JP2007247621A - Centrifugal fluid machine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent rotating stall in a centrifugal compressor for small flow rate high pressure usage. <P>SOLUTION: In the centrifugal compressor 100, a centrifugal impeller 3 is attached on a rotary shaft 2. A diffuser 8 part including vanes 13 between opposing two wall surfaces is formed in a downstream side of the centrifugal impeller. Many curved minute grooves 21 are formed on at least one of the opposing two surfaces with keeping intervals in a circumference direction. Depth of the curved grooves is preferably set to depth equivalent to thickness of a boundary layer growing between the opposing two wall surfaces. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a centrifugal fluid machine, and more particularly to a centrifugal compressor suitable for high pressure and low flow rate applications.

  An example of a conventional centrifugal fluid machine is described in Non-Patent Document 1. In this known document, in order to reduce the rotational stall of a low specific speed centrifugal compressor having a vaneless diffuser, a vaned diffuser with a low chord ratio is positioned between 1.03 and 1.2 times the impeller diameter. It is arranged on the downstream side. The chord ratio of the vanes of this small chord ratio diffuser is 0.68 to 0.87.

Nishida et al., "Study on Rotating Stall of Centrifugal Compressor (3rd Report, Suppression Method of Rotating Stall)", Transactions of the Japan Society of Mechanical Engineers (B), Vol.57, No.543, pp.154-160, November 1991

  In the centrifugal compressor described in Non-Patent Document 1, by installing a diffuser blade having a small chord joint ratio near the impeller, the static pressure recovery efficiency in the diffuser is improved in the small flow rate specification, and the rotation stall is also prevented. The effect of being able to be obtained is obtained. However, the closer the diffuser blade is to the impeller outlet, the more the periodic velocity distribution and pressure distribution due to the number of impeller blades acts on the diffuser blade as a large fluid fluctuation force at the impeller outlet. . This fluid fluctuation force increases noise and impact loss of the diffuser blades.

  By the way, in the centrifugal compressor of a small flow rate and a high pressure specification, since the ratio of the impeller exit width with respect to the impeller diameter is small, the casing thickness is relatively thick. Also, if there are diffuser blades, fluid flows along the diffuser blades between the diffuser blades, and therefore flows with relatively little friction loss. For this reason, the noise caused by the collision with the diffuser blade has not been considered much. However, it is known that turning the diffuser blade closer to the impeller exit can avoid turning stall, and therefore it is desirable to further bring the diffuser blade closer to the impeller. There is a risk of becoming.

  The present invention has been made in view of the above problems of the prior art, and an object thereof is to prevent an increase in noise while avoiding a rotating stall in a centrifugal compressor. Another object of the present invention is to prevent an increase in noise while increasing the pressure recovery efficiency in the diffuser.

  A feature of the present invention that achieves the above object is that, in a centrifugal fluid machine having a centrifugal impeller attached to a rotating shaft, a diffuser portion in which a large number of curved grooves are formed at intervals in the circumferential direction on the downstream side of the centrifugal impeller. It is to have done.

  In this feature, the diffuser portion has two wall surfaces facing each other, and it is preferable to form a large number of curved grooves on at least one wall surface, and in the mountain line groove, the flow exiting the centrifugal impeller occurs in the diffuser portion. The depth should be on the order of the thickness of the boundary layer. Further, the diffuser portion may have a large number of vanes arranged at intervals in the circumferential direction, and the curved groove may be formed between the vane and the impeller.

  Another feature of the present invention that achieves the above object is a centrifugal fluid machine having a centrifugal impeller attached to a rotating shaft, and having a diffuser portion formed of two wall surfaces facing the downstream side of the centrifugal impeller. The boundary layer thickness and depth generated on the wall surface forming the section are substantially the same, and a large number of curved grooves curved in the flow direction are formed at intervals in the circumferential direction.

