JP3380897B2 - Compressor - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a structure of a compressor, and more particularly to a technique for improving the performance and downsizing of the compressor.

[0002]

2. Description of the Related Art Conventional centrifugal compressors and mixed flow compressors used in gas turbines, superchargers, etc. are shown in FIGS.
As shown in (b), a plurality of moving blades D2 are arranged around the rotary drive shaft D1, and a plurality of stationary blades D3 are arranged downstream thereof.

The moving blade D2 is located in the curved flow path D4, and the stationary blade D3 is located downstream of the curved flow path D4. The moving blade D2 is located between the moving blade trailing edge portion D2a and the stationary blade front edge portion D3a. Are connected. The moving blade D2 and the stationary blade D3 are set such that the radial distance from the rotary driving axis D1 of the stationary blade D3 is larger than the radial distance of the moving blade trailing edge portion D2a from the rotary driving axis D1.

[0004]

In such a centrifugal compressor and mixed flow compressor, as shown in FIGS. 6 to 7, when the rotor blade D2 rotates in the rotation direction indicated by the arrow R at the speed V _R , At that time, the flow of the main flow DJ flowing out from the moving blade D2 has a velocity V _J. However, in each rotor blade D2,
The rear surface in the direction of rotation indicated by the arrow R becomes the suction surface, and the mainstream DJ
Wake DW having a velocity V _W less than the velocity V _{J of}
However, as shown in FIG. 6A, the curved flow path D4 of the rotor blade D2 is
Is formed in the intermediate portion of the wing DW, and its area becomes larger toward the wake, and the wake DW is in an expanded flow state at the blade outlet portion as shown in FIG. 6B.

The velocity V _{W at} which the fluid in the wake DW flows out of the rotor blade D2 is smaller than the velocity V _J of the main flow in the same direction as shown in FIG. The state in which the flow state changes with time is a waveform in which the velocity changes greatly with time as shown in FIG. 6C, and the different radial positions r ₁ and r shown in FIG. 6B are used. _{In 2} , the waveforms are also different.

[0006] In the moving blade trailing edge portion D2a, the rate of flow towards the stationary blade direction D3, as shown in FIG. 7, parallelogram diagonals formed by the velocity flowing from the rotor blade and the rotation speed V _R The velocity is V _J1 for the main stream DJ and V _W1 for the wake DK. Although the rotation speeds V _R are the same, the sizes of the velocity V _{J of} the main flow and the velocity V _W of the wake are different, and therefore the sizes of the velocity V _J1 and V _W1 of the flow toward the vane D3 are also slightly different. However, the direction changes greatly, and the angle difference θ shown in FIG. 7 is obtained.

Although such a flow state slows down while passing through the flow path D6, it reaches the vanes with a similar tendency, and as shown in FIG.
The flow of _J1 'and the wake DW portion becomes velocity _VW1 ', and both flow into the stator blade leading edge portion D3a after waiting for a large angle difference θ'corresponding to about the angle difference θ. Therefore, the stationary wing D3
The incident angle with respect to γ changes periodically with each passage of the moving blade D2, and as a result, the operating efficiency of the stationary blade D3 decreases,
Not only the compression performance deteriorates, but in the worst case, the stationary blade D
3 has a problem that the compressor stalls and the compressor enters an abnormal operation called a surge in which the flow rate stops or flows backward.

Further, in the conventional compressor, as shown in FIG.
As shown in FIGS. 5A and 5B, the moving blade trailing edge portion D2
Since the radius from the rotary drive axis D1 of the stationary blade D3 is larger than the radius from the rotary drive axis D1 of a, the moving blade trailing edge portion D2
The substantially circumferential flow passage cross-sectional area around the rotary drive shaft D1 near the vane leading edge portion D3a is larger than the substantially circumferential flow passage cross-sectional area around the rotary drive axis D1 near a. That is, in the flow path D6 between the moving blade trailing edge portion D2a and the stationary blade front edge portion D3a, the flow passage cross-sectional area expands in the flow direction, and the flow has a decelerating and pressure increasing effect. become.

