peripheral surface 61 </ b> A of the boss portion 61 to the outer peripheral portion 27 </ b> C of the base portion 27. The other surface on the opposite side of the surface 27A is a back surface 27B, and the surface 27A and the back surface 27B are, for example, perpendicular to the central axis of the rotating shaft 21 (FIG. 1). The aforementioned fitting hole 25 is formed in the boss portion 61. The configuration of the second-stage and third-stage centrifugal blades 24 (not shown in FIGS. 6A and 6B) is the same as the configuration of the first-stage centrifugal blade 24, but the number of spiral blades 26, The shape and the outer diameter of the boss portion 61 may be appropriately changed in the length of the flow path formed by the spiral blade 26.

  In order to manufacture the centrifugal blade 24, a method of forming the spiral blade 26 having a convex shape protruding from the base 27 by machining such as end milling from a disk-shaped material (not shown) has a blade dimensional accuracy. From the viewpoint of improvement and the use of high specific strength materials (for example, aluminum alloys, titanium alloys, ceramics, etc.), this is the most common method for rotating blades that perform high-speed rotation (for example, a peripheral speed of 300 to 600 m / s).

The configuration of the first stage fixed wing 28 will be described with reference to FIGS. 7 (a) and 7 (b). FIG. 7A is a plan view of the fixed blade 28 viewed from the intake nozzle 23A (FIG. 1) side. 7 (b) is a front sectional view. The fixed wing 28 includes a fixed wing body 30 having an outer peripheral wall 62 and a side wall 63, and a spiral guide 29 protruding from one surface 63A of the side wall 63 and having a convex cross section. Centrifugal blade 24 (
Rotation direction of FIG. 1) is clockwise in FIG. 7 (a).

  The spiral guide 29 is composed of a plurality (six) of spiral guides as shown in FIG. The spiral guide 29 has a structure extending in the gas flow direction forward (same direction as the rotation direction) with respect to the rotation direction. The spiral guide 29 extends from the inner peripheral portion 62 </ b> A of the outer peripheral wall 62 of the fixed wing 28 to the inner peripheral portion 63 </ b> C of the side wall 63. An end surface 29A of the spiral guide 29 on a plane perpendicular to the central axis of the rotation shaft 21 is a smooth surface. A back surface 63B of the side wall 63 located on the opposite side of the spiral guide 29 is a flat and smooth surface. Therefore, the back surface 63B of the fixed blade 28 that directly faces the spiral blade 26 of the centrifugal blade 24 (FIG. 6) is in the direction 65 (FIG. 6A) formed between the spiral blades 26 of the centrifugal blade 24. It does not disturb the flow of gas flowing along the flow path. The configuration of the second stage fixed wing 28 (not shown in FIGS. 7A and 7B) is the same as the configuration of the first stage fixed wing 28, but the number and shape of the spiral guides 29 are appropriately determined. You may change it.

  FIG. 8 shows a turbo turbine in which a single stage turbine blade 170 (a structure similar to the turbine blade 70 in FIG. 4) and a single stage centrifugal blade 124 (a structure similar to the centrifugal blade 24 in FIG. 6) are attached to the rotating shaft 121. 3 is a partial schematic cross-sectional view showing a mold vacuum pump 101. FIG. On the downstream side of the turbine blade 170, a turbine fixed blade 171 (a structure similar to the turbine fixed blade 71 in FIG. 5) is arranged. The turbine blade 170 has blades 175 and an outer diameter of Dt, and the centrifugal blade 124 has spiral blades 126. The turbine blade 170 and the centrifugal blade 124 are spaced apart by an axial distance Lx. The axial distance Lx between the turbine blade 170 and the centrifugal blade 124 is from the downstream end surface of the proximal end portion of the blade 175 of the turbine blade 170 to the front end surface of the proximal end portion of the spiral blade 126 of the centrifugal blade 24. Is an axial distance (parallel to the center line of the rotating shaft 121). The turbine blade 170 in the figure corresponds to the final stage turbine blade 70 in FIG. 1, and the centrifugal blade 124 in the figure corresponds to the first stage centrifugal blade 24 in FIG. 1. The turbo vacuum pump 101 is configured to be able to measure the intake side pressure Ps and the exhaust side pressure Pd of the turbine blade 170 (both pressure units are Torr).

  FIG. 9 shows that the above-mentioned Lx of the turbo vacuum pump 101 (FIG. 8) is made variable, and Lx / Dt (FIG. 8) is a parameter (8, 10, 12, 15%) (including the case of a single turbine blade). ), The horizontal axis represents the exhaust side pressure Pd of the turbine blade 170 and the vertical axis represents Pd / Ps, and as an experiment in a single stage of the turbine blade, a centrifugal partition wall 143 is disposed in the downstream of the turbine blade, It is the performance graph which investigated the influence on blade performance. When the centrifugal blade 124 is brought close to the turbine blade 170 and the axial distance Lx becomes small, the gas exhausted from the turbine blade 170 collides with the centrifugal partition wall 143 on the upstream side of the first stage centrifugal blade 124 and smoothly It is not sucked into the centrifugal blade 124. The fact that Pd / Ps is 1 or less on the performance graph indicates that the gas flows backward to the turbine blade 170. By increasing the axial distance Lx, the influence of the centrifugal partition wall 143 is reduced, and the flow direction of the gas flowing in the axial direction from the final stage of the turbine blade 170 is the turbine blade 170 and the centrifugal blade 124. Since the gas is smoothly sucked into the centrifugal blade 124 while flowing through the space (the axial distance) between them, the performance of the turbine blade 170 is improved. It approaches the performance of a single turbine blade.

By increasing the rigidity of the rotating shaft and increasing the vibration eigennumber, the axial dimension Lx between the turbine blade 170 and the centrifugal blade 124 can be increased, and the length of the rotating shaft 221 can be extended. It becomes possible.
From the performance chart shown in FIG., The following results were obtained. That is, if Lx / Dt is ensured to be about 15% or more, the axial dimension Lx between the final stage turbine blade 70 and the first stage centrifugal blade 24 will be described below with reference to the turbo vacuum pump 1 of FIG. If about 15% of the outer diameter of the final stage turbine blade 70 is secured, the performance of the turbine blade single stage 70 is almost the same, and the influence of the centrifugal partition wall 43 is almost eliminated. The axial dimension Lx should be as long as possible, but Lx / Dt should be 12% or more in consideration of pump dimension restrictions and restriction of the natural frequency of the entire pump rotor that can ensure stable control of the magnetic bearing. It is desirable to do. In this case, in this embodiment, Lx is substantially is equal to Ly, Ly / Dt is about 12% or more. In this case, the axial distance between the stationary blade 71 on the flow side immediately after the final stage of the turbine blade 70 and the centrifugal partition wall 43 is 9% or more of the outer diameter of the final stage of the turbine blade 70. In the present embodiment, the sum of the axial width of the fixed blade 71 and the axial distance between the fixed blade and the final stage of the turbine blade 70 is 3% of the outer diameter of the final stage of the turbine blade 70. It is comprised so that it may correspond.

  Next, the operation of the turbo vacuum pump 1 will be described with reference to FIGS.

  As the first stage turbine blade 70 rotates, gas is introduced from the intake nozzle 23A of the pump 1 in the axial direction in FIG. By using the turbine blade 70, the exhaust speed can be increased, and a relatively large amount of gas can be exhausted. The introduced gas is decelerated by the fixed wing 71 and the pressure rises. Similarly, the second stage and third stage turbine blades 70 and fixed blades 71 exhaust in the axial direction, and the pressure rises.

  Next, the gas is introduced in the axial direction by the rotation of the first stage centrifugal blade 24. The gas introduced into the first-stage centrifugal blade 24 is caused by the interaction between the first-stage centrifugal blade 24 and the first-stage fixed blade 28, that is, the drag action due to the viscosity of the gas, and the centrifugal action caused by the rotation of the centrifugal blade 24. The gas is compressed and exhausted along the surface 27A of the base 27 of the first stage centrifugal blade 24 toward the outer diameter side of the first stage centrifugal blade 24.

  That is, the gas introduced into the first-stage centrifugal blade 24 is introduced into the centrifugal blade 24 in the substantially axial direction 64 in FIG. 6B, and between the spiral blades 26 of the first-stage centrifugal blade 24. Flows in the direction toward the outer diameter side through the flow path 68 formed in, and is compressed and exhausted. The direction of this gas flow is the direction 65 shown in FIGS. 6A and 6B, and this direction is the gas flow direction with respect to the first stage centrifugal blade 24.

  The gas compressed toward the outer diameter side by the first-stage centrifugal blade 24 flows into the first-stage fixed blade 28, and is substantially axial in FIG. 7B by the inner peripheral portion 62A of the outer peripheral wall 62. The direction is changed to 66 and flows into the space in which the spiral guide 29 is provided. As the first stage centrifugal blade 24 rotates, the drag action caused by the gas viscosity between the end surface 29A of the spiral guide 29 of the fixed blade 28 and the back surface 27B of the base 27 of the first stage centrifugal blade 24 causes the first stage centrifugal blade 24 to rotate. The gas is compressed and exhausted along the surface 63A of the side wall 63 of the fixed wing 28 (the surface on which the spiral guide 29 of the side wall 63 is attached) toward the inner diameter side of the first stage fixed wing 28. The gas that has reached the inner diameter side of the first stage fixed blade 28 is changed in the direction of the substantially axial direction 64 in FIG. 6B by the outer peripheral surface 61A of the boss portion 61 of the first stage centrifugal blade 24, and the second stage The centrifugal blade 24 is introduced. The same compression and exhaust are performed, and the gas is discharged from the exhaust nozzle 23B through the third stage centrifugal blade 24. The intake pressure is a low pressure region of 1 to 1000 Pa, and the exhaust pressure is a high pressure region of 100 Pa to atmospheric pressure.

  The rotor blades attached to the rotating shaft 21 (centrifugal blade 24, circumferential flow blade 88 described later) may be fitted to the outer diameter portion of the rotating shaft 21 by interference fitting or clearance fitting. As advantages in the case of clearance fitting, (1) assembly of the centrifugal blade 24 to the rotating shaft 21 is easy. (2) Even after the centrifugal blade 24 is assembled to the rotary shaft 21, any centrifugal blade can be removed. Thus, for example, when breakage, deformation, corrosion, or the like occurs in the overhaul, only the wing element that is severely damaged can be replaced. Advantages in the case of interference fitting are as follows: (1) The natural frequency of the entire rotating body increases due to the increase in rigidity of the rotating body (rotor) due to the interference fitting effect, and the margin of rotational speed control increases.

  According to the pump 1 of the present embodiment, the turbine blade portion 73 can be formed from an integral material, and is configured to be fixed to the intake portion side end surface 15 of the rotating shaft 21. That is, the conventional pump 1C includes a stator S that accommodates the rotating shaft 104 and a rotor R having a hollow portion 105a. The stator S is accommodated in the hollow portion 105a, that is, the rotor R is disposed outside the stator S. The pump has a structure different from that of the pump 1C (FIG. 16). In the conventional pump 1C, the rotation speed is limited by the centrifugal stress generated in the inner diameter portion of the hollow portion 105a. However, in the pump 1 of the present embodiment, the inner diameter dimension of the through hole 58 of the turbine blade portion 73 may be set to a value that is sufficiently large to allow the hexagon bolt 78 to pass therethrough. It is formed smaller than the outer dimensions of The inner diameter of the hollow portion 12 of the turbine blade 73 is also slightly larger than the inner diameter of the through hole 58 and is smaller than the outer dimension of the rotating shaft 21. Therefore, both the inner diameter of the through hole 58 and the inner diameter of the hollow portion 12 can be made significantly smaller than the inner diameter of the hollow portion 105a (FIG. 16), and the generated centrifugal stress can be greatly reduced. High speed rotation is possible.

  Since the centrifugal blade 24 has a structure in which the rotary shaft 21 passes through the fitting hole 25 formed in the center and is laminated on the rotary shaft 21, the inner diameter of the fitting hole 25 is set to the hollow portion 105a. Since the centrifugal stress generated in the inner diameter of the fitting hole 25 can be significantly reduced similarly to the turbine blade 70, the rotation speed can be increased. It was. Further, by adopting such a structure, it is possible to make a structure that flows in the axial direction and flows along the flow path along the direction 65 in the outer diameter direction, and the flow path length can be greatly increased. The exhaust performance, particularly the compression performance can be improved. Further, the fixed blade 28 can be configured to flow in the flow path in the inner diameter direction, and the gas exhausted by the fixed blade 28 can be flowed along the long flow channel 67 to reduce the flow velocity. Compression performance can be improved.

  Since the minimum outer diameter Dtmin of the turbine blade 70 whose axial position is closer to the intake portion side than the intake portion side end face 15 is formed larger than the maximum outer diameter Dgmax of the centrifugal blade 24, the turbine blade 70 having the minimum outer diameter is provided. Exhaust performance can be improved, and the pump 1 having high exhaust performance can be obtained. The ratio of the minimum outer diameter Dtmin of the turbine blade 70 and the maximum outer diameter Dgmax of the centrifugal blade 24 is preferably 1.2 or more. In this way, the exhaust performance of the turbine blade 70 having the minimum outer diameter can be further improved.

  The centrifugal blade 24 and the fixed blade 28 have a multi-stage structure, and the spiral blade 26 and the surface 27A of each centrifugal blade 24 and the spiral guide 29, the surface 63A of each fixed blade 28, and the inner peripheral portion 62A of the outer peripheral wall 62 are pivoted. Since it can be accessed from the direction, the processing of the centrifugal blade 24 and the fixed blade 28 is facilitated, and the manufacturing cost can be reduced.

  Since the turbine blade portion 73 is structured to be attached to the intake portion side end face 15 of the rotary shaft 21, the turbine blade 70 and the centrifugal blade 24 which are the rotary blades have a separate structure, and either the turbine blade 70 or the centrifugal blade 24 is provided. If the rotor is damaged, deformed, corroded, or the like, the rotor blade that has been damaged may be replaced, and the entire rotor need not be replaced. Therefore, the pump can be advantageous for long-term use. Further, since the centrifugal blades 24 have a multi-stage structure and each centrifugal blade 24 has a separate structure, when any one of the centrifugal blades 24 is damaged or corroded, the centrifugal blade 24 that has been damaged, corroded, or the like is generated. Since it is not necessary to replace the entire rotor, the pump can be advantageous for long-term use.

  In addition, since the rotor blade is disassembled into a plurality of components as described above, it is extremely unlikely that all rotor blades will be destroyed at the same time in the event of a rotor blade failure. The impact acting on the pump can be reduced, the possibility of the pump casing being destroyed is minimized, and the impact force on peripheral devices that are directly or indirectly coupled to the pump can be reduced. Pump.

  Since the inner diameter dimension of the through hole 58 of the turbine blade portion 73 is smaller than the outer dimension of the rotating shaft 21, the stress generated in the inner diameter portion of the through hole 58 can be reduced, and the hollow portion 12 is excessively large. Since it can be set as the structure where stress does not generate | occur | produce, high speed rotation is possible. The outer diameter of the rotating shaft 21 is desirably designed to be as thick as the inner diameter stress of the centrifugal blade 24 or a circumferential flow blade 88 described later (FIG. 13) allows to make the natural frequency of the entire rotating body as high as possible. Since the turbine blade portion 73 is fixed to the intake portion side end face 15 of the rotary shaft 21, even if the natural frequency of the rotating body is reduced by this structure, the natural frequency is sufficiently separated from the operating rotational speed range. The outer diameter of the rotating shaft 21 is determined so as to be. For this reason, the inner diameter of the through hole 58 formed for attaching the turbine blade 73 to the intake portion side end surface 15 of the rotating shaft 21 by the hexagon bolt 78 is formed smaller than the outer diameter of the rotating shaft 21.

  Since the turbo type vacuum pump 1 is configured by combining the turbine blade 70 having high exhaust efficiency on the low pressure side and the centrifugal blade 24 having high exhaust efficiency on the high pressure side as described above, the exhaust efficiency of the entire pump is increased. it can. Further, since the centrifugal blade 24 exhausts gas in the radial direction, the flow path length can be increased without increasing the axial length. Therefore, since the length of the rotating shaft portion to which the turbine blade 70 and the centrifugal blade 24 are attached can be shortened, the natural frequency of the entire rotor is increased, and high-speed rotation is facilitated.

  According to the pump 1 of the first embodiment, since the centrifugal blade 24 is fixed to the rotating shaft 21 that penetrates the centrifugal blade 24, the diameter of the boss portion 61 of the centrifugal blade 24 can be reduced. Moreover, since the radial flow can be generated by the centrifugal blade 24 and the flow path length can be increased, the compression performance is improved. Further, since the turbine blade portion 73 having the turbine blade 70 is fixed to the intake portion side end face 15 of the rotating shaft 21, the diameter of the boss portion 74 of the turbine blade portion 73 can be reduced. Centrifugal force acting on the boss 74 can be reduced, and high-speed rotation can be achieved. As a result, even when a large flow rate gas is sucked in, the intake pressure can be reduced by the exhaust action of the turbine blades, and it can be compressed to a high pressure by the exhaust action of the centrifugal blades. Furthermore, since the turbine blade 70 and the centrifugal blade 24 are separated from each other by this structure, when the rotor blade is partially damaged, deformed, corroded, etc., the rotor blade may be replaced. Since it is not necessary to replace the entire rotor blade, the pump 1 can be advantageous for long-term use.

  In the structure in which the turbine blade 70 and the centrifugal blade 24 are fixed to the rotary shaft 21 in series in the axial direction as in the pump 1 of the present embodiment, the axial direction between the final stage turbine blade 70 and the first-stage centrifugal blade 24 is used. If the dimension Lx is too small (narrow), the performance of the turbine blade 70 at the final stage deteriorates. The reason is as follows. A disc-shaped centrifugal partition wall 43 is disposed upstream of the rotating first-stage centrifugal blade 24 so as to form a small axial clearance. The centrifugal partition wall 43 has an opening 43A. The gas exhausted from the final stage turbine blade 70 passes through the opening 43A, is sucked from the inner diameter side of the first stage centrifugal blade 24, and is compressed from the inner diameter side to the outer diameter side by centrifugal force and drag action. . At this time, if the axial dimension Lx between the turbine blade 70 and the centrifugal blade 24 is small, the gas exhausted from the turbine blade 70 collides with the centrifugal partition wall 43 and is not smoothly sucked into the inner diameter portion of the centrifugal blade 24. . Or the gas which collided flows back into the turbine blade 70.

  In order to increase this axial dimension Lx and to solve this problem, it is necessary to increase the axial length of the rotating shaft 21, but this reduces the natural frequency of the rotor, and the magnetic bearing 31. , 33 and 34 are difficult to stably rotate. Further, in order to increase the axial length of the rotating shaft 21 and increase the natural frequency, it is necessary to increase the outer diameter of the rotating shaft 21 and increase the thickness of the rotating shaft 21. However, since the centrifugal blade 24 is fixed to the rotating shaft 21, the inner diameter of the penetrating portion of the centrifugal blade 24 increases with an increase in the thickness of the rotating shaft 21, leading to an increase in stress at the inner diameter portion, resulting in high-speed rotation. Cannot be realized.

  In the present embodiment, the axial dimension Lx is increased without increasing the outer diameter of the rotating shaft 21 (the inner diameter of the penetrating portion of the centrifugal blade 24) by shrink-fitting (tightening) the circular ring ring 41 to the rotating shaft 21. As described above, the gas flow between the last stage turbine blade 70 and the first stage centrifugal blade 24 is smoothed, and the natural frequency of the rotor is increased to realize high speed rotation. Therefore, it is possible to provide a turbo vacuum pump 1 that can obtain excellent pump exhaust performance.

  As shown in FIG. 10, the turbine blade portion 73 of the pump 1 has an annular convex portion 83 with which the rotating shaft 21 is engaged with the lower end surface (end surface of the intake side) 11 </ b> B of the turbine blade portion 73. May be. The inner diameter of the annular convex portion 83 is formed to be equal to the outer diameter of the rotating shaft 21. The annular convex portion 83 facilitates concentric alignment of the turbine blade portion 73 with respect to the rotating shaft 21 and allows the turbine blade portion 73 to be attached without causing an inclination with the center axis aligned, so that unbalanced during high-speed rotation. Can be prevented, and stability during high-speed rotation can be obtained.

  As shown in FIG. 11, the turbine blade portion 73 of the pump 1 has a convex portion 85 in which a male screw is formed on the lower end surface (anti-intake side end surface) 11 </ b> B of the turbine blade portion 73. You may make it have the recessed part 84 in which the internal thread 15 was formed in the air intake part side end surface 15 by screwing with the convex part 85. As shown in FIG. If comprised in this way, the turbine blade part 73 can be made into a solid structure, it is not necessary to form the through-hole for attaching the turbine blade part 73 to the intake-part side end surface 15 of the rotating shaft 21, and a turbine blade part The stress generated in the boss part 74 (FIG. 4) 73 can be reduced, and further high speed rotation is possible.

As shown in FIG. 12, the pump 1 may be configured to include holding rings 86 </ b> A, B, and C that press the centrifugal blade 24 toward the radial axis center side. A front stepped portion 87A is formed on the suction portion side of the boss portion 61 (FIG. 6) of the centrifugal blade 24, and a rear stepped portion 87B is formed on the exhaust portion side. The front stepped portion 87A of the first stage centrifugal blade 24 is held by the holding ring 86A, and the rear stepped portion 87B of the first stage centrifugal blade 24 and the front stepped portion 87A of the second stage centrifugal blade 24 are moved by the holding ring 86B. The rear stepped portion 87B of the third centrifugal blade 24 and the front stepped portion 87A of the third stage centrifugal blade 24 are held by the holding ring 86B, and the rear stepped portion 87B of the third stage centrifugal blade 24 is held by the holding ring 86C. The structure is such that it is pressed to the side, and the distance between the centrifugal blades 24 is determined by the holding ring 86B. If it does in this way, it can be set as the structure which can fix the centrifugal blade 24 firmly to the rotating shaft 21, can suppress that the imbalance of a rotary body increases rapidly during high speed rotation, and copes with high-speed rotation. can do.
In the present embodiment, the circular pipe ring 41 </ b> A that is shrink-fitted on the rotary shaft 21 does not have to reach the intake portion side end face 15 at the top of the rotary shaft 21. The circular ring 41 </ b> A is shrink-fitted into a portion of the rotating shaft 21 where the three-stage centrifugal blade 24 is attached (a portion where the rotation of the rotating shaft 21 is caused by rotation), and protrudes from the centrifugal blade 24 of the rotating shaft 21. it may be used as the structure is not Toritsui circle tube ring. The way to be able to improve the rigidity of the shaft (rotary shaft 21 and the round tubular ring 41A) which rotates with a short round tubular ring 41A.

  FIG. 13 is a front sectional view showing a configuration of a turbo vacuum pump 1-1 using a two-stage circumferential flow blade 88 in place of the centrifugal blade according to the second embodiment of the present invention. Hereinafter, differences from the description related to the turbo type vacuum pump 1 according to the first embodiment of FIG. 1 will be described.

  The exhaust section 50-1 includes a fixed blade 71 having a plurality of stages (three stages), a turbine blade section 73 having a turbine blade 70 as a rotating blade having a plurality of stages (three stages), and a plurality of stages (two stages). And a circumferential flow blade 88 as a rotating blade. The fixed blade 71 is provided only on the downstream side of the turbine blade 70. Septums 89 are respectively provided before and after the circumferential flow blade 88. Since the rotor blade includes the circumferential flow blade 88, a turbo vacuum pump having high compression performance can be obtained, and a turbo vacuum pump with high back pressure can be obtained. In particular, it is more effective if the centrifugal blade is a circumferential flow blade in a high pressure region.

  As shown in FIG. 14, the circumferential flow vane 88 includes a boss portion 91 formed with a shaft hole 93 through which the rotating shaft 21 (FIG. 13) passes, a disc portion 92 formed outside the boss portion 91, and a circular shape. It has the blade | wing 90 attached to the outer peripheral part of the board part 92 radially.

  As shown in FIG. 15, the partition wall 89 is formed inside the partition wall 89 for sucking the gas exhausted from the circumferential flow blade 88, and is a flow that guides the gas sucked from the suction port 94 in the circumferential direction. A path 96 and a discharge port 95 for discharging the gas guided to the flow path 96 to the circumferential flow blade 88 on the downstream side are provided.

  Since the pump 1-1 of the present embodiment includes the circumferential flow blade 88, the pump 1-1 having high exhaust performance can be obtained, and a turbo-type vacuum pump with high back pressure can be obtained.

It is front sectional drawing of the turbo type vacuum pump which concerns on the 1st Embodiment of this invention. (A) is a perspective view of a circular ring, (b) is a partial perspective view of a rotary shaft, (c) is a partial perspective view of a rotary shaft with a circular tube ring shrink-fitted, and (d) is a circle. It is a fragmentary perspective view of the rotating shaft in which the pipe ring is not shrink-fitted. It is a figure explaining the minimum outer diameter of the turbine blade of the turbo type vacuum pump of FIG. 1, and the maximum outer diameter of a centrifugal blade. 1A is a plan view of a turbine blade portion of the turbo type vacuum pump of FIG. 1, and FIG. 2B is a partially developed view of the turbine blade viewed radially toward the center. (a) is a top view of the stationary blade | wing for turbine blades of the turbo type vacuum pump of FIG. 1, (b) is the same front view, (c) is XX sectional drawing of (a). (a) is a top view of the centrifugal blade of the turbo type vacuum pump of FIG. 1, (b) is a front sectional view of the same. (a) is a top view of the stationary blade for centrifugal blades of the turbo type vacuum pump of FIG. 1, and (b) is a front sectional view of the same. FIG. 3 is a partial schematic cross-sectional view showing a turbo vacuum pump in which a single stage turbine blade and a single stage centrifugal blade are attached to a rotating shaft. It is a performance graph of the turbo type vacuum pump of FIG. It is a figure at the time of providing a cyclic | annular convex part in the anti-intake part side end surface of the turbine blade part of the turbo type vacuum pump of FIG. It is a figure at the time of providing a screw-shaped convex part in the anti-intake-part side end surface of the turbine blade part of the turbo type vacuum pump of FIG. It is a figure at the time of providing the holding ring which presses down a centrifugal blade in the turbo type vacuum pump of FIG. It is front sectional drawing of the turbo type vacuum pump which concerns on the 2nd Embodiment of this invention. (a) is a top view of the circumferential flow blade of the turbo type vacuum pump of FIG. 13, and (b) is a front sectional view of the same. It is a partial top view of the partition of the turbo type vacuum pump of FIG. It is front sectional drawing of the conventional turbo type vacuum pump.

Explanation of symbols

1, 1-1 Turbo type vacuum pump 11B End face 15 Intake part side end face 21 Rotating shaft 23 Intake nozzle (intake part)
24 Centrifugal blade (rotary blade)
41 yen tubular ring 43 centrifugal partition wall (partition wall)
43A opening 58 through hole 70 turbine blades (rotor blades)
73 turbine impeller part 78 hexagonal bolt 83 annular protrusion 84 recess 85 protrusion 88 circumferential flow blades (rotor blades)
Lx, Ly Axial distance

Claims (7)

An intake section for sucking gas in the axial direction; an exhaust section in which rotating blades and fixed blades are alternately arranged; and a rotating shaft for rotating the rotating blades;
One or more stages of turbine blades for exhausting the sucked gas in the axial direction, and one stage for exhausting the exhausted gas further by centrifugal drag action It is configured to include a centrifugal impeller described above;
The centrifugal impeller is secured to the rotary shaft passing through the centrifugal blade;
Wherein arranged between the centrifugal impeller between said rotary shaft, further comprising an interference fit has been round tubular ring to the rotating shaft;
Turbo vacuum pump. The turbine blades are fixed to the end surface of the rotary shaft on the intake side;
The turbo type vacuum pump according to claim 1. An intake section for sucking gas in the axial direction; an exhaust section in which rotating blades and fixed blades are alternately arranged; and a rotating shaft for rotating the rotating blades;
One or more stages of turbine blades for exhausting the sucked gas in the axial direction, and one stage for exhausting the exhausted gas further by centrifugal drag action It is configured to include a centrifugal impeller described above;
The axial distance between the first stage of the centrifugal blade and the final stage of the turbine blade is 12% or more of the outer diameter of the final stage of the turbine blade;
Turbo vacuum pump. An intake section for sucking gas in the axial direction; an exhaust section in which rotating blades and fixed blades are alternately arranged; and a rotating shaft for rotating the rotating blades;
One or more stages of turbine blades for exhausting the sucked gas in the axial direction, and one stage for exhausting the exhausted gas further by centrifugal drag action It is configured to include a centrifugal impeller described above;
A partition having an opening on the upstream side of the first stage of the centrifugal blade;
A first stage of the centrifugal blade is arranged to suck the gas from the opening;
Axial distance between the last stage of the turbine blades and the partition wall, is about 12% or more of the outside diameter of the final stage of the turbine blades;
Turbo vacuum pump. The centrifugal impeller is secured to the rotary shaft passing through the centrifugal blade;
The turbine blade, is fixed to the suction-part-side end face of the rotating shaft;
Turbo vacuum pump according to claim 3 or claim 4. Wherein arranged between the centrifugal impeller between said rotary shaft, further comprising an interference fit has been round tubular ring to the rotating shaft;
The turbo vacuum pump according to claim 5. Ratio of the outer diameter of the rotary shaft to the outer diameter of the round tubular ring is at least 75%;
The turbo type vacuum pump according to claim 1, claim 2, or claim 6.

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