  And in this feature, the diffuser part has a plurality of diffuser vanes arranged at intervals in the circumferential direction, and the boundary layer thickness and depth are approximately on the wall surface of the vaneless part between the centrifugal impeller and the diffuser vane. The same curved groove should be provided, and the diffuser section has a plurality of diffuser vanes arranged at intervals in the circumferential direction, and the boundary layer thickness and depth are formed on the wall surface of the vaneless section downstream of the diffuser vane. May have substantially the same curved grooves. The curved groove may be formed on a wall surface facing the wall surface to which the diffuser vane is attached, and the flow path width of the diffuser is 4% or less of the radius of the impeller, and the facing wall surface forming the diffuser A curved groove may be formed in at least one of these.

  According to the present invention, since the curved groove of the boundary layer is provided in the vicinity of the impeller outlet of the small flow high pressure compressor, it is possible to prevent an increase in noise while improving the pressure recovery efficiency in the diffuser. Moreover, turning stall can be avoided.

  Hereinafter, an embodiment of a centrifugal compressor according to the present invention will be described with reference to the drawings. FIG. 4 is a longitudinal sectional view of a high-pressure centrifugal compressor 100 used in a petrochemical plant or the like. The centrifugal compressor 100 is a single-shaft multi-stage compressor, and five centrifugal impellers 3 are attached to one rotating shaft 2. Radial bearings 4 are provided at both ends of the rotary shaft 2, and the radial bearings 4 support a radial load. In the vicinity of one radial bearing 4, a thrust bearing 5 for supporting an axial load is provided. The rotating shaft 2 is rotationally driven by a driving machine such as a motor (not shown).

  A vaned or vaneless diffuser 8 is arranged on the downstream side of each impeller 3 in the radial direction, and converts the velocity energy of the working gas compressed by the impeller 3 into pressure energy. On the outlet side of the impeller 3, which is the inner peripheral side of the vaned diffuser 8, a curved groove whose details will be described later is formed. The flow of the working gas exiting the diffuser 8 flows inward in the radial direction in the return channel 9 through a U-shaped cross section flow path.

  A suction nozzle 7 is provided at one position in the radial direction on the suction side of the first stage impeller 3. On the downstream side of the diffuser 8 of the final stage impeller 3, there is provided a discharge nozzle for collectively sending the flow exiting the diffuser 8 to the outside of the apparatus. The casing 1 accommodates the impeller 3, the rotating shaft 2, and the bearings 4 and 5. The diffuser 8 and the return channel 9 are formed by diaphragms 11 and 12 in this embodiment.

  In the multistage centrifugal compressor 100 configured as described above, the working gas sucked from the suction nozzle 7 is compressed a plurality of times in the compression stage having the impeller 3, the diffuser 8, and the return channel 9, and five times in this embodiment. . However, the last stage does not have the return channel 9. The flow exiting the final stage diffuser 8 is discharged from the discharge nozzle 10 to the outside of the apparatus.

  FIG. 5 is a longitudinal sectional view showing details of one impeller 3 portion of the centrifugal compressor 100 shown in this embodiment. Further, FIG. 6 shows an AA cross section of FIG. FIG. 6 also shows a flow relationship. When the driving machine rotates the rotary shaft 2, energy is transmitted from the impeller 3 to the working fluid in the form of angular momentum.

  At this time, the working fluid in the impeller 3 rises in the form of an increase in static energy and kinetic energy in the form of velocity energy. A part of the velocity energy of the working fluid exiting from the impeller 3 is further converted into static pressure in the diffuser 8. The working fluid is gradually redirected to a radial flow in the return channel 9, and flows in the axial direction at the inlet of the next stage impeller 3.

  In the single-shaft multistage centrifugal compressor 100 shown in the present embodiment, a stall region (cell) is generated in the flow of the working fluid in the diffuser 8, and the stalled cell swirls in the rotation direction to cause a phenomenon of pressure pulsation and excessive rotor vibration. May result. The reason will be described with reference to FIG. In FIG. 5, u is a circumferential speed, w is a relative flow speed along the impeller 3, c is an absolute flow speed, and α is a flow angle. The subscript 2 indicates the outer peripheral portion of the impeller 3, the subscript u indicates the circumferential speed component, and the subscript m2 indicates the radial speed component. The radius of the impeller 3 is r2.

  At the outlet of the impeller 3, the flow can be expressed by a relative flow velocity w2 that flows substantially along the blade 3a that is disposed in the impeller 3 at intervals in the circumferential direction and an outlet peripheral velocity u2 of the impeller 3. . When the output flow of the impeller 3 is vector-combined, an absolute speed c2 is obtained, and the working fluid flows into the diffuser 8 at the absolute speed c2. The absolute velocity c2 at the exit of the impeller 3 can be decomposed into a radial flow velocity cm2 and a circumferential velocity cu2.

  Thus, the energy conservation law can be applied to the flow cu of the working fluid at an arbitrary radial position r in the diffuser 8 if there is no loss, and is expressed by the following equation from the law of constant angular momentum.

r × cu = r2 × cu2 (1)
Similarly, when the energy conservation law is applied to the radial velocity cm and the absolute velocity c in the diffuser 8, the radial velocity cm and the absolute velocity c are reduced according to the radial length. If the channel width of the diffuser 8 and the outlet channel width of the impeller 3 are set to be the same, the flow angle α does not change.

  In the diffuser 8, the absolute speed w is reduced as the working fluid flows in the outer diameter direction, and the amount of deceleration is converted into an increase in static pressure. When the flow is decelerated, a deceleration loss occurs, and the static pressure recovery is not as great as the value obtained based on the above equation (1). However, since the pressure increases in the flow direction, a reverse pressure gradient is generated for the flow of the working fluid.

  In the compressor 100 with a small flow rate, the flow path widths of the impeller 3 and the diffuser 8 are small. As a result, in the flow of the working fluid in the diffuser 8, the influence of the boundary layer on the main flow is large. Since the flow rate is small, the radial velocity component cm2 of the flow related to the flow rate tends to be small. In that case, the flow angle α of the diffuser 8 becomes small.

  In this small flow rate compressor 100, the radial velocity component cm of the flow is reduced by the wall friction in the vicinity of the wall surface of the diffuser 8. And if the flow velocity of the working fluid continues to decrease, the working fluid cannot withstand the reverse radial pressure gradient and starts to flow backward. This is the cause of the stalled cell. When a stall cell occurs, the effective flow area in the diffuser 8 decreases, and the average flow rate of the working fluid increases. As a result, the flow of the working fluid is stabilized in the healthy part other than the stalled cell part.

  When the stall cell moves in the direction in which the rotating shaft 2 rotates, the flow angle α of the working fluid increases on the side where the stall cell passes, and the static pressure recovery of the flow proceeds in the diffuser 8. On the other hand, in the direction in which the stall cell advances, the flow angle α of the working fluid decreases, the flow increases, and the static pressure recovery decreases. When the stall cell turns in the rotation direction of the rotating shaft 2, the static pressure distribution in the diffuser 8 deviates from the axially symmetric distribution, so that an external force of non-uniform force is applied to the rotating body having the rotating shaft 2 and the impeller 3. appear. This external force may induce forced vibration in the rotating body.

Therefore, as shown in FIG. 1, a so-called curved groove is provided on the outlet side of the impeller 3 where the inlet flow angle of the diffuser 8 is approximately 100 or less. FIG. 1 is a view corresponding to the AA cross-sectional view of FIG. 5, and the diffuser 8 is an example of the vaned diffuser 8. In order to efficiently convert velocity energy into pressure energy in the diffuser 8, a plurality of vanes 13 are arranged in the diffuser 8 at intervals in the circumferential direction. The flow angle α at the outlet of the vane 13 is larger than the inflow angle α2 at the inlet of the vane 13. Within the flow path of the diffuser 8, the flow length of the working fluid is shortened to reduce friction loss.