Therefore, in the inside of the flow path D6, as shown in FIG. 5, when the boundary layer D6BL develops on the inner side and the outer side in the radial direction, the compression performance is deteriorated, and the boundary layer separates. However, there is a problem in that it acts to close the flow path D6 and rushes into the surge. Further, in the flow toward the stationary blade D3 in the flow path D6, the pressure increases in the flow direction due to the diffuser effect,
It must flow to overcome the pressure on the downstream high pressure side.

FIG. 8 shows the main flow DJ and the windshield at the rotor blade outlet.
Wake DW speed V_J1And V_W1The circumferential velocity of the rotor blade rotation direction
Minute V_J1TAnd V_W1TAnd towards the vane D3 perpendicular to them
Vertical velocity component V, which is the direction of flow_J1LAnd V_W1LMinutes
A split view is shown. Vertical velocity component V of mainstream DJ
_J1LIn comparison with the vertical velocity component V of the wake DW. _W1LIs
Because it is small, it cannot overcome the pressure on the high pressure side.
In this case, backflow or the like may occur in the flow path D6.
Therefore, the effective flow path cross-sectional area cannot be secured and the flow path may be blocked.
If not only the compression performance deteriorates, but it is severe
Had a problem that led to a surge.

Further, as shown in FIGS. 5 (a) and 5 (b), in the conventional compressor, the rotary drive shaft D1 of the stator blade D3 is rotated.
Is located farther than the position of the trailing edge D2a of the moving blade, and it is necessary to install an air collecting pipe such as a scroll on the outer peripheral portion of the outlet of the stationary blade. It requires a very large compressor outer diameter, and has a compressor structure that is difficult to miniaturize.

The present invention has been made in view of the above circumstances, and aims to achieve the following objects. (1) To suppress the temporal variation of the incident angle with respect to the stationary blade. (2) To suppress the development or separation of the boundary layer in the flow path between the moving blade and the stationary blade. (3) To improve the performance of the compressor as described above. (4) To reduce the size of the compression structure.

[0013]

SUMMARY OF THE INVENTION The present invention rotationally drives a moving blade having a curved flow path that is bent radially outward from a rotary drive shaft, and radially moves from a stationary blade that is located radially downstream of the moving blade. In the conventional centrifugal compressor for discharging to and the conventional mixed flow compressor for discharging in a diagonal direction from a stationary blade located in a diagonal direction between the radial direction and the rotation drive shaft direction,
By bending the flow path near the outlet of the rotor blade in the axial direction, letting it flow out in the axial direction from the outlet of the rotor blade, and discharging in the substantially axial direction from the stator blade located at the same radial position as the outlet of the rotor blade,
The above problem is solved.

[0014]

BEST MODE FOR CARRYING OUT THE INVENTION The compressor according to the present invention is applied to a mixed flow compressor, a centrifugal compressor, a blower, a pump, and the like. Hereinafter, a mixed flow compressor will be described as an example with reference to the drawings. To do. FIG. 1 is a configuration diagram showing an example in which the compressor according to the present invention is applied to a turbojet engine. The turbojet engine of FIG. 1 compresses the air introduced from the direction of the arrow in the axial compressor (a) along the outer periphery of the rotary shaft (f) forming the streamline shape, and then according to the present invention described later. It is compressed by the compressor (b), guided to the combustion chamber (c) and burned, and the turbine (d) is driven to exhaust from the exhaust nozzle (e) in the arrow direction. 2 to 4 are
The compressor which concerns on this invention is shown, The code | symbol 1 is a rotary drive shaft, 2 is a moving blade, 3 is a stationary blade, 4 is a curved flow path, 5 is an axial flow path, 6 is a flow path between a moving blade and a stationary blade. , C are casings.

