阅读说明：本技术 涡轮分子泵 (Turbo molecular pump ) 是由 二木敬一 于 2020-01-08 设计创作，主要内容包括：本发明提供一种涡轮分子泵,其能够实现分子量小的氢气等的排气性能的提高。在涡轮分子泵的转子叶片(40)及定子叶片(30)的每一个,在圆周方向设置有多个扭曲叶片形状的浆叶(300、400),所述浆叶形成为放射状且内径侧叶片角度(θin)与外径侧叶片角度(θout)不同。关于叶片间距离(S)与浆叶长(b)的比即无因次参数(Xin)、无因次参数(Xc)、无因次参数(Xout),多层转子叶片(40)及多层定子叶片(30)的至少一者构成为满足以下条件中的任一个：第一条件“Xout＜Xc且Xin＜Xc”、第二条件“α·Xc Xin＞Xc＞Xout,其中α＝1.04”以及第三条件“Xin＜Xc＜Xout β·Xc,其中β＝1.04”。(The present invention provides a turbo-molecular pump, which can improve exhaust performance of hydrogen gas with small molecular weight, and the like, wherein a plurality of blades (300, 400) having twisted blade shapes are arranged in a circumferential direction in each of a rotor blade (40) and a stator blade (30) of the turbo-molecular pump, the blades are formed in a radial shape, and an inner diameter side blade angle (theta in) and an outer diameter side blade angle (theta out) are different from each other, regarding a ratio of an inter-blade distance (S) and a blade length (b), that is, a dimensionless parameter (Xin), a dimensionless parameter (Xc), a dimensionless parameter (Xout), at least one of a multi-layer rotor blade (40) and a multi-layer stator blade (30) is configured to satisfy any one of a first condition "Xout < Xc and Xin < Xc", a second condition "α. Xc Xin > Xc > Xout, wherein α" 1.04 "and a third condition" Xin < Xout < β. Xc < Xout, wherein β.04 "is 1.04".)

1. A turbo-molecular pump in which a plurality of layers of rotor blades and a plurality of layers of stator blades are alternately arranged in the order of rotor blades and stator blades from the side of a suction port of a pump, and

a plurality of blades having twisted blade shapes are provided in a circumferential direction in each of the rotor blade and the stator blade, the blades being formed in a radial shape and having an inner diameter side blade angle different from an outer diameter side blade angle, wherein

Regarding the ratio X of the blade length b, which is the width-direction dimension of the inclined surface of the blade, to the blade distance S, which is the circumferential distance between the blade and another blade adjacent in the circumferential direction, S/b, the value of the ratio X at the outer diameter-side end of the blade is Xout, the value of the ratio X at the inner diameter-side end of the blade is Xin, and the value of the ratio X at the intermediate position between the outer diameter-side end and the inner diameter-side end is Xc,

at least one of the multilayer rotor blade and the multilayer stator blade is configured to satisfy any one of a first condition "Xout < Xc and Xin < Xc", a second condition "α · Xc > Xout, where α ═ 1.04", and a third condition "Xin < Xc < Xout β · Xc, where β ═ 1.04".

2. The turbomolecular pump of claim 1, wherein

The multilayer stator blade is configured to satisfy any one of six conditions including a fourth condition "Xin < Xout < Xc", a fifth condition "Xin Xout < Xc", and a sixth condition "Xout Xin < Xc", in addition to the first condition, the second condition, and the third condition.

3. A turbo-molecular pump in which a plurality of layers of rotor blades and a plurality of layers of stator blades are alternately arranged in the order of rotor blades and stator blades from the side of a suction port of a pump, and

a plurality of blades having twisted blade shapes are provided in a circumferential direction on each of the rotor blade and the stator blade, the blades being formed in a radial shape and having an inner diameter side blade angle different from an outer diameter side blade angle; wherein

Regarding the ratio X of the blade length b, which is the width-direction dimension of the inclined surface of the blade, to the blade distance S, which is the circumferential distance between the blade and another blade adjacent in the circumferential direction, S/b, the value of the ratio X at the outer diameter-side end of the blade is Xout, the value of the ratio X at the inner diameter-side end of the blade is Xin, and the value of the ratio X at the intermediate position between the outer diameter-side end and the inner diameter-side end is Xc,

the multi-layered rotor blade and the multi-layered stator blade include: rotor blades and stator blades constituting an air suction section, rotor blades and stator blades constituting an intermediate section, and rotor blades and stator blades constituting an air discharge section,

the multilayer stator blade is configured to:

the suction segment satisfies a first condition "Xin < Xc < Xout",

the middle section satisfies a second condition "Xin < Xout < Xc",

the exhaust section satisfies any one of a third condition "α · Xc > Xout, where α ═ 1.04", a fourth condition "Xin Xout < Xc", and a fifth condition "Xout Xin < Xc".

Technical Field

The present invention relates to a turbo-molecular pump (turbo-molecular pump).

Background

The turbomolecular pump includes a multilayer stator vane (stator vane) and a rotor formed with a multilayer rotor vane (rotorvane). A plurality of blades (blades) formed in a radial shape are provided in the circumferential direction of the stator blade and the rotor blade. Each blade is inclined with respect to the horizontal direction, and the angle of inclination is referred to as the blade angle.

As the shape of the blade, a flat blade having a constant blade angle from the inner diameter side to the outer diameter side is known, but in the turbo molecular pump described in patent document 1, the blade angle of the blade is made maximum at the inner diameter portion and is continuously or intermittently decreased as the blade angle becomes the outer diameter in order to obtain high-efficiency exhaust performance.

[ Prior art documents ]

[ patent document ]

[ patent document 1] Japanese patent laid-open No. 2000-110771

Disclosure of Invention

[ problems to be solved by the invention ]

However, since the design is optimized to obtain sufficient exhaust performance in argon or nitrogen, there is a problem that sufficient exhaust performance cannot be obtained for hydrogen gas having a small molecular weight or the like. In particular, under conditions of high temperature, and further under conditions of high flow rate and high back pressure at high temperature, there is a disadvantage that the exhaust performance with respect to hydrogen gas or the like is significantly reduced.

[ means for solving problems ]

In a turbo-molecular pump according to a preferred embodiment of the present invention, a plurality of layers of rotor blades and a plurality of layers of stator blades are alternately arranged in the order of rotor blades and stator blades from a pump suction port side, and a plurality of twisted blade (twisted vane) shaped blades are provided in a circumferential direction in each of the rotor blades and the stator blades, the blades being formed in a radial shape and having an inner diameter side blade angle different from an outer diameter side blade angle, wherein, regarding a ratio X ═ S/b of a blade length b, which is a circumferential direction distance S between the blade and another blade adjacent to the blade in the circumferential direction, and a width direction dimension of an inclined surface of the blade, when the value of the ratio X at an outer diameter side end of the blade is Xout, the value of the ratio X at the inner diameter side end of the blade is Xin, and the value of the ratio X at an intermediate position between the outer diameter side end and the inner diameter side end of the blade is Xc, at least one of the plurality of layers of rotor blades and stator blades satisfies the conditions of "Xc < 3535 ≦ Xc < Xc > α ≦ Xc < 19 ≦ Xc < xxxv < 35 ≦ S > 2 ≦ Xc < xxxv.

In a more preferred aspect, the multilayer stator vane is configured to satisfy any one of six conditions including a fourth condition "Xin < Xout < Xc", a fifth condition "Xin Xout < Xc", and a sixth condition "Xout Xin < Xc" in addition to the first to third conditions.

In a turbo-molecular pump according to a preferred embodiment of the present invention, a plurality of layers of rotor blades and stator blades are alternately arranged in the order of rotor blades and stator blades from a pump suction port side, and a plurality of twisted blade-shaped blades are provided in a circumferential direction in each of the rotor blades and the stator blades, the blades being formed in a radial shape and having an inner diameter side blade angle different from an outer diameter side blade angle, wherein a ratio X of a blade length b, which is a circumferential direction distance S between the blade and another blade adjacent to the blade in the circumferential direction, and a width direction dimension of an inclined surface of the blade, is S/b, when the value of the ratio X at an outer diameter side end of the blade is Xout, the value of the ratio X at an inner diameter side end of the blade is Xin, and the value of the ratio X at an intermediate position between the outer diameter side end and the inner diameter side end of the blade is Xc, the plurality of layers of rotor blades and stator blades satisfies a condition of "Xc < 83, and a condition that" the stator blades satisfy a condition that "X < n < a fifth suction section and a stator blade < a" and "sub < a condition that" are satisfied "and" no < a condition that "Xc < 2, wherein" X < n < a condition is satisfied.

[ Effect of the invention ]

According to the present invention, the exhaust performance for hydrogen gas and the like having a small molecular weight can be improved.

Drawings

Fig. 1 is a sectional view showing an example of a turbomolecular pump.

FIG. 2 is a view of a first layer of rotor blades as viewed from the pump suction port side.

FIG. 3 is a schematic diagram illustrating blade design parameters.

Fig. 4(a) and 4(b) are views showing a part of the calculation results of the air discharge performance.

Fig. 5 is a diagram showing the performance improvement rate when the number of blades is set to 38 and the blade angle θ out on the outer diameter side and the blade angle θ in on the inner diameter side are changed.

Fig. 6(a) and 6(b) are graphs showing the relationship between the volume flow rate Qv and the pressure ratio Pr of the reference vane and the optimum solution (optimum solution) at the back pressure of 5 Pa.

Fig. 7(a) and 7(b) are diagrams showing the relationship between the volume flow rate Qv and the pressure ratio Pr of the reference vane and the optimum solution at the back pressure of 2 Pa.

Fig. 8 is a diagram showing changes in the dimensionless parameter X with respect to the optimal solution, the optimal solution candidate a, and the optimal solution candidate B, and the quasi-candidate F, and the quasi-candidate G.

Fig. 9 is a diagram showing changes in the dimensionless parameter X with respect to the optimal solution candidate C, the optimal solution candidate D, the optimal solution candidate E, and the quasi-candidate H.

Fig. 10 is a diagram showing changes in the dimensionless parameter X with respect to the quasi candidate I, the quasi candidate J, and the quasi candidate K.

[ description of symbols ]

1: turbo molecular pump

4 a: pump rotor

30: stator blade

40: rotor blade

300. 400: paddle blade

b: length of blade

S: space(s)

X, Xin, Xc, Xout: dimensionless parameter

θ, θ in, θ out: blade angle

Detailed Description

Hereinafter, embodiments for carrying out the present invention will be described with reference to the drawings.

First embodiment

Fig. 1 is a sectional view showing an example of a turbomolecular pump 1. In the present embodiment, a magnetic bearing type turbomolecular pump is described as an example, but the present invention is not limited to the magnetic bearing type, and various types of turbomolecular pumps can be applied. The turbomolecular pump 1 includes: a turbo pump stage including the stator blades 30 and the rotor blades 40, and a thread groove pump (threaded pump) stage including the cylindrical portion 41 and the stator 31.

In the example shown in FIG. 1, the turbopump segment includes 8 layers of stator blades 30 and 9 layers of rotor blades 40. In the thread groove pump segment, a thread groove is formed in the stator 31 or the cylindrical portion 41. The rotor blades 40 and the cylindrical portion 41 are formed on the pump rotor 4 a. The pump rotor 4a is fastened to a shaft (draft) 4b as a rotor shaft by a plurality of bolts 50. The rotary body 4 is formed by fastening the pump rotor 4a and the shaft 4b together with bolts 50.

The stator blades 30 are alternately arranged in a plurality of layers with respect to the rotor blades 40 arranged in the axial direction of the pump rotor 4 a. The stator blades 30 are stacked in the pump shaft direction via spacer rings (spacer rings) 33. The shaft 4b is supported in a noncontact manner by a magnetic bearing 34, a magnetic bearing 35, and a magnetic bearing 36 provided on the base 3. Although not shown in detail, each of the magnetic bearings 34 to 36 includes an electromagnet and a displacement sensor (displacement sensor). The floating position of the shaft 4b is detected by a displacement sensor.

The rotary body 4, which bolts the pump rotor 4a and the shaft 4b, is rotationally driven by the motor 10. When the magnetic bearings are not in operation, the shaft 4b is supported by emergency mechanical bearings (37 a, 37 b). When the rotor 4 is rotated at a high speed by the motor 10, gas on the pump intake side is sequentially exhausted through the turbo pump stage (rotor blades 40, stator blades 30) and the screw groove pump stage (cylindrical portion 41, stator 31) and is exhausted from the exhaust port (exhaust port) 38. A rear pump (back pump) is connected to the exhaust port 38.

Fig. 2 is a schematic view showing an example of the blade shape of the rotor blade 40, and is a view of the first-stage rotor blade 40 as viewed from the pump intake side. Each of the rotor blades 40 includes a plurality of blades 400 formed radially from the outer peripheral surface of the bell-shaped pump rotor 4 a. In general, the shape of the paddle 400 is: a flat blade having a constant blade angle from the inner diameter side to the outer diameter side of the blade 400, a twisted blade having a blade angle that varies depending on the position in the diameter direction of the blade 400, a tip blade having a blade width that decreases toward the tip, and the like. In the present embodiment, twisted blades are used.

The line indicated by the chain line 401 is a circle passing through the tip of the paddle 400, and the line indicated by the chain line 402 is a circle passing through the inner diameter side (near the root) of the paddle 400. The line indicated by the dashed-dotted line 403 is a circle passing through the middle position (average position) between the tip end and the inner diameter side of the paddle 400. Although not shown, each of the plurality of stator vanes 30 includes a plurality of blades 300 formed in a radial shape, as in the case of the rotor blade 40.

Fig. 3 is a schematic diagram illustrating blade design parameters of the stator blade 30 and the rotor blade 40, and shows a circumferential cross section (for example, a cross section along a dashed-dotted line 403 in fig. 2) of two adjacent blades 400 provided in the rotor blade 40.

The design parameters of the blade are as follows: a space S that is a circumferential interval (inter-blade distance) of the blades 400, a length (hereinafter referred to as a blade length) b from an intake side end to an exhaust side end of a slope of the blades 400, a blade height H that is a height of the blades 400 in the rotor shaft direction, a blade thickness t that is a thickness of the blades 400, a blade angle θ that is an inclination angle of the blades 400 with respect to a rotor shaft orthogonal surface, and a blade upper surface width W that is a circumferential width of a rotor shaft direction end surface of the blades 400. Regarding these blade design parameters, the parameter that most affects the exhaust performance is the ratio of the space S to the blade length b, i.e., the dimensionless parameter X ═ S/b (also referred to as the spacing code ratio).

The blade shapes used for the stator blade 30 and the rotor blade 40 are the flat blade, the tip blade, the twisted blade, and the like as described above. In the case of the flat blade, the blade length b of the blades arranged in the radial direction is set constant from the inner diameter side to the outer diameter side. Therefore, the dimensionless parameter Xin on the inner diameter side of the cross section along the dashed dotted line 402, the dimensionless parameter Xc on the average position (intermediate position between the inner diameter side and the outer diameter side) of the cross section along the dashed dotted line 403, and the dimensionless parameter Xout on the outer diameter side of the cross section along the dashed dotted line 401 in fig. 2 have the magnitude relationship as in the condition (1).

Xin < Xc < Xout … … Condition (1)

In general, in a turbo-molecular pump, vane design parameters are set in such a manner that the exhaust performance of argon or nitrogen is optimized (maximized). In this case, in the twisted blade having different blade angles on the inner diameter side and the outer diameter side, the following conditions are set as in the above condition (1): the dimensionless parameter X increases from the inner diameter side to the outer diameter side of the blade.

However, a turbo molecular pump designed to optimize the exhaust performance of argon or nitrogen has a problem that sufficient exhaust performance cannot be obtained for a gas having a small molecular weight such as hydrogen. Particularly, under conditions of high temperature, and further under conditions of high flow rate and high back pressure at high temperature, there is a problem that the exhaust performance for hydrogen gas and the like is significantly reduced.

In the present embodiment, in addition to the condition (1), the exhaust performance of a gas having a small molecular weight such as hydrogen can be further improved by setting the dimensionless parameter X of the stator blade 30 and the rotor blade 40 to the condition described below. Hereinafter, a hydrogen gas will be described as an example of a gas having a small molecular weight.

In optimizing the stator vane 30 and the rotor vane 40 according to the present embodiment, first, vane design parameters of the stator vane and the rotor vane are set so as to optimize the exhaust performance of argon gas or nitrogen gas. Hereinafter, the set stator blade and rotor blade are referred to as reference blades. Then, the vane design parameters of the reference vanes are changed, and the vane design parameters are optimized so as to improve the exhaust performance of the hydrogen gas. That is, optimization of the design parameters of the vane is achieved by using the reference vane as a base to improve the hydrogen gas exhaust performance.

Therefore, the following configuration that affects the exhaust performance except for the vane design parameters is set to the same conditions as the turbomolecular pump of the reference vane. That is, the number of rotor revolutions, the outer diameter of the rotor, the overall height of the turbo pump stage, the number of rotor blade layers, and the number of stator blade layers are set to the same conditions as those of the turbo molecular pump of the reference blade to be compared. The blade thickness t that affects the strength of the blade is also set to the same condition.

Under such precondition, considering the relationship between parameters such as H/b sin θ and t/W sin θ (see fig. 3), the dimensionless parameter X S/b is changed by changing the blade angle θ and the number of blades n.

(derivation of optimum conditions)

In the turbomolecular pump 1 shown in fig. 1, the turbopump section includes 8 layers of stator blades 30 and 9 layers of rotor blades 40. The turbo pump section is constituted by a suction section, an intermediate section, and a discharge section from the suction port side, and the vane design parameters of each section are different. Specifically, the 1 st to 2 nd layers (the 1 st layer of the rotor blade 40 and the 1 st layer of the stator blade 30) from the air intake side are air intake stages, the 3 rd to 6 th layers (the 2 nd to 3 rd layers of the rotor blade 40 and the 2 nd to 3 rd layers of the stator blade 30) are intermediate stages, and the 7 th to 17 th layers (the 4 th to 9 th layers of the rotor blade 40 and the 4 th to 8 th layers of the stator blade 30) are air discharge stages. The reference vane of the example has a rotor outer diameter of 304mm, which corresponds to a pump rotor of a turbomolecular pump of VG300 (Japanese Industrial Standards (JIS) standard).

Tables 1(a) and 1(b) show examples of the blade angles and the number of blades in the comparative examples and examples.

TABLE 1(a) reference vane (argon optimum)

TABLE 1(b) vanes of the embodiments (Hydrogen optimum)

Table 1(a) shows an example of the vane angle (inner diameter side vane angle, outer diameter side vane angle) and the number of vanes (the number of blades) of the stator vane 30 and the rotor vane 40 in the case where the reference vane of the vane design parameter is set so as to optimize the exhaust performance of argon gas. The blade shape is a twisted blade, and the blade angle on the inner diameter side is set to be larger than the blade angle on the outer diameter side. The blade angle θ is set so as to change at a constant rate in the radial direction. The blade angles and the number of blades of the air suction section, the intermediate section, and the air discharge section are set to values corresponding to the respective sections.

For example, the number of blades in the 2 nd to 3 rd layers of the rotor blade 40 and the 2 nd to 3 rd layers of the stator blade 30 constituting the intermediate stage is 36, the blade angle θ in on the inner diameter side is set to 50 degrees (deg), and the blade angle θ out on the outer diameter side is set to 30 deg. Although not described in table 1(a), since the blade angle θ is set so as to change at a constant rate in the radial direction as described above, the blade angle θ at the average position, which is the intermediate position between the inner diameter side and the outer diameter side, is set to 40deg ((50 deg +30 deg)/2).

On the other hand, table 1(b) shows the blade angle and the number of blades of the twisted blades according to the present embodiment. In the suction stage, the number of blades is reduced from 16 to 14 in the reference blade, and the blade angle on both the inner diameter side and the outer diameter side is made smaller than that of the reference blade. In the intermediate stage, the number of blades is increased from 36 to 38, and the blade angle is set to be the same as the reference blade. In the exhaust stage, the number of blades is increased from 34 to 38, and the blade angle on the inner diameter side is made larger than the reference blade. The blade angle on the outer diameter side is the same as the reference blade.

Table 2 shows an example of the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout in the examples.

TABLE 2

Kind of blade	Inner diameter side Xin	Average Xc	Outer diameter side Xout
				Air suction section	0.925	1.25	1.24
Middle section	0.960	1.18	1.19
				Exhaust section	1.06	1.12	1.04

Table 2 shows the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout when the blade angle and the number of blades are set as in table 1 (b). As shown in table 1(b), in the suction stage, the number of blades was adjusted to be reduced from 16 to 14 to increase the space S, and the blade angle θ was decreased to increase the blade length b. Thus, the values of the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout of the induction section are changed to values different from those of the induction section in table 1 (a).

In the intermediate stage, the number of blades was increased from 36 to 38, but the blade angle θ was set to the same value as in table 1(a), and therefore the space S decreased as the number of blades increased. As a result, the values of the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout are reduced as compared with the middle stage in table 1 (a).

In the exhaust stage, as shown in table 1(b), the number of blades was increased from 34 to 38, and the blade angle θ in on the inner diameter side was increased as compared with the case of table 1 (a). In this case, in the outer diameter side, the blade length b does not change, and the space S decreases. Therefore, the dimensionless parameter Xout (═ S/b) is reduced as compared with the case of table 1 (a). In the inner diameter side, since the blade angle θ in increases from 20deg to 23deg, the space S decreases, and the blade length b also decreases.

The blade design parameters (the number of blades and the blade angle) of each stage shown in table 1(b) were obtained by searching for an optimum solution in which the exhaust performance was improved as compared with that of the reference blade (in the case of table 1 (a)) by changing the blade angle and the number of blades based on the blade design parameters of the corresponding stage shown in table 1 (a). Hereinafter, the search results of the vane design parameters of the stator vanes 30 of the exhaust stage will be described as an example.

In the case of evaluating the exhaust performance of a single layer of the stator blade 30 caused by the change of the blade design parameter through the simulation, the exhaust performance is calculated using three stages including the changed stator blade 30, the rotor blade 40 disposed on the suction side of the stator blade 30, and the rotor blade 40 disposed on the exhaust side. As the vane design parameters of the rotor vanes 40 disposed on the intake side and the exhaust side of the stator vane 30, the vane design parameters described in table 1(a) as reference vanes were used. This enables the performance improvement with respect to the reference blade in the case of optimizing the stator blade 30 to be evaluated. Further, the criteria for blade temperature under high temperature conditions are: the stator blade temperature was 80 ℃ and the rotor blade temperature was 100 ℃. Therefore, in the case of the simulation calculation of the exhaust gas properties, the temperature is calculated to be 80 ℃ in the case of the stator vanes 30, and the temperature is calculated to be 100 ℃ in the case of the rotor vanes.

Fig. 4(a) and 4(b) are diagrams showing a part of the calculation results of the exhaust performance relating to the stator vanes of the exhaust stage, and the vertical axis represents the pressure ratio Pr and the horizontal axis represents the back pressure [ Pa ]. The exhaust performance is evaluated using a pressure ratio Pr which is a ratio of the exhaust side pressure to the suction side pressure (exhaust side pressure Pout/suction side pressure Pin). The flow rate of the discharged hydrogen gas was set constant at 300sccm (standard cc/min, 1atm), and the exhaust performance was determined at 3 points (3Pa, 5Pa, 8Pa) in the assumed operating pressure range of 2Pa to 8Pa with respect to the back pressure (pressure on the exhaust side of the three stages).

In fig. 4(a) and 4(b), the blade design parameter θ out, the blade design parameter θ in, and the number of blades of each line are represented as (15 to 20, 34). (15-20, 34 pieces) represent: the blade angle θ out on the outer diameter side was 15deg, the blade angle θ in on the inner diameter side was 20deg, and the number of blades was 34. Lines (15 to 20, 34) show the pressure ratio Pr in the case of the blade design parameters (blade angle and number of blades) described in the exhaust stage of table 1(a), and this line becomes a reference for searching the optimum solution.

Fig. 4(a) shows: lines (15-20, 34) of the reference blade, and lines (14-21, 34), lines (14-24, 38), lines (15-20, 36), lines (16-23, 38) and lines (14-23, 38) of the modified blade design parameter. Fig. 4(b) shows lines (15 to 20, 34) of the reference blade, lines (16 to 24, 38) of the modified blade design parameter, lines (15 to 24, 38), and lines (15 to 23, 38).

Fig. 5 shows the performance improvement rate when the number of blades is set to 38, and the blade angle θ out on the outer diameter side and the blade angle θ in on the inner diameter side are changed to various values. The performance improvement rate is a percentage of the pressure ratio Pr that is improved as compared with the line (15 to 20, 34) of the reference blade. In fig. 5, when the blade angle θ out and the blade angle θ in are expressed as the blade angle (θ out- θ in), the columns of the blade angles (12 to 23) are 1.3 to 7.7, for example. This indicates that the performance improvement rate is 1.3% at the minimum and 7.7% at the maximum, as compared with the line (15-20, 34) of the reference blade.

Fig. 5 also shows the performance improvement rate for combinations other than the blade angle (θ out — θ in) shown in fig. 4(a) and 4(b) when the number of blades is 38. As described later, the vane angles (15 to 23) are the optimal solutions in which the pressure ratio is the highest at the rate of increase from the reference vane (i.e., the rate of performance increase). Hereinafter, the leaf angles (θ out- θ in) denoted by the symbols a to E are referred to as optimal solution candidates near the optimal solution, and the leaf angles (θ out- θ in) denoted by the symbols F to K are referred to as quasi-candidates existing near the optimal solution candidates a to E.

Further, in the case of forming the stator vane by die casting (die cast), the vane angle θ in on the inner diameter side of the stator vane cannot be arbitrarily set depending on the inner diameter size of the stator vane. The inner diameter dimension of the stator vane also depends on the rotor diameter, so the smaller the rotor diameter, and the smaller the inner diameter dimension of the stator vane, the larger the blade angle θ in on the inner diameter side that can be machined. In the case of the stator blade of the present embodiment having the same inner diameter and outer diameter as those of the reference blade, when the number of blades is 38, it is difficult to machine the blade shape having the blade angle θ in on the inner diameter side of θ in22deg by die casting. However, since the simulation calculation can be performed even for a blade shape that is difficult to machine, the performance improvement rate is described with reference to the blade shape of θ in22deg in fig. 5.

Referring to fig. 4(a), 4(b), and 5, first, the pressure ratio Pr in the case where only the number n of blades is changed without changing the blade angle θ is compared. When the blade angles θ out and θ in are compared with the pressure ratio Pr in the case where the reference blades (15 to 20 and 34) are the same, it is understood that the pressure ratio Pr in the space S where X is smaller than S/b is larger by increasing the number of blades to one of 36 and 38 which is larger than the number of blades 34 of the reference blades. In addition, when the performance improvement rates in the case of the number of blades of 36 and 38 are compared in fig. 4(a) and 4(b), the performance improvement rate in the case of 38 is significantly improved.

When the performance improvement rate shown in fig. 5 is observed, the minimum value of the performance improvement rate is greater than 3 and the maximum value is increased to 8.5% to 9.0% in the vicinity of the blade angles (θ out- θ in) of (14-21), (14-22), and (15-22). On the other hand, as is clear from fig. 4(a) and 4(b), even when the number of blades is 34, the performance improvement rate is increased and the performance improvement rate is about (1% to 3%) in the case of the blade angle (14 to 21) as compared with the case of the blade angle (15 to 20). It is thus presumed that even when the number of vanes is changed, a peak in the exhaust performance appears in the vicinity of the same vane angle (θ out — θ in). Among these, the exhaust performance in the vicinity of the peak is higher for 38 pieces than for 34 pieces. In the case of the stator blade of the present embodiment, the blade machining becomes difficult if the number of blades is increased to 40 more depending on the relationship with the inner diameter of the stator blade, and therefore the upper limit of the number of blades is set to 38.

As can be seen from the above, the optimum solution is obtained in the case of the number of blades of 38 shown in fig. 5. In the case of forming the stator vane by die casting, as described above, in the case of the stator vane of the present embodiment, when the number of vanes is 38 in accordance with the inner diameter of the stator vane, it is difficult to perform machining when the vane angle θ in is 22deg or less. Therefore, an optimal solution is sought here for the blade angle θ in of the machinable θ in 23 deg. As a result, the optimal solution was obtained when θ out was 15deg, θ in was 23deg, and the number of blades was 38. Further, although the performance improvement rate is slightly smaller than the optimal solution, a sufficient performance improvement rate is obtained with respect to the blade design parameters (the blade angle and the number of blades) near the optimal solution.

In the example shown in fig. 5, the optimal solution is set to θ out of 15deg and θ in of 23deg within the range of θ in 23deg that can be processed, but for example, when the range of processing is θ in22deg, the optimal solution is set to θ out of 15deg and θ in of 22deg, or when θ out of 14deg and θ in of 22 deg. In this case, the search method described below can also be applied.

Fig. 6(a) and 6(b) and fig. 7(a) and 7(b) show the relationship between the volume flow rate Qv and the pressure ratio Pr in the case of the reference vanes (15 to 20, 34) and the optimum solution (15 to 23, 38). FIG. 6(a) shows the change of the pressure ratio Pr when the back pressure is fixed at 5Pa and the flow rate is changed between 100sccm and 500sccm, and FIG. 6(b) shows Pr (max) and qv (max) in this case. Further, fig. 7(a) shows the change of the pressure ratio Pr when the back pressure is fixed at 2Pa and the flow rate is changed between 100sccm and 500sccm, and fig. 7(b) shows Pr (max) and qv (max) in this case.

Note that the volume flow rate Qv on the horizontal axes in fig. 6(a) and 7(a) represents the flow rate of the gas having the same pressure as the suction-side pressure. Since the suction-side pressure in this case is different when the pressure ratio Pr is different, the flow rate measured in units of flow sccm is different when the pressure ratio Pr is different even when the value of the volume flow rate Qv is the same. In FIG. 6(a) and FIG. 7(a), a line representing a flow rate of 100sccm, a line representing a flow rate of 200sccm, and a line representing a flow rate of 500sccm are shown, respectively.

Here, 1.69 × 10 is added to the sccm^-3(Pa·m³/s), so if γ is 1.69 × 10^-3Then, y sccm is converted to y × γ (Pa · m)³Expressed as Pin (Pa), the flow rate is Qv × Pin (Pa m)³When the flow rate is equal to y sccm, the equation Qv × Pin is satisfied as y × γ, and Pr is Pout/Pin, the volume flow rate Qv becomes Qv y γ · (Pr/Pout).

In the case of a back pressure of 5Pa (═ Pout), a line of 100sccm is represented by an equation in which Pr ═ Pout/y γ) Qv ═ 5/0.169) Qv ═ 29.6 Qv. Similarly, a line of 200sccm is denoted as Pr of 14.8Qv, and a line of 500sccm is denoted as Pr of 5.92 Qv. In the case of a back pressure of 2Pa, a line indicating 100sccm indicates Pr ═ 2.37Qv (Pout/y γ) Qv ═ 11.8Qv, (a line indicating 200sccm indicates Pr ═ 5.92 Qv), and (a) a line indicating 500sccm indicates Pr ═ 2.37 Qv.

In fig. 6(a) and 7(a), lines 3921 and L are Qv-Pr lines estimated from the calculation data of the reference vane, lines 4622 and L are Qv-Pr lines estimated from the calculation data of the optimum solution, and the pressure ratio Pr at the point where the line L to L32 intersect the vertical axis is the pressure ratio when the volume flow rate Qv is zero, and indicates the maximum value Pr (max) of the pressure ratio, while the volume flow rate Qv at the point where the line L to L intersect the horizontal axis is the flow rate when the pressure ratio is 1, and indicates the maximum value Qv (max) of the volume flow rate Qv.

In the case of the back pressure 5Pa shown in fig. 6(a) and 6(b), pr (max) is 2.44 and qv (max) is 1.01 in the reference vane indicated by the line L21, and pr (max) is 2.60 and qv (max) is 0.99 in the optimal solution indicated by the line L22, and in the case of the back pressure 2Pa shown in fig. 7(a) and 7(b), pr (max) is 2.85 and qv (max) is 1.09 in the reference vane indicated by the line L31, and pr (max) is 3.12 and qv (max) is 1.06 in the optimal solution indicated by the line L32.

When the back pressure is 5Pa, the optimal solution is about 10% larger than the reference vane with respect to pr (max), and the reference vane shows a slightly larger value than the optimal solution with respect to qv (max). Under the actual flow rate of 100 sccm-200 sccm at the operating point, the optimal solution is about 6% better than the standard blade in performance. When the back pressure is 2Pa, the optimum solution is about 7% better than the reference blade at a flow rate of 100sccm to 200sccm at the operating point. Even when the flow rate is compared in the range of 100sccm to 500sccm, the blade shape of the optimal solution is higher than the performance of the reference blade when the back pressure is either 5Pa or 2 Pa.

Table 3 is a table showing an example of the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout.

TABLE 3

Table 3 shows the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout in the case where the blade design parameters are the reference blades (15 to 20 and 34), the optimal solutions (15 to 23 and 38) shown in fig. 5, the optimal solution candidate a to the optimal solution candidate E, and the quasi-candidate F to the quasi-candidate K.

Fig. 8 to 10 are diagrams showing changes in the dimensionless parameter X along the blade diameter direction, where the vertical axis shows the dimensionless parameter X (Xin, Xc, Xout), and the horizontal axis shows the blade diameter direction position, fig. 8 shows changes in the dimensionless parameter X concerning the optimal solution (line L) in which the blade angle θ in on the inner diameter side is 23deg, the optimal solution candidate a and the optimal solution candidate B (line L a, line L0B), and the quasi candidate F and the quasi candidate G (line L1F, line L G), fig. 9 shows changes in the dimensionless parameter X concerning the best solution candidate C, the optimal solution candidate D, and the optimal solution candidate E (line L C, line L D, and line L E), and the quasi candidate H (line 365H), and fig. 10 shows changes in the dimensionless parameter X concerning the quasi candidate I, the quasi candidate J, the quasi candidate L K, and the quasi-candidate H (line L K) in which the blade angle θ in on the inner diameter side is 25 deg.

On a line L0 showing the optimal solution (15 to 23, 38 pieces) in fig. 8, the dimensional relationship among the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout becomes "Xin < Xc and Xout < Xc", that is, the shape of the line L0 is a mountain line shape in which the average position is high and the inner diameter side and the outer diameter side on both sides are lower than the average position.

On the other hand, the optimal solution candidate a in which the blade angle θ in on the inner diameter side is 23deg, which is the same as the optimal solution, is obtained by changing the blade angle θ out on the outer diameter side of the optimal solution (15-23, 38 pieces) from 15deg to 14deg, and since the blade length b increases as the blade angle decreases, the blade lengths b on the outer diameter side and the average position increase in this case, and the dimensionless parameter Xout and the dimensionless parameter Xc decrease as a result, the inclination from the average position to the outer diameter side of the line L a becomes larger and the inclination from the average position to the inner diameter side becomes smaller as compared with the line L0, and further decreasing the blade angle θ out as the line L F of θ out 13deg results in the magnitude relationship of the dimensionless parameter X > Xc > Xout.

Conversely, when the blade angle θ out on the outer diameter side is increased from 15deg to 16deg as in the case of the line L B, the inclination of the line from the average position to the outer diameter side becomes smaller, and the inclination of the line from the average position to the inner diameter side becomes larger, and further, when the blade angle θ out is increased and set to θ out equal to 17deg as in the case of the line L G, the dimensional relationship of the dimensionless parameter X becomes Xin < Xc < Xout, contrary to the case where θ out is set to 13deg (in the case of the line L F).

In this case, when compared with the line L A of FIG. 8, in which θ out is 14deg as well, the dimensionless parameter Xin of the line L D, in which the blade angle θ in is larger, and the dimensionless parameter Xn, in which the dimensionless parameter Xc is larger than the dimensionless parameter Xin of the line L0A, the dimensionless parameter Xc. of the line L C, the line L E, and the line L H, in which L0, L B, and L G of FIG. 8 are the same as the blade angle θ out on the outer diameter side, the dimensionless parameter Xin, and the dimensionless parameter Xc become larger as compared with the line L, L G of FIG. 8, in which the blade angle θ out on the outer diameter side is the same, FIG. 8, line L G of FIG. 8 is Xin < Xout < Xc < Xout, but the relation of Xout < 9 is changed to Xout > L.

Further, when the vane angle θ in on the inner diameter side is increased to 25deg as shown by line L I, line L J, and line L K in fig. 10, the dimensionless parameter Xin and the dimensionless parameter Xc at the inner diameter side and the average position become larger than those of line L C, line L E, and line L H in table 3.

As described above, the relationship between the blade angles θ in and θ out and the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout of the line L and the line L a to the line L K shown in fig. 8 to 10 can be explained as follows, that is, if the blade angle θ out is reduced to 14deg and 13deg from the optimal solution (15-23, 38 pieces) of the table shown in fig. 5 to the upper side of the table, as shown in fig. 8, the relationship between the dimensions of the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout is "Xin > Xc" when θ out is 14deg, as in the case of 15deg, but changes to "Xin > Xc" when θ out is 13deg, whereas if the relationship between the optimal solution (15-23) to the table shown in fig. 5 is "Xin > Xc" when θ out is 15deg, the relationship between the dimensions of the dimensionless parameter Xc < 16deg "when θ out is 13deg, and the relationship between the dimensionless parameter Xout is" 16 deg.

Further, when the blade angle θ in on the inner diameter side is increased from 23deg to 24deg, the line L and the line L B of fig. 8 are compared with the line 3870C and the line L E of fig. 9, and when the line L and the line L B, and the line L C and the line L E are both "Xin < Xc and Xout < Xc", but when the line L C and the line L E having larger blade angles θ in are compared with each other, the difference between Xc and Xout is smaller, and the difference between Xc and Xout is larger, and when the line L G having larger blade angles θ out of 17deg is compared with the line L H, the magnitude relationship between xon, Xc, and Xout is changed from "Xin < Xout" to Xout "Xc < Xc" and Xc ". when the relationship between θ in and the line 632 is larger, the magnitude relationship between Xout is changed from" Xin < Xout "14" and Xout "L" and the magnitude of Xout "is larger.

As shown in fig. 5, it is understood that the performance improvement rate decreases as the distance from the optimal solution to the periphery (the vertical direction or the horizontal direction of the table) increases. For example, when the performance improvement rates of the optimal solution candidates a to G are compared under the condition that θ in is 23deg, the minimum value of the performance improvement rate is about 3 in the optimal solution candidates a to E, but the minimum value of the performance improvement rate is about 2 in the optimal solution candidates F and G adjacent to the outer peripheral sides thereof. When θ out is 12deg and 18deg, the minimum value decreases to about 1.

Here, considering a machining error during blade machining of a stator blade or a rotor blade, the minimum value needs to be about 2 in order not to make the minimum value of the performance improvement rate 1 or less due to the machining error, and since there is a relationship as shown in fig. 8 to 10 between the blade angle θ in and the blade angle θ out and the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout, when "Xin > Xc > Xout" as shown by a line L F in fig. 8 is taken as "Xin > Xc > Xout", the upper limit of Xin is taken as Xin in the case of the blade angle (13-23) shown in fig. 5, and when Xin at this time is taken as Xin α · Xc, the magnitude relationship of the dimensionless parameter Xc, and the dimensionless parameter Xout is taken as "α · Xc > Xout".

Further, as shown by a line L G in fig. 8, the upper limit of Xout in the case of "Xin < Xc < Xout" is Xout in the case of the blade angle (17-23) shown in fig. 5, when Xout at this time is expressed as Xout equal to β · Xc, the magnitude relation among the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout is expressed as "Xin < Xc < β · Xc", and in the example of the exhaust stage shown in fig. 5, α and β are about 1.03 to 1.04.

To sum up the above, the optimal solution candidates (including the optimal solution and the optimal solution candidates a to E) whose minimum value of the performance improvement rate is 2 or more satisfy the following conditions (2), (3), and (4). That is, by setting the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout to satisfy the conditions (2), (3), and (4), a turbo-molecular pump having excellent exhaust performance of hydrogen gas can be obtained. The optimal solution is set in consideration of whether or not the blade machining is possible. For example, in the example shown in fig. 5, the blade angles (14-21) and (14-22) are higher in the performance improvement rate than the blade angles (14-23) selected as the optimal solution, but in the case of the stator blade of the present embodiment, the blade processing is difficult except for this.

"Xin < Xc and Xout < Xc" … … condition (2)

"Xin < Xc < Xout β. Xc, wherein β ═ 1.04" … … conditions (3)

"α. Xc Xin > Xc > Xout, wherein α ═ 1.04" … … conditions (4)

In the description of fig. 4(a) and 4(b) to 10, the optimization of the stator vanes of the exhaust segment is described. Further, the optimum solution can be searched for the stator vanes in the intake stage and the intermediate stage by the same processing as in the case of the exhaust stage. Further, with respect to the stator vanes in the intake stage and the intermediate stage, by setting the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout to satisfy the conditions represented by the conditions (2) to (4), a turbo molecular pump having excellent exhaust performance for a gas having a small molecular weight such as hydrogen gas can be obtained. It is preferable that which of the conditions (2) to (4) is applied to each stator vane of the air suction stage, the intermediate stage, and the air discharge stage is appropriately selected in consideration of the vane height, the pressure condition, and the like of each stage.

Table 4 is a table showing the performance improvement rate of hydrogen gas in a single layer of the stator vane with respect to the respective optimum solutions of the suction stage, the intermediate stage, and the discharge stage.

TABLE 4

	H2(100sccm)	H2(300sccm)
			Air suction section	4.0～4.6％	4.2～8.5％
Middle section	3.4～5.7％	2.5～5.8％
			Exhaust section	1.7～4.6％	3.5～8.2％

Table 4 shows the respective optimum solutions for the intake stage, the intermediate stage, and the exhaust stage, and shows the performance improvement rate of hydrogen gas for a single layer of the stator vane. Further, a high temperature condition is assumed, and the temperature of the stator blade is set to 80 ℃. As is clear from Table 4, the performance improvement rate was increased at 200sccm with a larger hydrogen gas flow rate.

Table 4 shows the performance improvement rate for a single layer of the stator vane, but further performance improvement can be achieved by applying the optimal solution for each segment to the entire respective intake segment, intermediate segment, and exhaust segment.

Table 5 is a table showing the performance improvement rate of hydrogen in the case where the optimal solution is applied to all the layers of the stator vane.

TABLE 5

Table 5 shows the hydrogen performance improvement rate in the case where the optimal solution is applied to the entire layer of the stator vane. The greater the flow rate, the greater the rate of performance improvement. Further, the temperature of the stator vanes was calculated at 80 ℃.

In the above description, the performance improvement in the case of the rotor blade in which the reference blade is disposed above and below the stator blade of the optimal solution has been described, but the improvement in the hydrogen gas exhaust performance can be achieved by setting the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout to the conditions satisfying the conditions (2) to (4) even in the case where the rotor blade is set to the blade shape of the optimal solution and the stator blade is set to the blade shape of the reference blade.

In the case of applying the optimal solution to the rotor blade, it is not limited to hydrogen, but the performance improvement is also seen in argon.

Table 6 is a table showing the performance improvement rate of argon gas for a single layer of the rotor blade.

TABLE 6

	Ar(500sccm)
		Air suction section	3.3～7.5％
Middle section	3.4～11.5％
		Exhaust section	2.7～3.4％

Table 6 shows the performance improvement rate of argon gas in a single rotor blade layer (a simulation was performed with a three-stage configuration of stator blades in which reference blades are arranged above and below the rotor blade of the optimal solution). In the case of the rotor blade, the temperature was calculated with the temperature set to 100 ℃.

In the first embodiment described above, at least one of the multilayer rotor blades and the multilayer stator blades is configured to satisfy any one of the above-described conditions (2), (3), and (4), and thus, the exhaust performance under a high flow rate and a high back pressure condition in the hydrogen exhaust can be improved as compared with a conventional turbo molecular pump including the flat blades or twisted blades of the condition (1).

Second embodiment

In the first embodiment, it is described that the above-described conditions (2) to (4) are found as conditions for achieving the dimensionless parameter X of S/b that improves the exhaust gas characteristics of hydrogen gas. In the second embodiment, at least one of the multilayer rotor blade 40 and the multilayer stator blade 30 is configured to satisfy any one of the following six conditions, including the conditions (5) to (7) shown below in addition to the conditions (2) to (4).

Xin < Xout < Xc … … Condition (5)

Xin Xout < Xc … … Condition (6)

Xout Xin < Xc … … Condition (7)

With such a configuration, the exhaust performance under high flow rate and high back pressure conditions during hydrogen exhaust can be further improved, and the number of layers of the stator vanes 30 and the rotor vanes 40 can be reduced as compared with the first embodiment, thereby further reducing the size of the turbomolecular pump.

Third embodiment

In the third embodiment, the multilayer rotor blade 40 and the multilayer stator blade 30 include: the multi-layer rotor blade 40 and the multi-layer stator blade 30 are configured such that: the air suction section satisfies the existing condition (1), the intermediate section satisfies the condition (5), and the air discharge section satisfies any one of the conditions (4), (6), and (7).

Xin < Xc < Xout … … Condition (1)

Xin < Xout < Xc … … Condition (5)

"α. Xc Xin > Xc > Xout, wherein α ═ 1.04" … … conditions (4)

Xin Xout < Xc … … Condition (6)

Xout Xin < Xc … … Condition (7)

By dividing the three sections into the intake section, the intermediate section, and the exhaust section as described above and setting the optimum conditions for each section, the hydrogen exhaust performance under a large flow rate and high back pressure can be further improved as compared with the second embodiment.

Table 7 shows simulation results of exhaust performance in the case of the configuration of the second embodiment and the configuration of the third embodiment.

Formation of blade shape	Air suction pressure	Low rate of pressure reduction	Compression ratio
				Flat blade	1		1
Second embodiment	0.790	+21.0％	1.171
				Third embodiment	0.753	+24.7％	1.229

Table 7 shows the simulation results of the exhaust performance in the case of the configuration of the second embodiment and the configuration of the third embodiment, and the intake port pressure and the compression ratio under the condition that the hydrogen flow rate is 1500sccm are expressed by a ratio in which the exhaust performance of the conventional flat blade is 1. In the simulation, the exhaust performance was calculated for a configuration in which the number of stator vanes was 14, the number of rotor vanes was 15, and the total number was 29, but performance improvement was similarly achieved for other numbers of vanes.

As shown in table 7, the suction high pressure was reduced to 21.0% in the second embodiment and 24.7% in the third embodiment. In addition, the compression ratio is increased by 1.171 times in the second embodiment and 1.229 times in the third embodiment.

In the case of the second and third embodiments as well, similarly to the case of the first embodiment, the improvement of the hydrogen gas exhaust performance can be achieved even in the case where the above-described conditions relating to the dimensionless parameter Xin, the dimensionless parameter Xc, and the dimensionless parameter Xout are applied to the rotor blades 40 instead of the stator blades 30. Further, even when the above conditions are applied to all the layers of the stator blade and the rotor blade, the performance for hydrogen gas can be improved.

The present invention is not limited to the above-described embodiments, and other embodiments considered within the scope of the technical idea of the present invention are also included in the scope of the present invention. For example, although the magnetic bearing type turbomolecular pump having a screw groove pump section has been described as an example in the above embodiment, the present invention can be applied not only to the magnetic bearing type turbomolecular pump, but also to a turbomolecular pump having only a turbomolecular pump section without a screw groove pump section.