A vaneless diffuser is formed from the flow path inlet of the diffuser 8 to the vane 13 inlet. A large number of spiral grooves 21 are formed on the wall surface of the vaneless diffuser having the same entrance angle as the flow angle α at the exit of the impeller 3 and along the constant angular momentum line represented by the above formula (1). To do. The size of the groove 21 is almost the same as the thickness of the boundary layer, and is approximately 0.01.
From mm to 0.1 mm. The spiral grooves 21 are formed at intervals in the circumferential direction.

  Since the spiral groove 21 is formed at the outlet of the impeller 3, the wall surface friction is reduced and the friction loss is reduced. The spiral groove 21 stabilizes the turbulent boundary layer generated on the wall surface of the diffuser 8 and prevents separation. Thereby, the entrance diameter of the vane 13 of the diffuser 8 can be brought close to the exit of the impeller 8, and the pressure fluctuation or the speed fluctuation flow generated for each number of blades of the impeller 8 collides with the vane 13 of the diffuser 8. Can reduce pulsation, vibration and noise.

  In the above embodiment, the spiral groove 21 is provided on one side of the wall surface forming the flow path of the diffuser 8, but it may be provided on both wall surfaces. If the spiral groove 21 is provided on the wall surface facing the wall forming the vane 13 of the diffuser 8, the processing becomes easy.

  Another example of the flow stabilizing means used in the centrifugal compressor 100 according to the present invention is shown in a cross-sectional view in FIG. In this embodiment, a non-vane diffuser 8 formed at the outlet of the vane 13 of the diffuser 8 is provided with a spiral groove that maintains the flow angle α at the outlet of the vane 13. According to this embodiment, the friction loss in the vaneless diffuser 8 flow path is reduced, and the occurrence of peeling can be suppressed.

  In this embodiment, the spiral grooves are formed on the opposing wall surfaces forming the diffuser 8, but the spiral grooves may be formed only on the wall surface of the diffuser 8 where the vanes 13 are not formed. Further, as shown in FIG. 3, the diffuser 8 is a vaneless diffuser, and a large number of spiral grooves having a constant flow angle α are formed at intervals from the outlet of the impeller 3 in the radial direction. Also good.

  When the value obtained by dividing the flow path width of the vaneless diffuser 8 by the radius of the impeller 3 is approximately 4% or less, a curved groove extending radially outward from the flow path inlet of the diffuser 8 is formed. May be. The curved groove has a depth approximately equal to the thickness of the boundary layer formed on the flow path wall surface of the diffuser 8 and extends along the mainstream flow direction.

  The curved groove reduces friction loss in the diffuser 8 and stabilizes the turbulent boundary layer. As a result, even if the vane 13 is not formed in the diffuser 8, the flow passage width of the diffuser 8 can be made narrower than the exit width of the impeller 3, and the flow angle α can be increased. Can be avoided.

  As shown in the above embodiments, when a number of spiral grooves or curved grooves are formed in the diffuser 8 provided on the outlet side of the impeller 3 at intervals in the circumferential direction, the inlet of the vane 13 of the vaned diffuser 8 is formed. The side can be brought close to the exit of the impeller 3. FIG. 7 shows experimental results obtained by normalizing the horizontal axis by dividing the flow path width of the diffuser 8 and the vertical axis by dividing the backflow start position in the diffuser 8 by the radius of the impeller 3.

  The backflow start test result in the vaneless diffuser 8 is a line rrev / r2, and when the vane 13 of the diffuser 8 is attached below the line, no backflow occurs. In the figure, a circle mark is a condition in which a reverse flow is generated if a vaneless diffuser is used. However, since the inlet of the vane 13 is provided at the position indicated by the circle mark, the reverse flow is suppressed.

  If the vane 13 of the diffuser 8 is brought close to the impeller 3, collision loss increases and noise increases. Therefore, the installation radius position is set to about 1.1 times the radius of the impeller 3 to avoid them. In the small-flow high-pressure compressor, the vane 13 position of the diffuser 8 can be brought close to the impeller 3 to prevent the rotation stall. Further, it can be seen that in the small flow rate and high pressure compressor, the flow is separated due to the influence of wall surface friction even during a short distance from the flow path inlet of the diffuser 8 to the inlet of the vane 13. Therefore, it is also effective from this point of view to form a spiral groove or a curved groove and bring the inlet side of the vane 13 of the vaned diffuser 8 closer to the outlet of the impeller 3.

The cross-sectional view of one embodiment of the diffuser according to the present invention. The cross-sectional view of another embodiment of the diffuser according to the present invention. The cross-sectional view of the further another Example of the diffuser which concerns on this invention. The longitudinal cross-sectional view of one Example of the multistage centrifugal compressor which concerns on this invention. The fragmentary longitudinal cross-section of the multistage centrifugal compressor shown in FIG. It is AA sectional drawing of FIG. 4, and is a figure explaining a flow. The figure explaining the relationship between the inlet position of a diffuser, and a backflow.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Casing, 2 ... Rotating shaft, 3 ... Impeller, 3a ... Blade, 4 ... Radial bearing, 5 ... Thrust bearing, 6 ... Shaft seal device, 7 ... Suction nozzle, 8 ... Diffuser, 9 ... Return channel, 10 ... Discharge nozzle, 11 ... front diaphragm, 12 ... rear diaphragm, 13 ... diffuser vane, 21-23 ... minute groove, 100 ... compressor.

Claims (9)

  A centrifugal fluid machine having a centrifugal impeller attached to a rotating shaft, wherein a diffuser portion having a plurality of curved grooves formed at intervals in the circumferential direction is disposed downstream of the centrifugal impeller.   2. The centrifugal fluid machine according to claim 1, wherein the diffuser portion has two opposing wall surfaces, and the multiple curved grooves are formed on at least one of the wall surfaces.   2. The centrifugal fluid machine according to claim 1, wherein the curved groove has a depth on the order of the thickness of a boundary layer in which a flow exiting the centrifugal impeller is generated in a diffuser portion.   The centrifugal fluid machine according to claim 2, wherein the diffuser portion includes a plurality of vanes arranged at intervals in a circumferential direction, and the curved groove is formed between the vane and the impeller.   In a centrifugal fluid machine having a centrifugal impeller attached to a rotating shaft, a downstream portion of the centrifugal impeller has a diffuser portion formed by two opposing wall surfaces, and a boundary layer thickness generated on the wall surface forming the diffuser portion A centrifugal fluid machine characterized in that a large number of curved grooves having substantially the same depth and curved in the flow direction are formed at intervals in the circumferential direction.   The diffuser portion has a plurality of diffuser vanes arranged at intervals in the circumferential direction, and a boundary layer thickness and depth are substantially the same on the wall surface of the vaneless portion between the centrifugal impeller and the diffuser vane. The centrifugal fluid machine according to claim 5, further comprising a groove.   The diffuser portion has a plurality of diffuser vanes arranged at intervals in the circumferential direction, and a wall surface of a bladeless portion downstream of the diffuser vane is provided with a curved groove having substantially the same boundary layer thickness and depth. The centrifugal fluid machine according to claim 5.   The centrifugal fluid machine according to claim 6 or 7, wherein the curved groove is formed on a wall surface facing a wall surface to which the diffuser vane is attached. 6. The centrifugal fluid machine according to claim 5, wherein a flow path width of the diffuser is 4% or less of a radius of the impeller, and a curved groove is formed on at least one of opposing wall surfaces forming the diffuser. .

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