In the mixed flow compressor of this embodiment, as shown in FIG. 2, a rotation transmission member 1a for transmitting the rotation of the rotation drive shaft 1 is arranged around the rotation drive shaft 1, and the rotation transmission is carried out. A plurality of moving blades 2 are arranged on the member 1a, and a plurality of stationary blades 3 are arranged downstream of the member 1a substantially in parallel with the rotary drive shaft 1.

The curved flow path 4 includes a rotation transmitting member 1a and a moving blade 2.
It is formed by being surrounded by the casing C and the casing C, and is inclined outward in the radial direction from the axial direction of the rotary drive shaft 1. The curved portion 4 is provided on the downstream side of the curved flow path 4.
a is arranged and is connected to the axial flow path 5 via the curved portion 4a. The curved portion 4a is located between the curved flow passage 4 and the axial flow passage 5, and is smoothly curved in the axial direction. The upstream portion side is connected to the curved flow passage 4 and the downstream portion side is the axial flow passage. Connected to 5. In this axial flow path 5, the flow path 6 connecting the stationary blade 3 and the moving blade and the stationary blade 3 are located.

The outlet of the moving blade 2 is extended to the inlet or the inside of the axial passage 5 so that the trailing edge 2a of the moving blade is located at the inlet or the inside of the axial passage 5. The radius H _{1 of the} trailing edge portion 2a from the rotary drive shaft 1 is set to be substantially equal to the radius H _{3 of} the stationary blade 3 from the rotary drive shaft 1.

In such a compressor, as shown in FIG. 2, the rotary drive shaft 1 rotates to rotate the moving blades 2 to suck in air, and the inflowing air to the curved flow path 4 and the curved portion 4a. It is compressed by passing it. On this occasion,
In the curved flow path 4 and the curved portion 4a from the rotor blade inlet portion 2b to the rotor blade trailing edge portion 2a, the flow is from the rotor blade inlet portion 2b having a radial distance close to the rotary drive shaft 1 to the rear of the rotor blade having a radial distance. While moving to the edge portion 2a, it is accelerated in the circumferential direction by the rotation and imparts energy to the inflow air by the rotational driving force to compress the inflow air.

The relatively high speed discharge air from the moving blades is introduced into the axial flow passage 5, the speed is reduced by the stationary blades 3, and the dynamic pressure is converted into the static pressure to further increase the pressure. Is discharged.

[0020] Here, since the radius H ₁ of the blade trailing edge portion 2a with respect to the rotation drive shaft 1 as described above, is set to be substantially equal to the radius H ₃ of the stationary blade 3 with respect to the rotary drive shaft 1, the rotor blade In the vicinity of the trailing edge portion 2 a and the stationary blade 3, the fluid flowing from the moving blade 2 to the stationary blade 3 moves in the axial direction at a position substantially equal to the radial distance from the rotary drive shaft 1. Therefore, in the flow path 6 between the moving blade trailing edge portion 2a and the stationary blade 3, the flow path cross-sectional area does not expand in the radial direction from the rotary drive shaft 1 and the so-called diffuser effect is not exerted. Not only can the development of the boundary layer D6BL as shown in Fig. 5 be suppressed to improve the performance of the compressor, but also the separation of the boundary layer can be suppressed, so that the compressor is less likely to enter a surge. The performance of the compressor can be improved.

Further, in each moving blade 2, as shown in FIG. 3A, when the moving blade 2 rotates in the direction indicated by the arrow R,
The rear surface of each rotor blade 2 in the direction of rotation indicated by the arrow R serves as a negative pressure surface, and at the intermediate portion or the outlet portion of the curved flow path 4 (the inlet portion of the curved portion 4a), the wake DW having a velocity smaller than the velocity of the main flow DJ. Is formed, and the flow state is similar to that of the conventional compressor shown in FIG.

When the mainstream DJ and the wake DK pass through the bending portion 4a, the bending portion 4a is curved so as to be directed in the axial direction.
Centrifugal force acts on the outer side in the radial direction indicated by the arrow r. Here, the velocity of the mainstream DJ is V _J , the density of the mainstream DJ is ρ _J , the density of the wake DW is ρ w, and the velocity of the wake DW is V _W ,
When the radius of the curved portion 4a is r, the centrifugal force F of the main flow DJ generated in the radial direction indicated by the arrow r in the curved portion 4a.
_J becomes F _J = ρ _J V _J ² / r (Equation 1), and similarly, the centrifugal force F _W of the wake DW becomes F _W = ρwVw ² / r (Equation 2).

Here, the flow velocity V _J of the main flow DJ is larger than the flow velocity V _{w of the} wake DK, and in the incompressible fluid, the density ρ _{J of the} main flow DJ and the density ρ _{w of the} wake DW are equal, and Since they are almost equal even in the fluid, as is clear from the above equations 1 and 2, the main flow D
The centrifugal force F _{J of J} becomes larger than the centrifugal force Fw of the wake DW. Therefore, due to the difference between these centrifugal forces F _J and F _W , as shown by the arrow S in FIG. 3 (a), the main flow DJ moves radially outward in the curved portion 4 a, and the wake DW spreads radially inward. As shown in FIG. 3 (b), the main flow DJ is positioned radially outside from the rotary drive shaft 1 and the wake DK is positioned radially inward in the vicinity of the rotor blade outlet. Is discharged.

Therefore, at the outlet of the rotor blade, the mainstream DJ
And the wake DW are divided in the radial direction and are distributed in a substantially uniform state in the circumferential direction, as shown in FIG. As a result, the flow velocity V _{J of the} main flow DJ and the flow velocity V _{W of the} wakes DK have a substantially uniform distribution with little temporal variation, as shown in FIG. Here, small wakes due to the boundary layer developed on the blade surface of the moving blade 2 in each of the mainstream DJ and the wake DK are shown in FIGS.
In both cases (c), they are periodically present,
It is smaller than the wake DW.

Since the flow passage 6 forms an axial flow passage having substantially the same flow passage cross-sectional area, the state of the passing flow does not change so much, and the flow passage 6 is kept static while maintaining the flow state at the blade outlet. The incident angles of the fluid flowing into the vanes 3 into the main flow DJ and the wakes WD are substantially uniform at each radial position in the respective portions of the main flow DJ and the wakes WD, and the time variation of the incident angle of the fluid flowing into the vanes 3 can be suppressed. .

When the inlet shape of the stationary blade 3 is located radially inward of the rotary drive shaft 1, the inflow velocity V of the wake DW is V.
_The shape is set to correspond to the incident angle of _W1 , and the inflow velocity V of the main flow DJ is at the position radially outward from the rotary drive shaft 1.
By setting the shape corresponding to the incident angle of _J1 , it is possible to optimize the incident angle of the stationary blade 3, improve the performance of the stationary blade, and improve the performance of the compressor as a whole. Becomes

Further, the pressure ratio of the compressor depends on the difference in the radial position of the rotor blade leading edge portion 2b and the rotor blade trailing edge portion 2a from the rotary drive shaft 1 as shown in FIG. For comparison, FIG. 4 shows a conventional compressor shape capable of producing a pressure ratio equivalent to that of this embodiment by a broken line. As shown in FIG.
In this embodiment, the radius H _{3 of the} stationary blade 3 is equal to the trailing edge 2 of the moving blade.
It is set to be substantially equal to the radius H _{1 of a,} and an air collecting pipe such as a scroll can be arranged downstream in the axial direction thereof, and the outer diameter of the compressor can be made substantially equal to the outlet diameter of the moving blade. However, in the conventional type, the flow path 6 ′ connected to the outlet portion of the moving blade 2 ′ and the following stationary blade 3 ′ are connected in the direction in which the radius with respect to the rotary drive shaft 1 expands, and the radius H _{2 of the} stationary blade outlet portion is increased. Is larger than the stator blade outlet radius H ₃ of the present embodiment, and it is necessary to dispose an air collecting pipe such as a scroll further downstream thereof so as to expand in the radial direction.
The outer diameter of the compressor is much larger than the diameter of the blade outlet.
That is, in the present embodiment, the outer diameter dimension of the entire compression structure can be made smaller and the size can be reduced as compared with the conventional type which produces an equivalent pressure ratio.

[0028]

The compressor of the present invention has the following effects. (1) Since the rotor blade outlet portion is curved in the axial direction, the centrifugal force generated at the curved portion causes the main flow to be distributed radially outward and the wakes radially inward, so that the circumference at the rotor blade outlet is increased. It is possible to set the directional velocity distribution in a substantially uniform state in the circumferential direction, and as a result, it is possible to suppress the time variation of the incident angle with respect to the stationary blade. (2) The trailing edge portion of the moving blade outlet side is located at the inlet or inside of the axial flow passage, and the moving blade outlet portion, the succeeding flow passage and the stationary blade can be positioned on substantially the same radius, Since the flow path and the vane are not set so that the area expands in the radial direction and the area expands in the flow direction as in the past, and there is no so-called diffuser effect, it is possible to suppress the development or separation of the boundary layer. it can. (3) The performance of the compressor can be improved by the above. (4) Since the radius of the stationary blade can be set to be substantially equal to the radius of the moving blade outlet, the outer diameter of the compressor can be reduced without reducing the compression capacity of the compressor.

[Brief description of drawings]

FIG. 1 is a diagram showing a configuration in which a compressor according to the present invention is applied to a turbojet engine.

FIG. 2 is a sectional view showing a compressor according to an embodiment of the present invention with a part of the description omitted.

FIG. 3 shows an embodiment of a compressor according to the present invention, and (a) is a developed sectional view in the circumferential direction of the moving blade curved passage intermediate portion,
FIG. 4B is a development cross-sectional view of the rotor blade outlet portion in the circumferential direction, and FIG. 8C is a graph showing the temporal change of the fluid velocity at the rotor blade outlet.

FIG. 4 is a cross-sectional view showing a comparison between the compressor of FIG. 2 and a conventional type compressor capable of producing a pressure ratio equivalent to that, and omitting a part of the description.

FIG. 5 is a cross-sectional view in which a part of the conventional compressor is omitted.

6A and 6B show a conventional compressor, in which (a) a developed sectional view in the circumferential direction of a moving blade curved passage intermediate portion, (b) a developed sectional view in the circumferential direction of a moving blade outlet portion, and (c) a moving blade. It is a graph which shows the time change of the velocity of the fluid in an exit.

FIG. 7 is a schematic diagram showing the velocities of the main flow and the wake fluid at the rotor blade outlet and the stator blade inlet of the compressor, the velocity components, and their angular differences.

FIG. 8 is a schematic diagram in which the velocities of the main flow and the wake fluid at the rotor blade outlet of the compressor are divided into the rotor blade rotation direction (circumferential direction) and the velocity component in the flow direction (longitudinal direction) perpendicular to the rotor blades (vertical direction). is there.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 rotary drive shaft 1a rotation transmission member 2 moving blade 2a moving blade trailing edge 2b moving blade leading edge 3 stationary blade 4 curved flow path 4a curved portion 5 axial direction flow path 6 flow path C between moving blade and stationary blade Casing DJ Main flow DW Wake H Radial distance from rotary drive shaft R Rotating blade rotation direction r Moving blade outlet radius V _J Main flow velocity V _w Wake velocity

Claims (2)

(57) [Claims] 1. A compressor having a rotating blade and a stationary blade positioned downstream of the outlet of the rotating blade, the compressor having a curved flow path bent radially outward from a rotary drive shaft. The compressor is characterized in that the shape of the flow path near the outlet of the moving blade is curved in the axial direction so that the fluid flowing out of the compressor flows out substantially in the axial direction. 2. In the compressor, the vanes and the flow path connecting the rotor vane outlet and the stator vane are located substantially in the axial direction at a radial position substantially equal to the radius of the rotor vane outlet from the rotary drive shaft. The compressor according to claim 1, wherein: