JP6411118B2 - Centrifugal impeller, single-shaft multistage centrifugal compressor using the same, and method of manufacturing centrifugal impeller - Google Patents

Description

  The present invention relates to the structure of a centrifugal impeller of a single-shaft multistage centrifugal compressor.

Centrifugal compressors used for processes require high efficiency and a wide operating range. When the centrifugal compressor is operated at a constant rotational speed, as shown in the characteristic curve of the compressor in FIG. 10, the operating range is determined by the occurrence of surge on the small flow rate side and the generation of choke on the large flow rate side. At this time, the operating range on the small flow rate side with respect to the design point flow rate is called a surge margin, and the operating range on the large flow rate side is called a choke margin.
Attempting to operate at a flow rate smaller than the flow rate at which this surge occurs (hereinafter referred to as the surge point) causes a back flow due to stall or separation inside the centrifugal compressor, making stable compressor operation difficult, There is also the possibility of damage.
Even if the discharge pressure is lowered to increase the flow rate exceeding the flow rate of choke (hereinafter referred to as choke point), if the flow reaches the speed of sound in the flow path inside the compressor (choke phenomenon), the flow rate will increase further. The flow rate cannot be increased.

Several techniques for expanding the surge margin and choke margin have been devised.
For example, in the technology disclosed in Patent Document 1, by devising the blade angle distribution from the blade inlet to the outlet, in other words, the blade load distribution, the surge margin of the impeller can be increased and at the same time the efficiency of the impeller can be improved. Aiming. According to this technique, an effect of expanding the surge margin on the small flow rate side can be expected. However, the choke margin expansion on the large flow rate side is not considered.

In order to increase the choke margin of the impeller, it is known to increase the area of the narrowest section of the impeller passage, that is, the throat area. Specifically, if the blade angle is devised to increase the throat area, the choke margin can be expanded.
However, since the blade angle distribution for increasing the surge margin and the angle distribution for increasing the choke margin are considered to have a trade-off relationship, it is difficult to greatly increase the choke margin by this method.

  Another method for enlarging the choke margin is to employ an intermediate blade disclosed in Patent Document 2. In this technique, the blades of the impeller are constituted by main blades having a long blade length and intermediate blades having a relatively short blade length, and are arranged alternately in the circumferential direction with the trailing edge aligned. As a result, the leading edge of the intermediate blade is downstream of the throat position when all the main blades are configured, so that the number of blades at the throat position is halved and the throat area can be widened. .

  Further, in Patent Document 3, in a pump turbine, between adjacent main blades, there is a blade having a length L1 along the camber line that is shorter than the length L0 along the camber line of the main blade by a predetermined ratio. An impeller is disclosed in which at least one intermediate blade is provided to reduce the load per main blade, and to suppress the reduction of the flow velocity in the vicinity of the blade surface, thereby suppressing boundary layer growth and flow separation. It is described that the efficiency can be improved even when the water level fluctuation is large, and the cavitation characteristics are improved.

JP 2010-151126 A JP 2011-117346 A JP 2000-54944 A

  According to the technique disclosed in Patent Document 1, the surge margin can be increased, and the technique disclosed in Patent Document 2 can increase the choke margin. However, each technology alone cannot expand a margin on both the small flow rate side and the large flow rate side to obtain a centrifugal compressor with a wide operating range.

Further, the type of impeller using the intermediate blade of Patent Document 2 is often used for an open impeller without a shroud of the impeller. In the single-shaft multistage centrifugal compressor, there is a problem that it is difficult to manage the blade tip clearance, and the open impeller is hardly applied.
Furthermore, an impeller with a shroud having an intermediate blade applicable to a single-shaft multistage centrifugal compressor has the following manufacturing problems.

There are several ways to process the impeller. The most widely used method is a welding method, and there are other methods such as casting, diffusion bonding, brazing, or electric discharge machining.
For example, when the welding method is used, welding is not easy in the vicinity of the front edge of the intermediate blade. Even when slot welding in which grooves are formed in the shroud and welded from the outside is used, it is difficult to finish the weld bead in addition to the weld itself in the vicinity of the front edge. This is because the machining tool cannot be physically inserted.

Brazing is joined without welding beads, so there is no problem in finishing, but the strength of brazing itself is weak and stress concentration tends to occur, so there is a risk in applying it to a compressor that rotates at high speed. large. Although electric discharge machining can be performed into an arbitrary shape, there are problems in terms of machining time and machining cost.
When manufactured by casting, it is relatively easy to manufacture an impeller with intermediate blades. Therefore, since a plurality of impellers can be manufactured with one type of mold in a pump that handles incompressible fluid, this manufacturing method can be adopted. However, in the case of a multi-stage compressor, the impeller shape varies depending on the gas compressibility. It is not realistic in terms of cost to make all impellers from castings, differing in stages.

  An object of the present invention is to provide both a surge margin and a choke merge, provide a centrifugal impeller having a wide operating range, and provide a centrifugal impeller suitable for a single-shaft multistage centrifugal compressor.

In order to solve the above-mentioned problem, the centrifugal impeller for flowing out fluid in the centrifugal direction by rotation of a plurality of blades according to the present invention, wherein the plurality of blades have a constant blade thickness from the leading edge to the throat position. A main blade having a blade thickness increased from the throat position to the rear edge toward the downstream, and the main blade are alternately provided, and a leading edge thereof is provided downstream from the throat position, and the main blade and more short intermediate blade length, Ri consists, the main blade and the intermediate blade, between the shroud facing the impeller hub and the impeller hub, are arranged in the circumferential direction, the main blade, The impeller hub and the shroud are integrally joined, and the intermediate blade is integrally joined to either the impeller hub or the shroud.

  Further, in the single-shaft multistage centrifugal compressor in which a plurality of irregularly shaped centrifugal impellers are arranged in series on one rotating shaft of the present invention and the fluid is sequentially boosted, the centrifugal impeller is arranged at least in the most downstream stage. I made it.

  According to the present invention, it is possible to realize a centrifugal impeller with high performance and a wide operating range, and it is possible to achieve high performance, high reliability, and low cost of a single-shaft multistage centrifugal compressor.

It is a figure which shows the cross-sectional structure of the uniaxial multistage centrifugal compressor of an Example. It is a figure which shows the meridian surface shape of the impeller of an Example. It is a three-dimensional perspective view (wire frame) of the impeller of an Example. It is a figure which shows the load distribution of the blade | wing of an impeller. It is a figure which shows an example of the joining structure of the main blade | wing and intermediate | middle blade | wing of an Example. It is a three-dimensional perspective view (wire frame) which shows the blade | wing thickness of the impeller of an Example. It is a figure which shows the blade | wing thickness distribution of a main blade | wing and an intermediate blade. It is a figure which shows the other example of blade | wing thickness distribution of a main blade | wing and an intermediate blade | wing. It is a figure explaining the characteristic of the compressor of the first stage of a single axis multi stage centrifugal compressor. It is a figure explaining the characteristic of the compressor of the last stage of a uniaxial multistage centrifugal compressor. It is a figure explaining the characteristic curve of a compressor.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.
<Configuration of single-shaft multi-stage centrifugal compressor>
First, the configuration of the single-shaft multistage centrifugal compressor will be described. FIG. 1 is a cross-sectional configuration diagram of a single-shaft multistage centrifugal compressor according to an embodiment.

  The single-shaft multistage centrifugal compressor of the embodiment is an example of a five-stage machine in which five impellers 20-1 to 20-5 are provided on one rotary shaft 10. The compressed fluid flows from the suction nozzle 12, passes through the first stage impeller 20-1, the diffuser 1c, and the return channel 1d and flows into the second stage impeller 20-2. Hereinafter, similarly, the compressed fluid passes through the diffusers 2c, 3c, 4c and the return channels 2d, 3d, 4d, and the third stage impeller 20-3, the fourth stage impeller 20-4, and the fifth stage impeller. 20-5 flows into the discharge nozzle 13 through the diffuser 5c. During this time, the compressed fluid is sequentially pressurized by the impellers 20-1 to 20-5. Reference numeral 8 in the figure indicates the flow direction of the compressed fluid.

The impellers 20-1 to 20-5 are driven by a single rotation shaft 10 that rotates in the direction of reference numeral 11 and compresses the compressed fluid.
The rotary shaft 10 is supported on both sides by radial bearings 6 a and 6 b, and is supported by a thrust bearing 7 in the axial direction. The rotating shaft 10 is sealed by shaft end labyrinths 9a and 9b.
The impellers 20-1 to 20-5 are impellers with shrouds as shown in the figure, and between the rotating part and the stationary part, there are provided mouse labyrinths 1b to 5b and interstage labyrinths 1e to 4e, Seals pressure leakage between stages.

  The pressure of the compressed fluid is increased with the rotation of the rotating shaft 10 and the temperature of the compressed fluid is increased at the same time, so that the fluid temperature is higher in the downstream stage. Due to this temperature rise, the rotating shaft 10 is thermally expanded and extends a predetermined length in the axial direction, but the mouse labyrinths 1b to 5b and the interstage labyrinths 1e to 4e are sufficiently absorbed even if the rotating shaft extends in the axial direction. Because it can, the leak does not change.

Since the axial direction of the rotary shaft 10 is fixed by the thrust bearing 7 on the front stage side, the amount of displacement in the axial direction of the rear stage impeller is increased. If an open impeller without a shroud is used for the impeller, a blade tip gap corresponding to the extension of the shaft is generated, resulting in a large deterioration in performance.
For this reason, in a single-shaft multi-stage compressor, there may be cases where an open impeller is used only for the first stage or only for the first and second stages due to strength reasons. There is no case of applying.

<Structure of shroud with shroud>
Hereinafter, the structure of the impeller with shroud applied to the single-shaft multistage centrifugal compressor of this embodiment will be described with reference to FIGS. Although details will be described later, at least the impeller 20-5 of the single-shaft multi-stage centrifugal compressor shown in FIG. 1 has the following structure.
FIG. 2 is a diagram showing a meridional cross-sectional shape of the impeller 20 with the shroud. FIG. 3 is a three-dimensional perspective view (wire frame) of the impeller 20 with the shroud omitted in order to make the blade shape easy to see.

  The impeller 20 shown in FIG. 2 rotates integrally with a rotary shaft 10 (not shown) that rotates in a direction shown by reference numeral 11, and a main blade 20 c on an impeller hub 20 a formed in a flange shape. The intermediate blades 20d are alternately formed, and the impeller shroud 20b is formed so that the main blade 20c and the intermediate blade 20d are sandwiched between the impeller hubs 20a. Reference numeral 20f indicates the blade throat position when the intermediate blade 20d has the same shape as the main blade 20a.

《Expanding chalk merging》
The intermediate blade 20d of the embodiment is formed such that the front edge of the intermediate blade 20d is located downstream of the throat position 20f. Further, the trailing edge of the intermediate blade 20d is formed so as to be located at the trailing edge of the blade 20e, which is the same as the trailing edge of the main blade 20a.

As is clear from the three-dimensional perspective view of the impeller 20 in FIG. 3, the number of blades at the throat position 20f is halved by setting the positions of the front edges of the main blade 20a and the intermediate blade 20d to the above-described positions. Thus, the throat area corresponding to the thickness of the blade can be widened.
Of course, there is also a throat of the intermediate blade (minimum area between the intermediate blade and the main blade) on the downstream side of the leading edge of the intermediate blade, but this throat is formed at a position having a larger radius than the throat of the main blade. Therefore, it becomes larger than the throat area at the throat position 20f, and there is no influence on the operating range.

  In this embodiment, the choke margin can be increased by forming the above-described impeller 20 with the shroud having the main blade 20a and the intermediate blade 20d by the manufacturing method described in detail later.

《Expansion of surge merge》
Next, a method for enlarging surge merge will be described.
In the impeller 20 of the embodiment, the load of the main blade 20a and the intermediate blade 20d of the impeller becomes a minimum value at the leading edge of the blade, and increases from the minimum value along the camber line to a maximum value. The main blades so that the maximum value decreases from the maximum value along the camber line toward the trailing edge, and the minimum value of the load of the blades is large enough to suppress the backflow of the compressed fluid at the leading edge. The blade angles of 20a and 20d are distributed.

FIG. 4 is a diagram showing the load distribution of the blades of the embodiment. The horizontal axis in FIG. 4 represents a dimensionless number obtained by dividing the camber line length by the length (full length) of the shroud curve.
Specifically, the blade load is an index indicating the flow velocity difference or pressure difference of the compressed fluid flowing on both sides of the blade. The greater the blade load, the higher the deceleration rate of the compressed fluid flowing inside the impeller. For this reason, in order to delay the occurrence of the surge, it is preferable to moderate the deceleration of the relative speed at the front edge of the blade and suppress the backflow.

  Therefore, in this embodiment, as shown in FIG. 4, the bending point P1 of the blade load distribution is provided in the vicinity of the blade throat position. In other words, the blade has a blade angle with a distribution in which the blade load is suppressed to be smaller on the front edge side than the throat position and the blade load is rapidly increased on the rear edge side from the throat position. With such a configuration, it is possible to suppress the deceleration of the compressed fluid at the inlet of the main blade in the impeller, which is involved in the occurrence of surge, and to obtain an ideal relative velocity distribution that decelerates the sudden compressed fluid downstream from the throat position. it can.

  As shown in FIGS. 2 and 3, the leading edge of the intermediate blade 20d is positioned downstream of the throat position 20f in order to increase the choke margin, and the load distribution of the intermediate blade 20d is the bending of FIG. The blade has a load distribution downstream from the point P1.

The main blade 20c and the intermediate blade 20d of this embodiment determine the shape of the shroud curve by the inverse solution method so as to satisfy the blade load distribution along the shroud curve shown in FIG.
This inverse solution is, for example, a method of obtaining a desired blade load distribution first and determining the shape of the blade based on the distribution. The forward solution method of determining the blade shape first determines the desired blade load distribution. Easy to realize distribution.

  The load distribution in the camber line direction is a distribution indicating at which position of the blade the blade is to work, that is, the head is raised, and can be expressed by the following equation (Equation 1). (Equation 1) is that when the radius is r, the circumferential average absolute velocity of the compressed fluid is Cθ, and the camber wire length is x, the blade load λ is the circumferential average absolute velocity with respect to the change amount of the camber wire length x. This is the amount of change in the product “r · Cθ” of Cθ and the radius r.

  Therefore, when the blade load is determined, the relationship between the camber line length x and the radius r corresponding to the circumferential average absolute velocity Cθ of the compressed fluid can be calculated. For example, the blade angle β can be set based on the following (Formula 2). That is, when the blade load is determined, the blade angle β can be set by an inverse solution, and the shape of the shroud curve can be determined by setting the blade angle β continuously along the shroud curve.

(Equation 2) is a projection of the distance between an arbitrary point on the shroud curve and the axis center 5a to the meridian plane from the radius r, the angle between the radius r and the horizontal direction θ, and the leading edge of the shroud curve to the point. The blade angle β is shown when the length of m is m.

The shape of the blade shroud curve is determined by continuously setting the blade angle β on the shroud curve side from the leading edge to the trailing edge. Similarly, the shape of the hub curve is determined by continuously setting the blade angle β of the hub curve from the leading edge to the trailing edge.
Then, the shroud curve and the hub curve are smoothly connected, for example, linearly, to form a blade.

  The main blade 20c having the blade shape determined as described above has a small load from the leading edge of the blade to the throat portion and a light load at the inlet portion. Therefore, when the flow rate of the impeller decreases, this portion This prevents the stall and separation of the flow, and increases the surge margin.

  In this embodiment, since the intermediate blade is employed in the impeller 20 with the shroud, the throat area can be increased, thereby increasing the choke margin and expanding the entire operation range. The light load at the inlet of the main blade is the same as that disclosed in Patent Document 1, but the difference from Patent Document 1 is that not only the surge margin but also the choke margin is increased by adopting a configuration having an intermediate blade. It is in.

  In addition, the configuration and shape features of the main blade 20c and the intermediate blade 20d described above can also be applied to a mixed flow impeller having a flow of compressed fluid in the centrifugal direction of the impeller rotation.

  By the way, when an intermediate blade is applied to an impeller with a shroud, there is a manufacturing problem such as difficulty in finishing the weld bead in the vicinity of the front edge. Next, a solution will be described.

<Production method of impeller>
A manufacturing method of the shroud-equipped impeller 20 will be described with reference to an example of a joining structure of the main blade 20c and the intermediate blade 20d in FIG.
First, the main blade 20c and the intermediate blade 20d are cut into the shape determined as described above by cutting the impeller hub 20a.
Next, a separately manufactured impeller shroud 20b is joined by welding. At this time, only the main blade 20c is joined, and the intermediate blade 20d is not welded.

  Specifically, the height of the intermediate blade 20d is set lower by δ than the main blade 20c. This is because, when the main blade 20c is welded and cooled, the blade contracts, so that a slight gap is provided before welding in consideration of the contraction allowance. After the welding is completed, the residual stress due to the shrinkage of the main blade 20c after welding is relieved by providing a gap δ between the main blade 20c and the intermediate blade 20d in advance so that the gap becomes substantially zero.

This produces the following merit.
First, since half of the blades (intermediate blade 20d) are not welded, the welding work and the finishing work time of the bead after welding are greatly shortened.
Next, since the impeller entrance portion is only the main blade 20c, the space is wide and the workability of the welding of the front edge of the main blade 20c and the weld bead finish is improved as compared with the case where all the blades are the main blade 20c.

  In this example, the welding method may be welding from the doorway or slot welding from the outside of the shroud 20b. Further, in this example, the main blade 20c and the intermediate blade 20d are cut out on the impeller hub 20a, and the shroud 20b is welded. Conversely, the main blade 20c and the intermediate blade 20d are cut out on the shroud 20b side, Alternatively, the main blade 20c and the intermediate blade 20d may be joined to the side first by welding, and then only the main blade 20c may be joined to the impeller hub hub 20a.

  Alternatively, the main blade may be cut off on the impeller hub side, the intermediate blade may be cut off on the shroud side, and the main blade and the shroud may be welded. In this case, the intermediate blade may be cut off on the impeller hub side, and the main blade may be cut off on the shroud side. In either case, the impeller hub and the shroud are integrated by welding the main blade.

  As mentioned above, since the intermediate blade is not joined, welding near the leading edge of the intermediate blade and the finishing of the weld bead, which was a problem when adopting the intermediate blade, are unnecessary, and the impeller can be easily manufactured. Strength reliability can be ensured and production cost can be reduced.

<Thickness of the blade>
Next, the thickness of the blades of the impeller 20 of the embodiment will be described with reference to FIG.
FIG. 6 is a three-dimensional perspective view (wire frame) showing the blade thickness of the impeller of the embodiment. Although the shroud 20b is omitted, the main blade 20c is joined to the shroud 20b as described above.

In this embodiment, the main blade 20c and the intermediate blade 20d are thickened toward the downstream (rear edge).
By doing so, it is possible to relieve the stress generated in the blades and to ensure sufficient strength of the blades even when the force applied to each blade is large by joining only the main blade 20c.

  In particular, in a low specific speed stage with a small handling flow rate, by increasing the thickness of the blade, the ratio of the height and width of the flow channel, that is, the aspect ratio, can be made close to 1 in the flow channel that flows the same flow rate. . Since the friction loss becomes dominant at the low specific speed stage, this makes it possible to reduce the friction loss and improve the efficiency. That is, it is possible to provide an impeller that satisfies the relaxation of the blade stress and the reduction of the friction loss at the same time, is highly reliable, has high efficiency, and has a wider operation range than the conventional one.

Next, the distribution of the blade thickness of the main blade 20c and the intermediate blade 20d is shown in FIG. 7 and FIG.
In the example shown in FIG. 7, the main blade 20c has a constant blade thickness (small blade thickness) from the leading edge to the throat position 20f, and the blade thickness linearly toward the trailing edge after the throat position 20f. It is increasing.
As described above, the intermediate blade 20d is formed so that the front edge thereof is located downstream of the throat position 20f, and the blade is linearly directed toward the rear edge so that the rear edge has the same thickness as the main blade 20c. It is thick.
By doing so, it is possible to secure a wider throat area and to increase the choke flow rate.

In the example shown in FIG. 8, the main blade 20c has a constant blade thickness (small blade thickness) from the leading edge to the throat position 20f, as in FIG. In comparison, the thickness of the main blade is increased on the upstream side.
As in FIG. 7, the intermediate blade 20d is formed so that the front edge is located downstream of the throat position 20f and the front edge is located, and the rear blade has the same thickness as the main blade 20c. The blade thickness is increased linearly.

  This is aimed at reducing the stress of the blades. If the rotational speed is high and the strength is severe, the blade thickness distribution shown in FIG. High-speed and compact design can be achieved, which can contribute to improvement in compressor performance and cost reduction.

  Further, when only the main blade 20c is joined, the stress load on the blade supporting the shroud 20b is increased. By increasing the blade thickness of the main blade 20c toward the downstream as compared with the conventional case, the blade stress is increased. As a result, for the impeller 20 with a small handling flow rate, that is, a low specific speed impeller, the wet blotting area of the impeller passage can be reduced and the friction loss can be reduced, so that the fluid performance is also improved.

《Multi-stage configuration》
Next, the number of application stages effective when the above-described impeller 20 is applied to the single-shaft multistage centrifugal compressor of the present embodiment will be described. FIG. 1 shows a single-shaft multi-stage centrifugal compressor in which the above-described impeller 20 is applied to the fourth and fifth impellers 20-4 and 20-5. Although details will be described later, in the single-shaft multistage centrifugal compressor of the embodiment, the above-described impeller 20 is applied to at least the impeller 20-5 of the last stage. The reason will be described below.

  A single-shaft multistage centrifugal compressor is required to have a specified operating range, but each stage designed with the specified point flow rate will not match the flow matching of each stage when operating at a point where the flow rate is far from the specified point. Arise. This matching shift will be described with reference to the characteristic diagram of the first stage compressor in FIG. 9A and the characteristic diagram of the last stage compressor in FIG. 9B.

  As described above, the operation limit on the small flow rate side of the compressor is determined by the occurrence of surging. On the other hand, the operation limit on the large flow rate side is determined by the generation of choke.

When the surging point is the surge point Ps, the flow rate at that time is Qs, the choke point is the choke point Pc, and the flow rate at that time is Qc, the flow range from Qs to Qc is the operating range. The compressor is in operation.
In the middle of this operating range, there is a specification point P _Des which is also a design point. The boundary of the flow rate _{Q Des} of this specification point _{P Des,} is called surge margin the flow rate range from the surge flow rate Qs to the specification point flow rate _{Q Des,} the flow rate range from choke flow rate Qc to the specification point flow rate _{Q Des} choke margin .
When there is a specification point P _Des that is also a design point in the middle of this operating range, and the first stage compressor is operated at the specification point flow rate Q _Des (volume flow rate ratio = 1.0), the final stage compressor is also specified. It is operated at a point flow rate Q _Des .

On the other hand, in FIG. 9A, when the flow rate of the first stage compressor is increased to the flow rate at the large flow side operating point A in order to increase the flow rate, the head H of the first stage compressor becomes HA, and the head H at the specification point P _{Des Lower} than _Des . That is, since the compressed fluid is not compressed as compared with the operation at the specification point P _Des , the downstream impeller operates at a larger volume flow rate ratio.
As a result, as shown in FIG. 9B, the final stage compressor is operated at the operating point A ′ on the large flow rate side in the volume flow rate ratio than the first stage compressor. That is, the operation is performed at a position where the deviation from the specification point flow rate Q _Des is large.

  On the contrary, in FIG. 9A, when the first stage compressor is operated at the operation point B on the smaller flow rate side than the specification point PDes, the head H becomes higher (HB) than the head HDes at the specification point PDes, so that the gas Are more compressed, and the compressors in the subsequent stages are operated at an operating point with a smaller volume flow rate ratio. For example, in the final stage, the operation is performed at the operation point B 'on the small flow rate side close to the surge point Ps.

  As described above, the operating range of the single-shaft multi-stage compressor greatly depends on the operating range of the downstream compressor, particularly the final compressor, of the constituting compressors. Therefore, in order to obtain a single-shaft multistage compressor with a wide operating range, it is necessary to increase the operating range toward the downstream side of the compressor.

  The single-shaft multi-stage compressor in FIG. 1 uses the impeller having a wide operating range of the present invention on the rear side (in this case, the fourth and fifth stages). As a result, the operating range of the compressor as a whole can be expanded.

  The present invention is not limited to the above-described embodiments, and includes various modifications. The above-described embodiments have been described in detail for easy understanding in the present invention, and are not necessarily limited to those having all the configurations described. Further, a part of the configuration of an embodiment can be replaced with the configuration of another embodiment, and the configuration of another embodiment can be added to the configuration of an embodiment.

8 Direction of flow 20 (20-1, 20-2, 20-3, 20-4, 20-5) Impeller 20a Impeller hub 20b Impeller shroud 20c Main blade 20d Intermediate blade 20e Blade trailing edge 20f Throat position

Claims (4)

A centrifugal impeller that discharges fluid in a centrifugal direction by rotation of a plurality of blades, wherein the plurality of blades are
A main blade having a constant blade thickness from the leading edge to the throat position, and increasing the blade thickness from the throat position to the trailing edge toward the trailing edge ;
An intermediate blade provided alternately with the main blade, a leading edge thereof provided downstream from the throat position, and a length shorter than the main blade,
Ri consists of,
The main blade and the intermediate blade are disposed in a circumferential direction between an impeller hub and a shroud facing the impeller hub,
The main blade is integrally joined to the impeller hub and the shroud,
The centrifugal impeller , wherein the intermediate blade is integrally joined to one of the impeller hub and the shroud . The centrifugal impeller according to claim 1 ,
The centrifugal impeller characterized in that the height of the intermediate blade is a predetermined amount lower than the height of the main blade. In a single-shaft multistage centrifugal compressor in which a plurality of centrifugal impellers are arranged in series on one rotating shaft and the fluid is boosted in sequence,
The centrifugal impeller according to claim 1 or 2 is arranged at least at the most downstream stage, and is a single-shaft multistage centrifugal compressor. It is a manufacturing method of the centrifugal impeller of Claim 1,
Cutting the impeller hub and cutting out the main blade and the intermediate blade;
Joining only the main blade to the impeller shroud by welding;
A method of manufacturing a centrifugal impeller characterized by comprising: