JP7543153B2 - Centrifugal Compressor - Google Patents

Description

The present invention relates to a centrifugal compressor, and in particular to a centrifugal compressor suitable for one that has a return vane in the return flow passage that constitutes the stationary flow passage.

Centrifugal fluid machines with rotating centrifugal impellers have traditionally been used in a variety of plants, air conditioning equipment, liquid pressure pumps, turbochargers, etc.

In response to the growing demand for reducing environmental impact in recent years, these fluid machines are required to be more efficient and have a wider operating range than ever before, while at the same time, from the perspective of reducing costs and saving space in the plant, there is a demand for the centrifugal compressors themselves to be more compact.

In order to achieve high efficiency, a wide operating range, and compact size in fluid machinery, it is important to reduce the outer diameter of the stationary flow passage. In a centrifugal compressor, the stationary flow passage is a flow passage provided downstream of the discharge port of the rotating impeller, and is composed of a diffuser flow passage and a return flow passage. Of these, the return flow passage is a flow passage that removes the swirling component of the flow that has passed through the diffuser flow passage, and guides the flow without pre-swirl to the next stage impeller.

However, when the outer diameter of the stationary flow passage is reduced, the length of the return flow passage that constitutes the stationary flow passage is also shortened, so it is necessary to redirect the flow over a shorter distance to eliminate pre-swirl of the flow. In order to redirect the flow efficiently in the return flow passage that constitutes the stationary flow passage, blades called return vanes are usually provided at equal intervals in the circumferential direction.

The return flow passage has blades called return vanes that are arranged at equal intervals in the circumferential direction, and the blades described in Patent Documents 1 to 3 have been proposed.

The above-mentioned Patent Document 1 describes that in order to obtain a centrifugal turbomachine with return vanes of a shape that can suppress a decrease in efficiency when miniaturized, the blade surfaces of the return vanes arranged in a multiple circular cascade around a center line with the axial direction of the rotating shaft as the height direction in a return flow passage where fluid flows in the return direction toward the rotating shaft are curved surfaces that redirect the flow of fluid in the return flow passage from the circumferential direction around the center line to the radial direction toward the rotating shaft, and that the camber line of the blade cross section of the outer blade located at the most upstream side of the return vanes, cut on a plane perpendicular to the axial direction of the rotating shaft, exhibits a curved shape that varies in the height direction.

In addition, the above-mentioned Patent Document 2 describes a centrifugal pump that suppresses the remaining swirling flow when the fluid is introduced to the next impeller while removing the swirling component of the fluid flow with the return vane, and includes a rotating shaft that rotates around its axis, multiple impellers arranged in the axial direction on the rotating shaft and compressing the fluid by centrifugal force, a flow path that reverses the fluid compressed radially outward by the upstream impeller to the radially inward direction and flows into the downstream impeller, and multiple return vanes that are arranged at intervals in the circumferential direction in the flow path after the fluid is reversed and curve to turn the fluid radially inward, and the return vanes have a first communication portion that communicates the pressure surface and the negative pressure surface so that the return vanes are inclined downstream as they move from the pressure surface of the return vane to the negative pressure surface.

Furthermore, the above-mentioned Patent Document 3 describes that in order to obtain a multi-stage centrifugal compressor that can reduce the occurrence of flow separation from the guide vane surface without increasing costs, the multi-stage centrifugal compressor has impellers arranged in multiple stages, diffusers arranged downstream of each impeller, and a return flow path arranged downstream of the diffuser to guide the flow to the next stage impeller, and has a first circular blade row consisting of a plurality of first guide vanes arranged on the outer circumferential side of the return flow path to redirect the flow flowing in from the diffuser by a first angle, and a second circular blade row consisting of a plurality of second guide vanes arranged on the inner circumferential side of the first circular blade row to further redirect the flow flowing in from the first circular blade row by a second angle, and the first and second circular blade rows are arranged in a staggered pattern.

Patent No. 06339794 Patent No. 06097487 JP 2001-200797 A

However, if the radial length of the return vane is reduced to further reduce the size of the centrifugal compressor, the amount of flow deflection required between the inlet and outlet of the return vane becomes larger relative to the length of the vane.

In the return vanes of the centrifugal compressors and centrifugal pumps described in Patent Documents 1 to 3 mentioned above, as centrifugal compressors and centrifugal pumps become smaller, it becomes necessary to increase the curvature of the camber line (a line connecting points equidistant from the upper and lower surfaces of the blade) of the cross section (airfoil shape) of the blade cut by a plane perpendicular to the axial direction of the main shaft (rotating shaft), which increases the likelihood of flow separation.

In order to avoid the above-mentioned flow separation, the above-mentioned Patent Documents 1 to 3 provide a double blade row, but when considering further miniaturization of centrifugal compressors, the load acting on each blade will be excessive if only the shape of the blades alone is considered. Therefore, even if the blades are simply provided in double or triple layers, there is a risk that the flow will separate from the blade surface, and efficiency may not be improved.

The present invention has been made in consideration of the above points, and its purpose is to provide a centrifugal compressor that can reduce the outer diameter of the stationary flow path while maintaining and improving efficiency.

In order to achieve the above object, a centrifugal compressor of the present invention comprises a rotating shaft, a plurality of centrifugal impellers attached to the rotating shaft, a diffuser through which a fluid flowing out of the centrifugal impeller flows in a centrifugal direction away from the rotating shaft, a return flow passage provided downstream of the diffuser and through which the fluid flowing from the diffuser into the centrifugal impeller at a subsequent stage flows in a return direction toward the rotating shaft, a plurality of return vanes disposed in a circular cascade centered on a center line of the rotating shaft and installed in the return flow passage, and a turning portion at which the flow of the fluid that has flowed through the diffuser turns from the centrifugal direction to an axial direction, and further turns from the axial direction to the return direction,
a centrifugal compressor in which the return vanes, in which the circular blade rows are provided in multiple rows, are arranged in two rows from the upstream side to the downstream side of the flow of the fluid in the return flow passage, wherein an inlet blade angle (β) of a rear blade provided on the downstream side of the return vanes is inclined in the circumferential direction relative to an inlet blade angle (α) of a front blade provided on the upstream side of the return vanes ,
A plurality of the return vanes having a blade shape are provided in the return flow passage in a circumferential direction as a front blade row and a rear blade row on the upstream and downstream sides of the return flow passage, respectively;
the trailing vane is provided offset toward the pressure surface side of the leading vane in order to guide the flow on the pressure surface side of the leading vane to the suction surface of the trailing vane,
The angle (θ) between the leading edge of the lead vane and the trailing edge of the trail vane is smaller than the angle (γ) between the leading edge of the lead vane and the leading edge of the lead vane adjacent in the circumferential direction, and
The maximum camber position of the camber line of the fore-placed wing (the position on the chord line at which the distance (camber) of a perpendicular line extending vertically from any position on the straight line (chord line) connecting the leading edge and trailing edge of the wing to the camber line is maximum) is located in the rear half of the chord line .

The present invention provides a centrifugal compressor that can reduce the outer diameter of the stationary flow passage while maintaining and improving efficiency.

FIG. 1 is a meridional cross-sectional view showing an upper half of an overall configuration of a typical centrifugal compressor. FIG. 2 is a partially enlarged cross-sectional view of the centrifugal compressor shown in FIG. 1 . 3 is a diagram showing half of the periphery of the return vane shown in FIGS. 1 and 2 as viewed from the downstream side in the axial direction of the rotary shaft. FIG. FIG. 2 is a diagram showing half of a return vane and its periphery in the centrifugal compressor according to the first embodiment of the present invention, as viewed from the downstream side in the axial direction of the rotary shaft. FIG. 2 is a schematic diagram showing the positional relationship between the front and rear blades of a return vane in the centrifugal compressor according to the first embodiment of the present invention. FIG. 1 is a diagram showing a comparison of the angular distribution of the flow around the return vane in the centrifugal compressor according to the first embodiment of the present invention. FIG. 2 is a diagram showing the relationship between the radial length of the trailing edge of the front blade of the return vane and the radial length of the leading edge of the rear blade of the centrifugal compressor according to the first embodiment of the present invention. FIG. 2 is a diagram showing the relationship between the dimensionless radial position (horizontal axis) and the blade angle distribution (vertical axis) of the front blade of the return vane in the centrifugal compressor according to the first embodiment of the present invention. FIG. 11 is a diagram showing the shape features of the front blades of the return vane in the centrifugal compressor according to the second embodiment of the present invention.

The centrifugal compressor of the present invention will be described below based on the illustrated embodiment. Note that the same reference numerals are used for the same components in each figure.

Before describing the first embodiment of the centrifugal compressor of the present invention, a general centrifugal compressor will be described with reference to Figures 1 to 3.

As shown in the figure, the centrifugal compressor 100 is roughly composed of a centrifugal impeller 1 that imparts rotational energy to the fluid, a rotating shaft 4 to which the centrifugal impeller 1 is attached, and a diffuser 5 that is located radially outside the centrifugal impeller 1 and converts the dynamic pressure of the fluid flowing out from the centrifugal impeller 1 into static pressure. In addition, a return flow path 6 is provided downstream of the diffuser 5 to guide the fluid to the subsequent centrifugal impeller 1.

Although not specifically shown, the centrifugal impeller 1 typically has a disk (hub) that is fastened to the rotating shaft 4, a side plate (shroud) that is arranged opposite the hub, and multiple blades that are located between the hub and the shroud and spaced apart in the circumferential direction (perpendicular to the plane of the paper in FIG. 2).

The diffuser 5 can be either a vaned diffuser having multiple blades arranged at approximately equal pitch in the circumferential direction, or a vaneless diffuser that does not have blades (not shown in Figure 2).

The return flow passage 6 is composed of turning sections 7a and 7b, where the flow of the fluid that has passed through the diffuser 5 is turned from the centrifugal direction to the axial direction, and further turned from the axial direction to the return direction, and a return vane 8 (see Figure 2). The return vane 8 turns the fluid that has passed through the diffuser 5 from the radially outward direction to the radially inward direction, and further removes the swirling component of the fluid, rectifying the fluid while allowing it to flow into the centrifugal impeller 1 of the next stage.

As shown in FIG. 2, the turning sections 7a and 7b, which turn from the axial direction to the return direction, are formed as U-shaped curved flow passages surrounded by surrounding structures in the meridian plane, with the turning section inlet 9 defined by an approximately cylindrical surface corresponding to the outlet of the diffuser 5, and the turning section outlet 10 defined as the section from the turning section inlet 9 to the turning section outlet 10, which is defined by an approximately cylindrical surface corresponding to the end of the meridian plane curved flow passage located immediately upstream of the return vane leading edge 12.

The return vane 8 is composed of multiple blades arranged at approximately equal pitch in the circumferential direction around the rotating shaft 4. In addition, although not shown, radial bearings that rotatably support the rotating shaft 4 are arranged on both ends of the rotating shaft 4 in the centrifugal compressor 100.

In addition, a multi-stage compression centrifugal impeller (six centrifugal impellers in FIG. 1) 1 is attached to the rotating shaft 4, and a diffuser 5 and a return passage 6 are provided downstream of each centrifugal impeller 1, as shown in FIG. 2.

The centrifugal impeller 1, diffuser 5, and return passage 6 are housed in a casing 19, which is supported by flanges 20a and 20b. A suction passage 15 is provided on the suction side of the casing 19, and a discharge passage 16 is provided on the discharge side of the casing 19.

In the centrifugal compressor 100 configured in this manner, as shown in FIG. 1, the fluid sucked in from the suction passage 15 is pressurized as it passes through the centrifugal impeller 1, the diffuser 5, and the return passage 6 of each stage, and is finally discharged from the discharge passage 16 at a predetermined pressure.

However, in the centrifugal compressor 100 configured in this manner, as described above, if the radial length of the return vane 8 is reduced to further reduce the size, the amount of flow deflection required between the inlet and outlet of the return vane 8 becomes large relative to the length of the centrifugal impeller 1, which may cause flow separation and prevent efficiency from being improved.

The centrifugal compressor 100 of this embodiment solves this problem, and its details are explained below using Figures 4 and 5.

Figure 4 shows half of the periphery of the return vane 8 in the centrifugal compressor 100 according to the first embodiment of the present invention, as viewed from the downstream side in the axial direction of the rotating shaft 4, and Figure 5 is a schematic diagram showing the positional relationship between the front blade 8A and the rear blade 8B of the return vane 8 in the centrifugal compressor 100 according to the first embodiment of the present invention.

The centrifugal compressor 100 of this embodiment shown in Figures 4 and 5 is a centrifugal compressor in which return vanes 8, each having a multiple circular blade row, are arranged in two rows from the upstream side to the downstream side of the fluid flow in the return flow passage 6. In this embodiment, the inlet blade angle (β) of the rear blade 8B located downstream of the return vane 8 is tilted in the circumferential direction relative to the inlet blade angle (α) of the front blade 8A located upstream of the return vane 8. Specifically, the inlet blade angle (β) of the rear blade 8B of the return vane 8 and the inlet blade angle (α) of the front blade 8A have a relationship of β<α.

As shown in Figure 6, the distribution of flow angles around the return vane 8 obtained by numerical analysis shows that the flow angle around the leading blade 8A of the return vane 8 hardly changes from the leading edge 8A3 of the leading blade 8A to near the leading edge 8B2 of the trailing blade 8B on the pressure surface 8A1 side of the leading blade 8A. This means that the leading blade 8A of the return vane 8 does not form a throat with the blade, so the flow is only partially redirected.

This shows that the blade angle of the leading edge 8B2 of the trailing blade 8B of the return vane 8 must be at least as large as the blade angle of the leading blade 8A of the return vane 8.

In this embodiment, a plurality of wing-shaped return vanes 8 are installed in the circumferential direction in the return flow passage 6 as a front blade row and a rear blade row on the upstream and downstream sides of the return flow passage 6, respectively. In order to guide the flow on the pressure surface 8A1 side of the front blade 8A to the negative pressure surface 8B1 of the rear blade 8B of the return vane 8, the rear blade 8B of the return vane 8 is offset toward the pressure surface 8A1 side of the front blade 8A, and as shown in FIG. 7, the leading edge 8B2 of the rear blade 8B of the return vane 8 is shorter in radial length from the center of the rotating shaft 4 than the trailing edge 8A2 of the front blade 8A (there is a relationship of L1>L2 shown in FIG. 7).

In addition, the angle (θ) between the leading edge 8A3 of the front blade 8A of the return vane 8 and the trailing edge 8B3 of the trailing blade 8B is smaller than the angle (γ) between the leading edge 8A3 of the front blade 8A of the return vane 8 and the leading edge 8A3 of the circumferentially adjacent front blade 8A.

Furthermore, in this embodiment, the camber line 8A4 (a line connecting points equidistant from the upper and lower surfaces of the blade) of the forward blade 8A of the return vane 8 has a constant blade angle for more than 50% of the first half from the leading edge 8A3 to the trailing edge 8A2 of the forward blade 8A.

This is because the angle of the camber line 8A4 of the front blade 8A of the return vane 8 shown in Figure 5 does not change over more than half (50% or more) of the leading edge 8A3 side from the leading edge 8A3 to the trailing edge 8A2 of the front blade 8A of the return vane 8. As can be seen from the relationship between the dimensionless radial position (horizontal axis) and the blade angle distribution (vertical axis) of the front blade 8A of the return vane 8 shown in Figure 8, the angle of the camber line 8A4 of the front blade 8A of the return vane 8 does not change over more than half (50% or more) of the leading edge 8A3 side from the leading edge 8A3 to the trailing edge 8A2 of the front blade 8A of the return vane 8.

The effects of the centrifugal compressor 100 of this embodiment configured in this way are as follows:

That is, the inlet blade angle (α) of the leading blade 8A, which is located upstream of the return vane 8, is tilted in the circumferential direction. Specifically, by making the inlet blade angle (β) of the trailing blade 8B of the return vane 8 and the inlet blade angle (α) of the leading blade 8A have a relationship of β < α, the fluid flows and flows in from the suction surface 8B1 side of the trailing blade 8B.

This increases the pressure in the flow passage formed between the leading blade 8A and trailing blade 8B of the return vane 8, increasing the flow velocity of the flow passing through this passage. As the flow velocity increases, the momentum of the flow passing through the negative pressure surface 8B1 of the trailing blade 8B increases, making it possible to suppress flow separation that occurs on the negative pressure surface 8B1 of the trailing blade 8B. By suppressing flow separation, it is possible to both suppress the decrease in efficiency associated with separation and redirect the flow.

In addition, by colliding the flow with the negative pressure surface 8B1 of the trailing blade 8B of the return vane 8, the pressure on the pressure surface 8B4 of the trailing blade 8B is relatively reduced, so that the pressure difference between the pressure surface 8B4 of the trailing blade 8B and the negative pressure surface 8B1 of the adjacent trailing blade 8B is reduced.

This suppresses the secondary flow that occurs between the pressure surface 8B4 of the trailing blade 8B of the return vane 8 and the suction surface 8B1 of the adjacent trailing blade 8B. This suppression of the secondary flow makes it possible to reduce loss of the flow field due to the secondary flow.

Furthermore, the camber line 8A4 of the front blade 8A of the return vane 8 is set to a constant blade angle for at least 50% of the first half from the leading edge 8A3 to the trailing edge 8A2 of the front blade 8A, thereby ensuring a longer blade chord length.

This reduces the blade loading of the leading blade 8A of the return vane 8, suppressing separation on the blade negative pressure surface, and by lengthening the distance between the leading edge 8A3 of the leading blade 8A and the leading edge 8B2 of the trailing blade 8B, the flow direction pressure gradient between adjacent blades becomes gentler, making it possible to suppress separation of the boundary layer that develops on the sidewall in the return flow passage 6.

Therefore, according to the centrifugal compressor 100 of this embodiment, it is possible to maintain and improve efficiency while reducing the outer diameter of the stationary flow path, which is expected to reduce costs and improve operational efficiency. In addition, the reduction in the outer diameter also makes it possible to reduce the floor space occupied by the centrifugal compressor 100 within the facility.

The second embodiment of the centrifugal compressor of the present invention will be described below with reference to Figures 4, 5, and 9.

The centrifugal compressor 100 of this embodiment is a centrifugal compressor in which the return vanes 8, in which the circular blade rows shown in Figures 4 and 5 are provided in multiple rows, are arranged in two rows from the upstream side to the downstream side of the fluid flow in the return flow passage 6, as in the first embodiment. In this embodiment, the inlet blade angle β of the rear blade 8B provided on the downstream side of the return vane 8 is more inclined in the circumferential direction than the inlet blade angle α of the front blade 8A provided on the upstream side of the return vane 8. Specifically, the inlet blade angle β of the rear blade 8B of the return vane 8 and the inlet blade angle α of the front blade 8A have a relationship of β<α.

Next, in this embodiment, a plurality of airfoil-shaped return vanes 8 are installed in the circumferential direction in the return flow passage 6 as a leading blade row and a trailing blade row on the upstream and downstream sides of the return flow passage 6, and the trailing blade 8B of the return vane 8 is offset toward the pressure surface 8A1 of the leading blade 8A in order to guide the flow on the pressure surface 8A1 side of the leading blade 8A to the negative pressure surface 8B1 of the trailing blade 8B of the return vane 8.

In addition, the angle (θ) between the leading edge 8A3 of the front blade 8A of the return vane 8 and the trailing edge 8B3 of the trailing blade 8B is smaller than the angle (γ) between the leading edge 8A3 of the front blade 8A of the return vane 8 and the leading edge 8A3 of the circumferentially adjacent front blade 8A.

Furthermore, in this embodiment, the maximum camber position of the front blade 8A is set in the latter half of the blade chord. The shape characteristics of the front blade 8A of the return vane 8 in the centrifugal compressor 100 of this embodiment will be explained using Figure 9.

Figure 9 is a diagram showing the shape characteristics of the front blade 8A of the return vane 8 in the centrifugal compressor 100 according to the second embodiment of the present invention.

The dashed line 8A6 shown in the figure indicates the chord line, which is a straight line connecting the leading edge 8A3 and the trailing edge 8A2 of the fore-end wing 8A, and the dotted line 8A4 shown in the figure indicates the camber line of the fore-end wing 8A. The arrow 8A7 shown in the figure indicates the camber of the fore-end wing 8A, which is the distance from an arbitrary position on the chord line 8A6 to the camber line 8A4. The arrow 8A8 shown in the figure indicates the maximum camber at which the camber of the fore-end wing 8A is at its maximum.

On the chord line 8A6 in FIG. 9, the distance from the leading edge 8A3 of the fore-placement blade 8A to the maximum camber 8A8 is called the maximum camber position. The maximum camber position is expressed as a percentage (dimensionless chord position) of the length of the chord line 8A6 (chord length L). Here, the leading edge 8A3 of the fore-placement blade 8A corresponds to a position where the dimensionless chord position is 0%, and the trailing edge 8A2 corresponds to a position where the dimensionless chord position is 100%.

In this embodiment, the maximum camber position of the forward wing 8A is set closer to the trailing edge 8A2 than the center of the chord (the position where the dimensionless chord position is 50%), i.e., in the latter half of the chord.

The effects of the centrifugal compressor 100 of this embodiment configured in this manner are similar to those described in the first embodiment, but by setting the maximum camber position of the front blade 8A to the rear half of the blade chord, the following additional effects are obtained:

That is, as shown in FIG. 9, the camber line 8A4 of the leading blade 8A is bent sharply near the trailing edge 8A2, so that the flow direction along the pressure surface 8A1 of the leading blade 8A is toward the suction surface 8B1 of the trailing blade 8B shown in FIG. 5.

This flow presses the flow along the suction surface 8B1 of the trailing blade 8B toward the blade surface, suppressing flow separation that occurs on the suction surface 8B1 of the trailing blade 8B. By suppressing flow separation that occurs on the suction surface 8B1 of the trailing blade 8B, it is possible to both suppress the decrease in efficiency associated with separation and redirect the flow.

If the camber line 8A4 of the fore-end vane 8A were to bend sharply, the flow would be more likely to separate on the negative pressure surface 8A5 of the fore-end vane 8A in that vicinity. However, in this embodiment, the sharp bend in the camber line 8A4 of the fore-end vane 8A is limited to the vicinity of the trailing edge 8A2, so the separation area on the negative pressure surface 8A5 is limited to the area near the trailing edge 8A2.

This makes it possible to effectively suppress flow separation on the negative pressure surface 8B1 of the trailing blade 8B while minimizing the increase in pressure loss at the leading blade 8A.

In addition, in this embodiment, as in the first embodiment, it is more preferable that the leading edge 8B2 of the trailing blade 8B of the return vane 8 be shorter in radial length from the center of the rotating shaft 4 than the trailing edge 8A2 of the leading blade 8A, as shown in FIG. 7 (the relationship L1>L2 shown in FIG. 7).

This is for the following reason. In other words, in order to suppress separation of the flow occurring on the suction surface 8B1 of the trailing blade 8B, it is most effective to narrow the flow passage width formed between the rear half of the pressure surface 8A1 of the leading blade 8A and the front half of the suction surface 8B1 of the trailing blade 8B as much as possible, and to direct the flow from the pressure surface 8A1 of the leading blade 8A to the vicinity of the front half of the blade where the flow speed on the blade surface is most decelerated on the suction surface 8B1 and separation is most likely to occur.

On the other hand, if the width of the flow path between the rear part of the pressure surface 8A1 of the front blade 8A and the front part of the negative pressure surface 8B1 of the rear blade 8B is narrowed too much, a small diameter machining tool must be used to cut this part, which deteriorates the machinability. Therefore, in order to ensure the width of the flow path between the rear part of the pressure surface 8A1 of the front blade 8A and the front part of the negative pressure surface 8B1 of the rear blade 8B to an extent that does not deteriorate the machinability, it is necessary to reduce the angle (θ) between the leading edge 8A3 of the front blade 8A and the trailing edge 8B3 of the rear blade 8B of the return vane 8 and increase the offset amount of the trailing blade 8B toward the pressure surface 8A1 of the front blade 8A, or to shorten the radial length from the center of the rotating shaft 4 of the leading edge 8B2 of the rear blade 8B relative to the trailing edge 8A2 of the front blade 8A to provide a gap in the radial direction.

When the camber line 8A4 of the leading blade 8A is curved sharply near the trailing edge 8A2 as in this embodiment, if one attempts to secure the flow passage width between the rear half of the pressure surface 8A1 of the leading blade 8A and the front half of the suction surface 8B1 of the trailing blade 8B by only reducing the angle (θ) between the leading edge 8A3 of the leading blade 8A of the return vane 8 and the trailing edge 8B3 of the trailing blade 8B, one must increase the reduction in the angle (θ) between the leading edge 8A3 of the leading blade 8A of the return vane 8 and the trailing edge 8B3 of the trailing blade 8B.

In this case, the direction of the flow from the pressure surface 8A1 of the leading blade 8A moves downstream from the front half of the suction surface 8B1 of the trailing blade 8B, where the flow speed on the blade surface slows most and separation is most likely to occur, reducing the effect of suppressing flow separation on the suction surface 8B1.

To avoid this and ensure a flow passage width between the rear half of the pressure surface 8A1 of the leading blade 8A and the front half of the suction surface 8B1 of the trailing blade 8B without deteriorating workability, it is advisable to make the leading edge 8B2 of the trailing blade 8B shorter in radial length from the center of the rotating shaft 4 than the trailing edge 8A2 of the leading blade 8A. In other words, it is better to adopt a means for providing a radial gap between the leading edge 8B2 of the trailing blade 8B and the trailing edge 8A2 of the leading blade 8A.

The centrifugal compressor 100 of this embodiment can maintain and improve efficiency while reducing the outer diameter of the stationary flow path, which is expected to reduce costs and improve operational efficiency. In addition, the reduction in the outer diameter also makes it possible to reduce the floor space required by the centrifugal compressor 100 within the facility.

The present invention is not limited to the above-described embodiments, but includes various modified examples. For example, the above-described embodiments have been described in detail to clearly explain the present invention, and are not necessarily limited to those having all of the configurations described. It is also possible to replace part of the configuration of one embodiment with the configuration of another embodiment, and it is also possible to add the configuration of another embodiment to the configuration of one embodiment. It is also possible to add, delete, or replace part of the configuration of each embodiment with other configurations.

1... centrifugal impeller, 4... rotating shaft, 5... diffuser, 6... return flow passage, 7a, 7b... turning section, 8... return vane, 8A... leading blade of return vane, 8A1... pressure surface of leading blade of return vane, 8A2... trailing edge of leading blade of return vane, 8A3... leading edge of leading blade of return vane, 8A4... camber line of leading blade of return vane, 8A5... negative pressure surface of leading blade of return vane, 8A6... chord line of leading blade of return vane, 8A7... camber of leading blade of return vane, 8A8... maximum camber of leading blade of return vane, 8B... trailing blade of return vane, 8B1... return vane suction surface of the trailing blade of the return vane, 8B2...leading edge of the trailing blade of the return vane, 8B3...trailing edge of the trailing blade of the return vane, 8B4...pressure surface of the trailing blade of the return vane, 9...turning section inlet, 10...turning section outlet, 12...leading edge of the return vane, 15...suction passage, 16...discharge passage, 19...casing, 20a, 20b...flange, 100...centrifugal compressor, L...blade chord length, α...inlet blade angle of the leading blade of the return vane, β...inlet blade angle of the trailing blade of the return vane, θ...angle between the leading edge of the leading blade of the return vane and the trailing edge of the trailing blade, γ...angle between the leading edge of the leading blade of the return vane and the leading edge of the leading blade adjacent in the circumferential direction.

Claims (6)

a return flow passage provided downstream of the diffuser and through which the fluid flowing from the diffuser into the centrifugal impeller at a subsequent stage flows in a return direction toward the rotary shaft; a plurality of return vanes disposed in a circular cascade about a center line of the rotary shaft and installed in the return flow passage; and a turning portion at which the flow of the fluid having flowed through the diffuser turns from the centrifugal direction to an axial direction, and further turns from the axial direction to the return direction ,
a centrifugal compressor in which the return vanes, in which the circular blade rows are provided in multiple rows, are arranged in two rows from the upstream side to the downstream side of the flow of the fluid in the return flow passage, wherein an inlet blade angle (β) of a rear blade provided on the downstream side of the return vanes is inclined in the circumferential direction relative to an inlet blade angle (α) of a front blade provided on the upstream side of the return vanes ,
A plurality of the return vanes having a blade shape are provided in the return flow passage in a circumferential direction as a front blade row and a rear blade row on the upstream and downstream sides of the return flow passage, respectively;
the trailing vane is provided offset toward the pressure surface side of the leading vane in order to guide the flow on the pressure surface side of the leading vane to the suction surface of the trailing vane,
The angle (θ) between the leading edge of the lead vane and the trailing edge of the trail vane is smaller than the angle (γ) between the leading edge of the lead vane and the leading edge of the lead vane adjacent in the circumferential direction, and
A centrifugal compressor characterized in that the maximum camber position of the camber line of the front blade (the position on the chord line at which the distance (camber) of a perpendicular line extending vertically from any position on the straight line (chord line) connecting the leading edge and trailing edge of the blade to the camber line is maximum) is located in the rear half of the chord line . 2. The centrifugal compressor according to claim 1,
a first inlet blade angle (β) of the rear blade and a second inlet blade angle (α) of the front blade that satisfy a relationship of β<α; 3. The centrifugal compressor according to claim 1 or 2,
A plurality of the return vanes having a blade shape are provided in the return flow passage in a circumferential direction as a front blade row and a rear blade row on the upstream and downstream sides of the return flow passage, respectively;
a leading edge of the trailing blade is shorter in radial length from the center of the rotation shaft than a trailing edge of the leading blade, in order to guide a flow on the pressure surface side of the leading blade to the negative pressure surface of the trailing blade. 4. The centrifugal compressor according to claim 3,
a leading edge of the leading blade and a trailing edge of the trailing blade being smaller than an angle (γ) between a leading edge of the leading blade and a leading edge of a circumferentially adjacent leading blade. 5. The centrifugal compressor according to claim 4,
A centrifugal compressor characterized in that a camber line (a line connecting points equidistant from the upper and lower surfaces of the blade) of the front blade has a constant blade angle for 50% or more of a first half portion from the leading edge to the trailing edge of the front blade. 2. The centrifugal compressor according to claim 1 ,
A centrifugal compressor, characterized in that a leading edge of the rear blade has a radial length shorter than a trailing edge of the front blade.

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