JP2014109193A - Centrifugal fluid machine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To maintain an operation range at a large flow-rate side to the same level of that of a conventional radial impeller while enlarging an operation range at a low flow-rate side more than that of the conventional radial impeller, in a centrifugal fluid machine.SOLUTION: A centrifugal fluid machine has a plurality of vanes arranged at a hub at intervals in the circumferential direction. When a vane angle distribution of the vanes is shown with a hub-side camber line connecting a hub-side front edge and a hub-side rear edge as a lateral axis, and a hub-side vane angle of the vane as a vertical axis, the hub-side vane angle becomes maximum in a position nearer the hub-side front edge than a center point of the hub-side camber line, and a distribution of the hub-side vane angles becomes protrusively convex up to a position in which the hub-side vane angle becomes maximum from the hub-side front edge. When the vane angle distribution is shown with a shroud-side camber line connecting a shroud-side front edge and a shroud-side rear edge as a lateral axis, and a shroud-side vane angle as a vertical axis, the shroud-side vane angle becomes minimum in a position near the shroud-side front edge than a center point of the shroud-side camber line.

Description

  The present invention relates to a centrifugal fluid machine having a centrifugal impeller, and more particularly to a blade shape of a centrifugal impeller.

  An example of a conventional centrifugal fluid machine is described in Patent Document 1. In the centrifugal compressor described in this publication, the blade angle of the blades of the impeller is set as follows in order to expand the operating range, improve the efficiency, and increase the peripheral speed of the impeller.

  That is, the blade angle in the blade shroud curve has a minimum value in the vicinity of the leading edge, increases toward the trailing edge, and reaches a maximum between the midpoint and the trailing edge of the shroud curve. On the other hand, the blade angle in the hub curve of the blade increases from the leading edge toward the trailing edge, and reaches a maximum value between the midpoint and the leading edge of the hub curve.

  In addition, as a design method related to the blades included in the impeller of the centrifugal fluid machine, those described in Patent Document 2 and Non-Patent Document 1 are known.

JP 2010-151126 A Japanese Patent No. 3693121

M. Zangeneh, A. Goto, and H. Harada: “On the Design Criteria for Suppression of Secondary Flows in Centrifugal and Mixed Flow Impellers”, ASME Journal of Turbomachinery, vol. 120, pp. 723-735, Oct., 1998

  In the centrifugal compressor described in Patent Document 1, the shroud side leading edge is in a state where the blade angle is minimized in the vicinity of the blade leading edge on the shroud side of the impeller and the impeller is viewed from the suction side in the axial direction. In the vicinity, the shape of the wing is closer to the circumferential direction. Therefore, the throat area, which is the minimum flow path cross-sectional area between two adjacent blades, decreases particularly on the shroud side. Therefore, the flow velocity in the vicinity of the throat increases and choke is likely to occur. When choke is generated, the operating range on the large flow rate side of the centrifugal compressor is narrowed.

  On the other hand, as described in Patent Document 2, when the hub side is inclined in the vicinity of the blade trailing edge of the impeller (near the impeller exit) so as to precede the shroud side in the rotational direction of the impeller, Non-Patent Document 1 As shown in Fig. 5, the operating range on the low flow rate side is narrowed.

  The present invention has been made in view of the above-mentioned problems of the prior art. The purpose of the centrifugal fluid machine is to increase the operating range on the large flow rate side while expanding the operating range on the low flow rate side compared to the conventional centrifugal impeller. The purpose is to maintain the same level as a conventional centrifugal impeller. And it aims also at improving the fluid efficiency of a centrifugal fluid machine, expanding the operating range of a centrifugal impeller in this way.

  A feature of the present invention that achieves the above object is that the blade angle distribution of the plurality of blades of a centrifugal fluid machine in which a centrifugal impeller having a plurality of blades arranged on a hub at circumferential intervals is attached to a rotating shaft. When the hub-side camber line connecting the hub-side front edge that is the suction-side end and the hub-side rear edge that is the discharge end is shown on the horizontal axis and the hub-side blade angle of the blade is shown on the vertical axis, The hub-side blade angle is maximum at a position closer to the hub-side front edge than the center point of the side camber line, and the hub-side blade angle distribution is convex upward between the hub-side front edge and the hub-side blade angle. Yes, the shroud-side camber line connecting the shroud-side front edge, which is the suction side end of the shroud side, which is the opposite side of the hub, and the shroud-side rear edge, which is the discharge end, is on the horizontal axis, and the blade shroud-side blade angle is on the vertical axis The center point of the shroud camber line In the section including the point where the shroud side blade angle is minimum at the position close to the shroud side front edge and the shroud side blade angle is minimum from the shroud side front edge, the blade angle distribution is convex downward. The portion on the shroud side trailing edge side from the blade angle distribution section is projected upward to the shroud side trailing edge.

  According to the present invention, the hub blade angle of the centrifugal impeller blades included in the centrifugal fluid machine is maximized between the blade leading edge and the flow direction intermediate point, and the blade angle is maximized from the blade leading edge. Since the blade angle distribution up to the point where there is a value is made not to have an inflection point, the operating range on the large flow rate side is expanded as compared with the conventional centrifugal impeller while expanding the operating range on the low flow rate side compared to the conventional centrifugal impeller. It can be maintained at the same level as the impeller. In addition, the fluid efficiency of the centrifugal fluid machine is improved while expanding the operating range of the centrifugal impeller.

It is a graph explaining the shape distribution of the blade | wing of the centrifugal impeller in one Example (Example 1) of the centrifugal fluid machine which concerns on this invention. FIG. 2 is an axial view of a centrifugal impeller blade having the blade angle distribution shown in FIG. 1. It is a figure explaining the definition of the shape of a centrifugal impeller. It is a figure explaining the speed triangle of the flow in an impeller. It is a figure explaining the shape distribution of the blade | wing of the centrifugal impeller in other Example (Example 2) of the centrifugal fluid machine which concerns on this invention. FIG. 5 is an axial view of the blades of the centrifugal impeller having the blade angle distribution shown in FIG. 4. It is a figure explaining the definition of the blade shape at the time of seeing a centrifugal impeller axially. It is a figure explaining the blade | wing shape distribution of the centrifugal impeller in further another Example (Example 3) of the centrifugal fluid machine which concerns on this invention. It is a figure explaining the flow which flows between two adjacent wing | blades of a centrifugal impeller. It is a figure explaining the flow which flows between two adjacent wing | blades of a centrifugal impeller. It is a graph which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a graph which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a figure which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a figure which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a graph which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a graph which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is a graph which shows an example of the result of having analyzed the flow in a centrifugal impeller. It is an axial view of the blades of the centrifugal impeller having the blade angle distribution shown in FIG. FIG. 13B is a meridional shape of a wing included in the centrifugal impeller shown in FIG. 13A. It is an axial view of the modification of the wing | blade with which the further another Example of the centrifugal impeller which concerns on this invention is provided. FIG. 14B is a meridional shape diagram of a wing included in the centrifugal impeller shown in FIG. 14A. It is a figure explaining the method to deform | transform the centrifugal impeller which concerns on this invention. It is a figure which shows the result of having analyzed the flow in the centrifugal impeller which has the blade | wing shown to FIG. 13A. It is a figure which shows the result of having analyzed the flow in the centrifugal impeller which has the blade | wing shown to FIG. 14A. It is meridional plane sectional drawing which shows a part of centrifugal fluid machine which concerns on this invention. It is meridional sectional drawing of the centrifugal fluid machine which concerns on this invention.

Hereinafter, some embodiments of a centrifugal fluid machine according to the present invention will be described with reference to the drawings.
First, a centrifugal fluid machine will be described with reference to FIGS. 17 and 18.

FIG. 17 is a meridional cross-sectional view of a part of the centrifugal fluid machine 200 according to the present invention.
FIG. 18 is a meridional cross-sectional view of the entire centrifugal fluid machine 200 including the portion shown in FIG. 17, and is an example of a single-shaft multi-stage centrifugal compressor.

  The centrifugal fluid machine 200 includes a centrifugal impeller 170 that imparts rotational energy to the fluid, a rotating shaft 174 to which the centrifugal impeller 170 is attached, and a radial outer side of the centrifugal impeller 170 that flows out of the centrifugal impeller 170. A diffuser 175 for converting the dynamic pressure of the fluid into static pressure. Downstream of the diffuser 175, a downstream flow path 178 for guiding fluid to the downstream centrifugal impeller 170 or the outside of the machine, and a return channel 176 for guiding fluid from the diffuser 175 to the downstream flow path 178 are provided.

  The impeller 170 is located between the hub 172 and the shroud 173 with a disc (hub) 172 fastened to the rotating shaft 174, a side plate (shroud) 173 disposed opposite the hub 172, and spaced in the circumferential direction. It has a plurality of wings 171 arranged. In addition, although the case where it has the shroud 173 is shown in FIG. 17, what is called a half shroud type impeller which does not have a shroud is also used.

  As shown in FIG. 17, the diffuser 175 includes a vane diffuser having a plurality of blades arranged at substantially equal pitches in the circumferential direction, and a vane-less diffuser not having a blade, which is not shown in FIG. Either one is used.

  FIG. 18 shows a multistage centrifugal compressor 200 in which the compression stages shown in FIG. 17 are stacked in a plurality of stages in the axial direction. Radial bearings 187 that rotatably support the rotating shaft 181 are disposed on both ends of the rotating shaft 181. A thrust bearing 188 that supports the rotating shaft 181 in the axial direction is disposed at one end of the rotating shaft 181.

  A plurality of compression stage impellers 180, which are four impellers 180 in FIG. 18, are fixedly attached to the rotary shaft 181. As shown in FIG. 17, a diffuser 182 and a return channel 183 are provided on the downstream side of each impeller 180. The impeller 180, the diffuser 182, and the return channel 183 are accommodated in the casing 190. A suction casing 184 is provided on the suction side of the casing 190, and a discharge casing 186 is provided on the discharge side of the casing 190. An inlet guide vane 185 is provided on the upstream side of the first stage impeller 180 to impart a pre-turn to the suction flow. The discharge casing 186 has a scroll 191.

  In the centrifugal fluid machine 200 configured in this manner, the fluid sucked from the impeller suction port 177 is increased every time it passes through the impeller 180, the diffuser 182, and the return channel 183 of each stage, and is compressed to the next stage compression stage. Sent to. The compressed fluid finally reaches a predetermined pressure and is discharged to the downstream flow path 178.

  By the way, in the centrifugal fluid machine 200, particularly in the centrifugal fluid machine that handles gas, the flow is stalled by the centrifugal impeller 1800 and the diffuser 182 as the flow rate decreases, and the pressure does not increase any more even if the flow rate is reduced. A phenomenon occurs that generates large pressure fluctuations and flow fluctuations. This is called surging and becomes a limit point on the small flow rate side of the centrifugal fluid machine 200. On the other hand, when the flow rate adjustment valve or the like is opened from the surging generation limit flow rate to increase the flow rate, a phenomenon occurs in which the discharge pressure decreases and the flow rate does not increase further. This is called a choke and becomes a limit point on the large flow rate side of the centrifugal fluid machine 200. The distance between these two limits, surging and choke, is called the operating range, and there is always a need to increase the operating range without reducing the efficiency of the centrifugal fluid machine. In the following, some examples of increasing the operating range without reducing the efficiency in the centrifugal fluid machine 200 according to the present invention will be described.

  A first embodiment of a centrifugal fluid machine 200 according to the present invention will be described with reference to FIGS. 1 to 3. FIG. 1 is a view showing a blade 20 shape distribution of an impeller 180 provided in the centrifugal fluid machine 200. The horizontal axis of FIG. 1 shows the length S of the blade center line (camber line) connecting points where the distances from the pressure surface and the suction surface of the blade 20 are equal to each other, respectively, for the hub end and the shroud end. This is the length S shown as non-dimensionalized by the total camber line length from the edge to the rear edge. The vertical axis represents the blade angle β (°) (see FIG. 2). The blade angle β at the hub side end of the blade 20 is indicated by a broken line, and the blade angle β at the shroud side end is indicated by a solid line.

  FIG. 2 is a view in which only one blade 120 included in the impeller 180 is taken out and viewed in the axial direction. A shroud side end of the blade 20 is indicated by a curve 24, and a hub side end of the blade 20 is indicated by a curve 23. In FIG. 2 and subsequent figures, the camber line is used as a representative curve of the blade 20. The front edge 21 that is the suction side end of the impeller 180 and the rear edge 22 that is the discharge side end of the impeller are straight.

The blade angle β is expressed as an inclination from the circumferential direction. For example, the blade angle β _{s at} the position of the radius R on the shroud side is expressed by a ratio between the circumferential small length R · dθ and the distance dm on the meridian plane. It is. Here, the distance dm on the meridian plane is defined as the points s ₁ and s ₂ on the assumption that the shroud side end 24 is changed from the point s ₁ to the point s ₂ between the circumferential small lengths R · dθ on the wing 20. This is the distance between points obtained by projecting on the meridian plane (RZ plane) of the impeller 180. Therefore, the blade angle β on the camber line between the points s ₁ and s ₂ is expressed by the following equation (1). In FIG. 2, N is the rotation direction, and O is the origin.

β = tan ⁻¹ (dm / (R · dθ)) (1)
Figure 3A is a phase adjacent two vanes A, an axial plan view at an arbitrary radial position B, the dashed line is a case the blade angle beta is greater and beta = beta _G, the solid line and beta = beta _S This is the case. The suction surfaces of the blades A and B are denoted by reference numerals 31A and 31B, respectively, and the pressure surfaces are denoted by reference numerals 32A and 32B. A perpendicular drawn from one blade A of two adjacent blades A and B onto the suction surface of the other blade B is defined as an inter-blade channel width L. Wings pipe passage width L, when the blade angle beta = beta _G is L _G, in the case of blade angle beta = beta _S is L _S.

FIG. 3B is a vector diagram showing the velocity triangle of the flow in the impeller 180. When the circumferential speed of the impeller 180 is U and the blade angle β is β _G , the relative speed of the flow in the impeller 180 is W, and the absolute speed of the flow in the impeller is C. If the blade angle beta is beta _S, the relative speed of the impeller 180 flows W ', the absolute velocity of the impeller in flow C' varies, Cm is the meridional direction component of the absolute velocity, the flow rate The associated velocity component.

Returning to Figure 1, the distribution curve 10 of the blade angle beta _s of the shroud side of the blade 20 shown in this embodiment, the minimum beta _{s_min} in leading edge s _{L_S,} increases as it goes downstream. The blade angle beta _s of the shroud side is convex down in the range of leading edge s camber line from _{L_s} length S _A, from the viewpoint of the camber line length S _A to the wing trailing edge s _{T_s} camber In the range of the line length S _B , the distribution is convex upward. Here, the camber line length S _A is less than the flow direction midpoint (dimensionless camber line length S = 0.5).

On the other hand, the distribution curve 11 of the blade angle beta _h of the hub side is maximum beta _{H_max} and between the leading edge s _{L_h} from the flow direction midpoint (dimensionless camber line length S = 0.5) s _m. And between the blade leading edge s _{L_h} and the blade angle β _{h_max} , the hub-side blade angle β _h distribution curve 11 has no inflection point. Thus, the reason for setting the shape of the wing 20 is as follows.

In FIG. 3A, the difference between the blade angles β _G and β _S appears as a difference in the shape of the velocity triangle in FIG. 3B. When the magnitudes of the meridional direction direction components C _m of the absolute flow velocities C and C ′ in FIG. 3B are set to be substantially the same at the same radial position, the magnitude of the relative velocity vector W ′ in the case of β _{S with} a small blade angle β. Is larger than the magnitude of the relative velocity vector W in the case of β _G where the blade angle β is large.

  In the ordinary centrifugal impeller 180, the shroud side relative flow velocity has a larger deceleration rate than the hub side relative flow velocity. Therefore, impeller efficiency and impeller stall characteristics determined from values such as wall friction loss and deceleration loss (loss due to the relative flow velocity decelerating and the wall boundary layer thickness increasing toward the downstream in the flow direction) This can be improved by appropriately setting the deceleration of the relative flow velocity on the shroud side.

The blade angle beta _s of the shroud-side to minimize the leading edge, as convex profile under the section of the camber line length S _A, wing in easily shroud side half deceleration rate is large wing 20 of relative flow velocity stall The increase in the angle β _s is suppressed, and the deceleration rate of the relative flow velocity is reduced. Thereby, the stall of an impeller can be suppressed to the lower flow rate side.

If so at the shroud side of the front edge of the wing 20 (camber line length S _A portion) does not decelerate the relative flow rates, areas of high relative velocity is increased from the blade leading edge 21 toward the flow direction downstream side. In a region where the relative flow velocity is high, wall friction loss is large, and an increase in the region where the relative flow velocity is high causes a decrease in impeller efficiency. Therefore, on the blade trailing edge 22 side on the shroud side (camber line length S _B portion), the blade angle β _s is made a convex distribution, the relative flow velocity is reduced, and the wall friction loss further increases. To prevent.

That is, in the blade leading edge side of the shroud side (camber line length S _A unit), prior to suppress an increase in blade angle beta _s in the vicinity of the rim 21, abruptly increased thereafter blade angle beta _s, the relative velocity Increase the deceleration rate. That is, in the region where the increase in the blade angle β _s is suppressed, the relative flow velocity becomes higher as shown in FIG. 3B, and the region with the higher relative flow velocity is expanded to the downstream side. As a result, impeller stall on the low flow rate side caused by a decrease in relative flow velocity is suppressed, and the efficiency of the impeller can be improved.

Incidentally, in the impeller of this embodiment, since suppressing the increase in the blade angle beta _s on the shroud side of the front edge (range of camber line length S _A), the shroud-side leading edge side (the camber line length S _A In the range), the inter-blade channel width L is narrowed as shown in FIG. 3A. Regarding the direction of the camber line length S, the inter-blade channel width L is the narrowest at the blade leading edge 21 and further shorter on the shroud 23 side than on the hub 24 side.

  In the inter-blade channel formed by two adjacent blades A and B, the portion having the narrowest channel cross-sectional area in the direction of the camber line length S is called a throat. If the Mach number of the relative flow velocity exceeds 1 at this throat, choke is generated and the flow rate cannot be increased any further. Therefore, in the large flow rate side operation of the centrifugal fluid machine in which the relative flow rate increases, the operating range is narrowed.

In order to avoid this problem, in this embodiment, the blade angle β _h on the hub side is changed from the blade leading edge (non-dimensional camber line length S = 0) to the non-dimensional camber line length S = 0.5. It set so that it might become the maximum ((beta) _{h_max} ) in between. Further, while the blade angle β _h is maximized from the blade leading edge 21, the distribution curve of the blade angle β _h on the hub side does not have an inflection point.

Thus, the blade angle β _h on the hub side between the throat that is often formed up to the dimensionless camber line length S = 0.5 and the blade leading edge 21 (dimensionless camber line length S = 0). Increases smoothly and rapidly. As a result, the blade angle β _{h_throat} on the hub side in the throat increases, and the inter-blade channel width L _h increases near the hub side of the throat. Since the inter-blade channel width L _h increases near the hub side, the throat area can be maintained even if the inter-blade channel width L _s narrows on the shroud side. Increasing the inter-blade passage width L _h of the hub-side does not have a inflection point hub-side blade angle distribution is achieved by a upwardly convex shape. As a result, the flow rate region in which the Mach number of the relative flow velocity exceeds 1 can be extended to the large flow rate side, choke generation in the impeller 180 can be suppressed, and the operating range on the large flow rate side of the centrifugal fluid machine can be secured.

In order to increase the hub side blade angle β _h at the throat, the maximum value β _{h —} max of the hub side blade angle is set as close to 90 ° as possible to the extent that no separation occurs on the hub side _surface of the blade 20. With such approach the maximum value beta _{H_max} the hub-side blade angle 90 °, the more that the maximum value beta _{H_max} the hub-side blade angle is greater than the outlet blade angle beta _{H_T} the hub side. Therefore, it is desirable to smoothly reduce the blade angle β _h distribution from the point at which the hub side blade angle reaches the maximum value β _{h —} max to the hub side outlet.

Further, it is desirable to be smaller than the previous blade hub side En径the R _{H_L} shroud side blade leading En径R _{S_L.} The reason for this is as follows.

In the centrifugal impeller, since the curvature radius ρ _s on the meridian surface on the shroud side is smaller than the curvature radius ρ _h on the hub side (see FIG. 17), the meridional surface direction flow velocity component at the blade leading edge 21 is large on the shroud side. So small. Therefore, if the hub-side blade leading edge _{diameter Rh_L} is substantially equal to the shroud-side blade leading edge _{diameter Rs_L} , the peripheral speed U of the impeller at the hub-side blade leading edge 21 becomes excessive. Further, since the meridional direction flow velocity is small at the hub-side blade leading edge 21, the fluid flowing into the hub-side blade leading edge 21 flows into the hub-side blade leading edge 21 from a more inclined direction.

On the other hand, in order to reduce the pressure loss generated at the blade leading edge 21, the relative inflow angle of the flow flowing into the blade leading edge 21 and the blade angle β of the blade leading edge 21 are made equal (state of no angle of attack). Is preferred. If the hub-side blade leading edge diameter _{Rh_L} and the shroud-side blade leading edge diameter _{Rs_L} are equal, it is ideal that the blade angle β at the hub-side and shroud-side blade leading edge 21 be the same. And Sankounisuru 1 will have to reduce the blade angle beta _{H_L} the blade leading edge 21 at the hub side, until the shroud side of the blade leading edge of the blade angle _β s_L ₌ β _{s_min.} As a result, the blade angle β increases abruptly from the blade leading edge 21, the amount of increase in the blade angle β becomes excessive, the flow does not follow the blade 20, and there is a risk that the efficiency will deteriorate due to separation. In order to avoid this problem, it is preferable to make the hub side blade leading edge diameter R _{h_L} smaller than the shroud side blade leading edge diameter R _{s_L} .

  Another embodiment of the centrifugal fluid machine 200 according to the present invention will be described with reference to FIGS. The present embodiment differs from the centrifugal fluid machine shown in the first embodiment in that the position of the minimum value in the blade angle distribution on the shroud side of the blades of the centrifugal impeller 180 is changed.

FIG. 4 shows an example of the blade angle distribution of the centrifugal impeller 180 according to the present embodiment. The hub-side blade angle distribution 41 is the same as in the first embodiment. In contrast, the blade angle distribution 40 on the shroud side once decreases from the blade leading edge s _{L_s} toward the downstream side in the flow direction, reaches a minimum value β _{s_min} , and then increases. Further, the blade angle between the shroud-side blade leading edge s _{L_s} and the blade trailing edge s _{T_s} is initially convex downward in the camber line direction and convex upward at the end. In Figure 4, convex below in the section S _C positioned on the upstream side of the midpoint Sm (camber line length S = 0.5), is convex upward in the section S _D following the period S _C However, the downwardly convex section Sc may exceed the intermediate point Sm.

The centrifugal impeller 180 of the present embodiment further reduces the reduction rate of the relative flow velocity near the shroud-side front edge of the impeller 180 compared to the centrifugal impeller 180 shown in the first embodiment. Thereby, the operating range on the low flow rate side of the impeller can be further expanded. The inter-blade channel width L is further smaller than the impeller shown in the first embodiment on the shroud side of the throat. Therefore, in this embodiment, the maximum blade angle β _{h —} max on the hub side is set to be equal to or greater than the value in the first embodiment in order to ensure the operating range on the large flow rate side of the centrifugal impeller 180. The maximum blade angle beta _{H_max} the hub side, because often greater than the hub side Tsubasa Ideguchi angle beta _{T_h,} between the position of maximum blade angle beta _h _ _max of the hub side to the hub side outlet S _{T_h} The blade angle is distributed so as to decrease smoothly.

  FIG. 5 is an axial view when only one blade 50 of the centrifugal impeller having the blade angle distribution shown in FIG. 4 is taken out. A camber line 54 on the shroud side of the blade 50 has a substantially S shape having a portion A5A in which the blade leading edge 51 is convex outward in the radial direction. On the other hand, the camber line 53 on the hub side of the blade 50 has a substantially S shape having a portion A5B in which the blade leading edge 51 side is convex inward in the radial direction. The reason for this will be described with reference to FIG.

  FIG. 6 is a coordinate system related to the centrifugal impeller 180 and is an axial view. FIG. 6 is a view seen from the suction side. The centrifugal impeller 180 rotates in the rotation direction N around the rotation axis O. In order to facilitate the explanation of the operation of the blades of the centrifugal impeller 180, the blade 60 having a straight blade camber line is taken up.

The blade angle at the leading edge 61 of the wing when the beta _L, is a diagram showing a blade angle beta at the position P ₆₂ on the downstream side of the front edge 61. The position P ₆₂ is separated from the blade leading edge 61 by Δθ in the circumferential direction, and the blade angle β at the position P ₆₂ is expressed by β = β _L + Δθ from a geometrical relationship. In a blade with a straight blade camber line, the blade angle β increases linearly with respect to the circumferential angle θ as it goes downstream from the blade leading edge 61.

The blade angle β does not change linearly with respect to the circumferential direction angle θ of the blade camber line, and when the blade angle β gradually increases from the position _P62 toward the downstream side, and the blade angle β increases. Consider the case where the amount increases. From the position P _62, increase the blade angle β is if smaller with respect to the circumferential direction an angle θ camber line, the shape of the camber line is a curve 63 in FIG. 6. That is, the contact with the straight camber line passing through the position P _62, is convex radially outward. On the other hand, when the increase amount of the blade angle β increases rapidly with respect to the circumferential angle θ of the camber line, the shape of the camber line is as shown by a curve 64 in FIG. That is, the contact with the straight camber line passing through the position P _62, is convex radially inwardly.

  In the centrifugal impeller 180 having the blade angle distribution shown in FIG. 4, the blade angle on the shroud side once decreases from the blade leading edge toward the downstream side and then becomes minimum, and then increases. As shown, a substantially S-shape is shown in which the blade leading edge 51 side is convex outward in the radial direction. In addition, the blade angle on the hub side is maximum without an inflection point between the leading edge 51 and the intermediate point in the flow direction, and smoothly decreases downstream from the maximum position. The camber line has a substantially S-shape that protrudes radially inward on the blade leading edge 51 side.

  Also in the present embodiment, for the same reason as in the first embodiment, it is desirable to make the blade leading edge diameter on the hub side smaller than the blade leading edge diameter on the shroud side.

  Still another embodiment of the centrifugal fluid machine 200 according to the present invention will be described with reference to FIG. This embodiment is different from the centrifugal fluid machine 200 shown in the first and second embodiments in that the inclination direction of the blade 70 at the blade trailing edge 72 of the centrifugal impeller 180 is inclined backward with respect to the rotation direction. . Thereby, when the centrifugal impeller 180 is viewed in the axial direction, the shroud side camber line and the hub side camber line of the blade intersect.

  FIG. 7 is an axial view when only one blade 70 of the centrifugal impeller 180 is taken out. At the blade trailing edge 72 of the blade 70, the shroud side camber line 74 is tilted backward with respect to the rotation direction from the hub side camber line 73. The blade angle distribution of the hub-side camber line 73 and the shroud-side camber line 74 is the same distribution as in the first or second embodiment.

Below, the effect | action of the centrifugal impeller 180 shown in a present Example is demonstrated.
FIG. 8A is a view of an impeller 180 (hereinafter also referred to as a forward inclined impeller) in which the camber line on the shroud side 83 is inclined forward relative to the camber line on the hub side 84 at the blade trailing edge 86. It is the figure which took out the two adjacent wing | blade 80 which forms. At the trailing edge 86 of the wing 80, the shroud side 83 is inclined forward with respect to the rotation direction with respect to the hub side 84.

  Thus, if the shroud side 83 of the blade 80 is tilted forward in the rotational direction from the hub side 84 at the blade trailing edge 86, the centrifugal stress acting on the blade 80 can be reduced. The blade force F acting on the fluid from each blade 80 acts in a direction perpendicular to the blade pressure surface 81, in other words, in the direction of the hub side 84 of the blade suction surface 82. Since the static pressure increases in the direction in which the blade force F acts, the static pressure increases on the hub side 84 of the blade suction surface 82. On the other hand, the static pressure decreases on the shroud side 83 of the blade suction surface 82.

  In the inter-blade flow path of the centrifugal impeller 180, a wall surface velocity boundary layer having a lower flow velocity and lower energy than the main flow velocity is generated near the wall surface. The fluid in the wall surface velocity boundary layer cannot overcome the static pressure gradient in the cross section between the blades, and flows from a region having a high static pressure to a region having a low static pressure. Here, the inter-blade channel cross section is a cross section obtained by cutting the inter-blade channel from the center of the rotation axis with a cylindrical surface having a radius r = constant. This flowing flow forms a secondary flow having a flow velocity component in a direction perpendicular to the main flow in the cross section between the blades.

  As described above, in the vicinity of the wall surface velocity boundary layer in the cross section between the blades of the centrifugal impeller 180, a secondary flow is generated from the blade pressure surface 81 having a high static pressure toward the blade negative pressure surface 82 having a low static pressure. In the forward inclined impeller, a secondary flow from the hub side 84 toward the shroud side 83 is also generated near the wall surface velocity boundary layer of the blade suction surface 82. Therefore, the low energy fluid accumulates on the shroud side 83 of the blade suction surface 82 and pressure loss increases. At the same time, the uniformity of the flow in the cross section between the blades is deteriorated, and the loss of the diffuser and the return channel portion on the downstream side of the impeller 180 is increased.

  FIG. 8B is a view of an impeller 180 (hereinafter also referred to as a backward inclined impeller) in which the camber line on the shroud side 83 is inclined backward relative to the camber line on the hub side 84 at the blade trailing edge 86. It is the figure which took out the two adjacent wing | blade 80 which forms. In the rearwardly inclined impeller, the blade force F acts in the direction of the shroud side 83 of the blade suction surface 82. Accordingly, the static pressure is reduced on the hub side 84 of the blade suction surface 82, and the static pressure is increased on the shroud side 83 of the blade suction surface 82. As a result, the secondary flow toward the shroud side 83 of the blade suction surface 82 can be suppressed, the uniformity of the flow in the cross section between the blades is improved, and the efficiency of the centrifugal impeller 180 is improved.

FIG. 9A is a diagram showing a pressure characteristic, which is a three-dimensional viscous fluid analysis result of a forward inclined impeller and a backward inclined impeller.
FIG. 9B is a diagram showing the efficiency characteristics of the forward tilt impeller and the rear tilt impeller whose pressure characteristics are shown in FIG. 9A, and is a three-dimensional viscous fluid analysis result. The forward inclined impeller and the backward inclined impeller shown in these drawings have substantially the same blade angle distribution as the impeller described in Patent Document 1 in the section of the related art, which is a conventional impeller. A thin line 91 is a case of a forward inclined impeller, and a thick line 92 is a case of a backward inclined impeller.

  The distribution of the blade angle β of the forward inclined blade impeller and the rear inclined blade impeller 180 is the same, and only the inclination direction of the blade trailing edge 86 is different. The three-dimensional viscous fluid analysis was performed on a single-stage centrifugal fluid machine in which the above-described two types of centrifugal impeller 180 were combined with a vaneless diffuser having the same shape and a return channel. The horizontal axis represents the flow rate ratio normalized with the design point flow rate of the forward inclined impeller as 1. The vertical axis represents the pressure ratio and efficiency ratio normalized with the design point pressure and design point efficiency of the forward inclined impeller as 1. It can be seen that both the pressure and efficiency near the design point of the rear inclined impeller are improved as compared to the forward inclined impeller.

FIG. 10A is a diagram illustrating an example of a flow velocity distribution at a design point of a forward inclined impeller. A cylindrical surface including the impeller blade trailing edge 102 is shown in an expanded manner, and is a non-dimensional radial flow velocity distribution CR / U ₂ in which the radial flow velocity distribution CR is made non-dimensional with the impeller outlet peripheral speed U ₂ . The flow path between blades between two adjacent blades is shown.
FIG. 10B is a diagram illustrating an example of a flow velocity distribution at a design point of the backward inclined impeller. It is a figure corresponding to FIG. 10A.

  In the forward inclined impeller, as shown in FIG. 10A, a low speed region (black portion) appears prominently near the blade suction surface 104 on the shroud 101 side, but in the backward inclined impeller, as shown in FIG. The basin is shrinking. On the other hand, as shown in FIG. 9A, the rearwardly inclined impeller has a pressure that rapidly decreases at a smaller flow rate than the forwardly inclined impeller, and the stall is accelerated. From the analysis results, it was found that the fact that the stall was accelerated in the centrifugal compression stage having the rearward inclined impeller was caused by the fact that the stall was accelerated in the rearward inclined impeller.

The reason why the stall in the backward inclined impeller is accelerated will be described with reference to FIG.
FIG. 11 is a graph showing the distribution of the dimensionless relative velocity W / U ₂ near the shroud blade surface in the forward inclined impeller and the backward inclined impeller, and is based on the analysis result at the specification point. The horizontal axis of FIG. 11 is the dimensionless camber line length S, the left end is the blade leading edge 85 and the right end is the blade trailing edge 86. Dotted lines 112 and 113 are for forward tilt impellers, and solid lines 110 and 111 are for backward tilt impellers. Lines 110 and 112 are relative flow rates at the suction surface 82, and lines 111 and 113 are relative flow rates at the pressure surface.

  In the rearwardly inclined impeller, the shroud side is inclined rearward in the rotational direction relative to the hub side at the blade trailing edge 86 of the centrifugal impeller 180 (see FIG. 8A), so that static pressure is applied to the hub side 84 of the blade suction surface 82. The static pressure rises on the shroud side 83 of the blade suction surface 82. Therefore, on the shroud side 83 of the blade suction surface 82, the amount of static pressure increase in the flow direction increases from the blade leading edge 85 toward the trailing edge 86. When the amount of increase in static pressure in the flow direction increases, deceleration of the relative flow velocity in the flow direction is promoted. In the rearwardly inclined impeller, the reduction rate of the relative flow velocity from the blade leading edge 85 increases particularly in the front half portion on the shroud side 83 of the blade suction surface 82 where the blade 80 is likely to stall (see the solid line 110 in FIG. 11). That is, the blade suction surface relative flow velocity 110 of the rearward inclined impeller has a larger decrease in the relative flow velocity in the front half of the blade 80 than the blade suction surface relative flow velocity 112 of the forward inclined impeller.

  A reverse pressure gradient is generated due to the static pressure rising from the blade leading edge 85 toward the blade trailing edge 86. However, when the relative flow velocity rapidly decreases near the surface of the blade 80, the reverse pressure gradient is overcome and the downstream side is overcome. Insufficient energy to flow into As a result, on the suction surface 82 on the shroud side 83 of the blade 80, the blade surface is likely to be peeled off in the front half portion near the blade leading edge 85, and the impeller stall related to the surging limit occurs on the larger flow rate side. . Therefore, in the rearwardly inclined impeller, as shown in FIG. 10B, the impeller stall is accelerated even though the accumulation of the low-energy fluid on the suction surface 104 on the blade shroud 101 side is suppressed. Normally, a low energy fluid is accumulated by the secondary flow and induces large-scale separation on the blade surface.

In view of these disadvantages of the prior art, the present embodiment employs the distribution of the blade angle β described in the first or second embodiment. That is, the blade angle β _s on the shroud side 24 is minimized at the blade leading edge 21 and then increased. At the same time, the blade angle β _s between the blade leading edge 21 and the blade trailing edge 22 on the shroud side 24 has a convex distribution downward on the blade leading edge 21 side and a convex distribution on the blade trailing edge 22 side. . The position of the maximum blade angle β _{h —} max on the hub side is located in the front half of the flow direction from the blade leading edge 21 and changed to the distribution of the blade angle β _h on the hub side 23 from the leading edge 21 to the position of the maximum blade angle β _{h —} max. The music point should not appear (see FIGS. 1 and 2).

Alternatively, the blade angle β _s on the shroud side 54 is once decreased to the minimum blade angle β _{s_min} from the blade leading edge 51 toward the downstream side in the flow direction, and then increased. At the same time, the blade angle β _s between the blade leading edge 51 and the blade trailing edge 52 on the shroud side 54 is convex downward on the blade leading edge 51 side and convex upward on the blade trailing edge 52 side. The position of the maximum blade angle β _{h_max} on the hub side 53 is the position from the blade leading edge 51 to the intermediate point in the flow direction, and the blade angle β _{h on} the hub side 53 is changed to the distribution upstream of the position of the maximum blade angle β _{h_max.} Make sure that no music points appear (see 4 and 5).

  By adopting one of the blade angle distributions described above, the blade shroud-side suction surface is generated by the impeller whose shroud side is tilted backward in the rotational direction from the hub side at the trailing edge of the impeller because the efficiency was improved by adopting one of the blade angle distributions described above. A sudden deceleration of the relative flow velocity at can be avoided. Furthermore, the operating range on the low flow rate side can be secured to the same level as that of the conventional impeller, and the operating range on the large flow rate side can be secured to the level of the conventional impeller.

  As described above, in the present embodiment, any blade angle distribution described in the first embodiment and the second embodiment is applicable. However, when a significant improvement in the efficiency of the centrifugal fluid machine is required, it is necessary to increase the degree that the shroud side tilts backward in the rotational direction from the hub side at the blade trailing edge. In such a case, in order to ensure the operating range on the low flow rate side, it is necessary to significantly reduce the deceleration rate of the relative flow velocity at the front half of the suction surface on the blade shroud side, which is described in Example 2. The distribution of the blade angle β is preferable because the deceleration rate of the relative flow velocity can be reduced.

FIG. 12A shows the pressure characteristics of a backward tilt impeller (hereinafter also referred to as a corrected post tilt impeller) having a blade angle β distribution described in the second embodiment and the conventional forward tilt impeller. The thin line 121 is the same as the thin line 91 shown in FIG. 9A, and the thick line 122 is a case of a tilted impeller after correction.
FIG. 12B is a diagram showing efficiency characteristics for the impeller whose pressure characteristics are shown in FIG. 12A. The thin line 123 is the same as the thin line 93 in FIG. 9B, and the thick line 124 is the efficiency curve for the corrected tilted impeller. Both FIGS. 12A and 12B are obtained from a three-dimensional viscous fluid analysis, and are a single-stage centrifugal type in which a vaneless diffuser and a return channel having the same shape are combined with a corrected inclined impeller impeller and a forward inclined impeller. About fluid machinery. The corrected tilted impeller is improved in both pressure and efficiency near the design point than the forward tilted impeller. In addition, the operating range on the low flow rate side and the operating range on the large flow rate side can be ensured to approximately the same extent as that of the conventional forward inclined impeller.

  By the way, in each said Example, the shape of the blades 20 and 50 which the centrifugal impeller 180 has is prescribed | regulated using the blade shape of the hub side end and shroud side end of the centrifugal impeller 180. FIG. In order to obtain a highly efficient centrifugal impeller, a three-dimensional blade shape is essential. Therefore, it is necessary to three-dimensionalize the blade shape defined by the hub side end and the shroud side end by appropriately connecting the hub side end and the shroud side end.

Several examples of connection methods between the hub side end and the shroud side end will be described below with reference to the drawings.
FIG. 13A is an axial view of a blade 130 having a linear element included in the centrifugal impeller 180. FIG. 13B is a meridional section view of a centrifugal impeller having blades 130 shown in FIG. 13A. Dashed line 136 shows an example of meridional streamlines. Ax indicates the center line of the rotation axis. The hub-side camber line 133 and the shroud-side camber line 134 are divided from the blade leading edge 131 to the blade trailing edge 132 at a plurality of points P _si , _Ph i (i = 0, n; n = total dividing points), The dividing points P _s i and P _h i are connected by a straight line. The connecting element 135 that connects the dividing points P _si and _Ph i is a linear element.

  Since the blades 130 of the centrifugal impeller 180 are defined by the assembly of the linear elements 135, when a centrifugal impeller is manufactured, machining with a cutting tool such as an end mill can be performed using a 5-axis NC processing machine or the like. If the side face of the end mill is positioned on the first linear element 135 and the cutting tool is smoothly moved along the hub-side camber line 133 and the shroud-side camber line 134 from the blade leading edge 131 to the blade trailing edge 132, 3 The dimensioned wing shape can be manufactured by NC processing, which improves manufacturability.

Another example of a method for connecting the hub-side camber line 143 and the shroud-side camber line 144 will be described with reference to FIGS. 14A and 14B.
FIG. 14A is an axial view of a blade 140 having a curved element 145 included in the centrifugal impeller 180. 14B is a meridional section view of a centrifugal impeller having blades 140 shown in FIG. 14A. Dashed line 146 shows an example of meridional streamlines. Unlike the impeller 180 shown in FIGS. 13A and 13B, the wing 140 is defined by an assembly of curved elements 145. As in the case of the linear element, the hub-side camber line 143 and the shroud-side camber line 144 are connected to a plurality of points P _s i, P _h i (i = 0, m; m = m = m) from the blade leading edge 141 to the blade trailing edge 142. divided by the total division points), to connect the dividing point _P s i, the _{P h} i to each other in the curve. A connecting element 145 that connects the dividing points P _si and _Ph i is a curved element.

Some specific examples of connection curves of the hub-side camber line 143 and the shroud-side camber line 144 will be described below.
First, a method for obtaining a curved element by deforming a linear element will be described. In the impeller 180 having the linear element blade 130 described in FIG. 13B, the meridional streamline 136 formed between the shroud side camber line 134 and the hub side camber line 133 is rotated around the rotation axis Ax of the impeller 180. By doing so, a rotating flow surface is obtained.

  FIG. 15 is a three-dimensional perspective view of the rotating flow surface 150. The blade section 151 is enlarged or contracted (reduced by Rd in this embodiment) in the chord S direction connecting the blade leading edge 152A and the blade trailing edge 153 of the blade section 151 formed on the rotating flow surface 150. A new blade profile 155 is obtained. In the blade section 155 of the present embodiment, the blade trailing edge 153 is fixed, the blade leading edge 152A is changed to the blade leading edge 152B, and the blade leading edge diameter 156A is changed to the blade leading edge diameter 156B.

  A plurality of linear elements 135 shown in FIG. 13B are divided in the blade height direction by a plurality of streamlines formed between the shroud-side camber line 134 and the hub-side camber line, and the rotary flow surface 150 is similarly formed. The obtained blade cross section 151 is similarly expanded or contracted in the chord direction to obtain a new blade cross section 155. This procedure is repeated, and a plurality of new blade sections 155 obtained are stacked in the blade height direction.

  The work (theoretical pressure increase) of the impeller 180 is maintained without changing the blade angle distribution β with respect to the blade trailing edge diameter 157 and the dimensionless camber line length S. At the time of stacking, the similar enlargement amount and the similar reduction amount are changed at each blade height position, and the blade sections obtained are stacked in the blade height direction. As a result, curved elements are formed.

  In the centrifugal impeller in which the blade leading edge shape becomes a curved shape on the meridian surface, particularly when the blade leading edge 146 is concave with respect to the suction port direction on the meridian surface shown in FIG. 14B. The inventors have found that the throat area increases. Therefore, the use of the curved element blade shown in FIGS. 14A and 14B can further improve the operating range on the large flow rate side.

  As another example of wing curve elementization, multiple rotating flow surfaces obtained by rotating meridional streamlines around the rotation axis are obtained in the blade height direction, and the blade cross section on each rotating flow surface is circumferential To obtain a new blade section, and stack the blade sections after the slide in the blade height direction. If the blades are curved elements in this way, it is possible to promote more uniform impeller outlet flow than straight element blades that are simply tilted back in the rotational direction at the trailing edge of the impeller blades than the hub side. There is sex.

FIG. 16A is a diagram showing a distribution of dimensionless radial flow velocity CR / U ₂ at the impeller exit of the linear element impeller.

FIG. 16B is a diagram showing a distribution of dimensionless radial flow velocity CR / U ₂ at the impeller exit of the curved element impeller.
In the impeller shown in FIG. 16B, the sliding amount in the circumferential direction of the blade cross section at each blade height position is adjusted so that the blade suction surface 164 has a convex shape with respect to the rotation direction. Both of these figures are the results of a three-dimensional viscosity analysis at the specification point. In the linear element wing of FIG. 16A, a low speed region (black area) appeared near the hub 160 side of the blade suction surface 164, but disappeared in the curved element wing of FIG. 16B.

  According to this embodiment, although the productivity is somewhat sacrificed, when connecting the hub end and shroud end of the blade, the likelihood of the connection method is increased, the blade surface can be defined by a free-form surface, and further throat It is possible to enlarge the area and control the precise secondary flow.

  According to each of the above embodiments, at least one of the blade angle distribution and the blade trailing edge shape on the hub side and the shroud side is changed, and a linear element and a curved element are used as blade elements connecting the hub side end and the shroud side end. Since either of these can be applied, the impeller can be further improved in performance. In the above embodiment, the case where two types of methods are used independently as the curve elementization has been described, but a curve elementization method combining these methods may be used. Further, in the third embodiment, for the same reason as in the first and second embodiments, it is desirable to make the hub side blade leading edge diameter smaller than the shroud side blade leading edge diameter.

  In each of the above embodiments, a single-shaft multi-stage centrifugal compressor has been described as an example, but it goes without saying that the present invention can also be applied to an end-stage centrifugal compressor. Moreover, although the centrifugal impeller in which the impeller has a shroud (side plate) has been described, the present invention can be similarly applied to an open type impeller that does not have a shroud (side plate).

  Summarizing the above embodiments, the position at which the shroud side blade angle is minimum may be the shroud side leading edge, and the hub side blade angle distribution does not have an inflection point in the section convex above the hub side blade angle distribution. Is essential.

  In addition, on the shroud side, which is the opposite side of the hub of a centrifugal fluid machine with a centrifugal impeller having a plurality of blades arranged on the hub at intervals in the circumferential direction, on the rotating shaft, the axial suction side of the rotating shaft The blade is characterized in that the section from the leading edge to the middle point in the flow direction is convex upward, and the blade is seen at the front end at the hub side as viewed from the axial suction side of the rotating shaft. The section from the edge to the midpoint in the flow direction should be convex downward.

  Further, if the rotation direction of the rotating shaft is rightward, the blade has an S-shaped shroud camber line and an inverted S-shaped hub-side camber line as viewed from the axial suction side, and the rotation direction is leftward. If it is present, it is more preferable if it has the opposite shape. It is preferable that the shroud side of the blade is inclined backward in the rotational direction from the hub side at the trailing edge which is the discharge end portion of the blade, and the shroud side camber wire and the hub side camber wire of the blade are on the axial suction side of the rotating shaft. From the viewpoint of crossing, it is more preferable if they intersect.

  Furthermore, the wing may be configured by smoothly connecting a plurality of linear elements, or the wing may be configured by smoothly connecting a plurality of curved elements. A plurality of the above centrifugal impellers may be attached to the same rotating shaft or may be a single stage.

  In addition, what was shown in each said Example is an illustration, Comprising: It is not limited. The scope of the invention is indicated in the appended claims, and all variations that come within the true spirit and scope of the invention are included within the scope of the claims.

DESCRIPTION OF SYMBOLS 10 ... Shroud side blade angle, 11 ... Hub side blade angle, 20 ... Centrifugal impeller blade, 21 ... Blade front edge, 22 ... Blade trailing edge, 23 ... Hub side camber wire, 24 ... Shroud side camber wire, 40 ... Shroud side blade Angle, 41 ... Hub side blade angle, 50 ... Centrifugal impeller blade, 51 ... Blade front edge, 52 ... Blade trailing edge, 53 ... Hub side camber wire, 54 ... Shroud side camber wire, 60 ... Centrifugal impeller blade, 61: Blade front edge, 62 ... Arbitrary blade position downstream of blade front edge, 63, 74 ... Camber wire, 70 ... Centrifugal impeller blade, 71 ... Blade front edge, 72 ... Blade rear edge, 73 ... Hub side Camber wire, 74 ... shroud side camber wire, 80 ... centrifugal impeller blade, 81 ... blade pressure surface, 82 ... blade suction surface, 83 ... shroud side, 84 ... hub side, 85 ... blade leading edge, 86 ... blade rear Edge, 100 ... Hub, 101 ... Shroud, 1 2 ... Blade trailing edge, 103 ... Blade pressure surface, 104 ... Blade suction surface, 110, 112 ... Relative speed near blade suction surface, 111, 113 ... Relative speed near blade pressure surface, 130 ... Centrifugal impeller blade, 131 ... Blade leading edge, 132 ... Blade trailing edge, 133 ... Hub side camber wire, 134 ... Shroud side camber wire, 135 ... Linear element, 136 ... Meridian streamline at arbitrary blade height position, 140 ... Centrifugal impeller Wings, 141 ... Wing leading edge, 142 ... Wing trailing edge, 143 ... Hub side camber wire, 144 ... Shroud side camber wire, 145 ... Linear element, 146 ... Meridion streamline at arbitrary blade height position, 150 ... Rotating flow Surface 151, blade cross section, 152 blade leading edge, 153 blade trailing edge, 154 blade chord, 155 blade reduced cross section in the chord direction, 160 hub, 161 shroud, 162 blade rear Edge, 163 ... Wings Force surface, 164 ... blade suction surface, 170 ... centrifugal impeller, 171 ... centrifugal impeller blade, 172 ... hub, 173 ... shroud, 174 ... rotating shaft, 175 ... diffuser, 176 ... return channel, 177 ... impeller suction 178 ... stage downstream flow path, 180 ... centrifugal impeller, 181 ... rotary shaft, 182 ... diffuser, 183 ... return channel, 184 ... suction casing, 185 ... inlet guide vane, 186 ... discharge casing, 200 ... centrifugal fluid Machine, C, C ′: Absolute flow velocity, Cm: Absolute flow velocity meridional direction component, CR: Absolute flow velocity radial component, F: Blade force, L: Interblade channel width, m: Meridional length, R ... radius, s ... dimensionless camber line length, s L _... leading edge, s _{L_S ...} shroud side blade leading edge, s _{L_S ...} hub-side blade front edge, s _m ... blade midpoint, s _T ... trailing edge, s _{T_s} Shroud side blade trailing edge, s _{T_s ...} hub-side blade trailing edge, U ... impeller peripheral speed, U 2 _... impeller outlet peripheral velocity, W, W '... relative velocity, beta ... blade angle, beta _{H_max} ... hub-side blade angle maximum , Β _L ... leading edge blade angle, β _{s_min} ... shroud side blade angle minimum value, θ ... angle, ω ... impeller rotation angular velocity.

Claims (11)

  The blade angle distribution of the plurality of blades of a centrifugal fluid machine in which a centrifugal impeller having a plurality of blades disposed in a circumferential direction at a hub is attached to a rotating shaft, When the hub-side camber line connecting the edge and the hub-side trailing edge, which is the discharge end portion, is shown on the horizontal axis and the hub-side blade angle of the blade is shown on the vertical axis, the hub is more than the center point of the hub-side camber line. The hub-side blade angle is maximized at a position close to the side front edge, and the distribution of the hub-side blade angle is convex upward until the position between the hub-side front edge and the hub-side blade angle is maximum, and is on the anti-hub side. When the shroud side camber line connecting the shroud side leading edge which is the suction side end on the shroud side and the shroud side trailing edge which is the discharge end is shown on the horizontal axis, and the shroud side blade angle of the blade is shown on the vertical axis, Shroud side of the center point of the shroud side camber line The shroud side blade angle is minimum at a position close to the edge, and the section including the point where the shroud side blade angle is minimum from the shroud side front edge has a downwardly convex blade angle distribution, and the convex blade angle distribution section below this The centrifugal fluid machine is characterized in that the shroud side blade angle distribution is convex upward to the shroud side rear edge at a portion closer to the shroud side rear edge.   2. The centrifugal fluid machine according to claim 1, wherein the position at which the shroud side blade angle is minimum is a shroud side leading edge.   3. The centrifugal fluid machine according to claim 1, wherein the hub side blade angle distribution does not have an inflection point in a section that is convex upward of the hub side blade angle distribution. 4.   The axial direction of the rotary shaft is on the shroud side which is the end opposite to the hub side of the blade of a centrifugal fluid machine in which a centrifugal impeller having a plurality of blades arranged in the hub at circumferential intervals is attached to the rotary shaft. A centrifugal fluid machine characterized in that a section from the leading edge to the intermediate point in the flow direction has a convex shape when viewed from the suction side.   5. The section of the blade according to claim 4, wherein when viewed from the axial suction side of the rotating shaft, the section from the front edge to the intermediate point in the flow direction is convex downward at the hub side end. Centrifugal fluid machine.   If the rotation direction of the rotating shaft is rightward, the blade has a S-shaped camber line on the shroud side and an inverted S-shaped camber line on the hub side when viewed from the axial suction side, and the rotation direction is leftward. 6. The centrifugal fluid machine according to claim 4 or 5, wherein if it exists, the reverse shape is obtained.   The centrifugal fluid machine according to any one of claims 4 to 6, wherein a shroud side of the blade is inclined backward in a rotational direction from the hub side at a trailing edge which is a discharge end portion of the blade. .   The centrifugal type according to any one of claims 4 to 7, wherein the shroud-side camber line and the hub-side camber line of the blade intersect each other when viewed from the axial suction side of the rotating shaft. Fluid machinery.   The centrifugal fluid machine according to any one of claims 1 to 8, wherein the blade is configured by smoothly connecting a plurality of linear elements.   The centrifugal fluid machine according to any one of claims 1 to 8, wherein the blade is configured by smoothly connecting a plurality of curved elements.   The centrifugal fluid machine according to claim 1, wherein a plurality of centrifugal impellers are attached to the same rotating shaft.

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