JP6860331B2 - Diffuser, discharge channel, and centrifugal turbomachinery - Google Patents

Description

The present invention relates to a diffuser, a discharge channel, and a centrifugal turbomachine.

As the diffuser and the discharge flow path of the conventional centrifugal compressor, for example, there is a general one described in Patent Document 1. The centrifugal compressor described in Patent Document 1 includes a diffuser composed of an axial suction flow path, a centrifugal impeller, a radial flow path, and a curved flow path connecting the radial flow path and the discharge flow path. It is composed of a discharge flow path having a circular flow path cross-sectional shape with a constant flow path outer diameter.

However, as shown in FIG. 1 of Patent Document 1, by applying the conventional technique of extending the diffuser vane to the curved flow path connecting the radial flow path and the discharge flow path to the diffuser, the deceleration effect of the diffuser is conventionally obtained. Is larger than. As a result, the flow velocity when the compressed fluid flows out to the discharge flow path can be slowed down as compared with the conventional case, so that high efficiency can be obtained without increasing the pressure loss.

Japanese Unexamined Patent Publication No. 2009-264136 (paragraphs 0024, 0025, FIG. 1, etc.)

By the way, in the diffuser and the discharge flow path described in Patent Document 1, the deceleration effect of the diffuser is further enhanced by extending the diffuser vane to the curved flow path connecting the radial flow path and the discharge flow path. However, if the deceleration effect is increased by the diffuser vane, the efficiency at the design flow rate will improve. On the other hand, when the centrifugal compressor is operated on the flow rate side smaller than the design flow rate, backflow is likely to occur in the diffuser portion where the deceleration is large. Therefore, the operable flow rate range from the design point flow rate to the small flow rate, which is limited by the turning stall in which the flow is turbulent or backflow occurs, or the surging in which the backflow is high and vibration is generated, becomes narrow.

The present invention has been devised in view of the above circumstances, and an object of the present invention is to provide a diffuser, a discharge flow path, and a centrifugal turbomachine that can improve the efficiency of the centrifugal turbomachine without narrowing the operable flow rate range by a simple structure. And.

In order to solve the above problems, the discharge flow path of the first centrifugal turbo machine of the present invention connects the radial flow path, the radial flow path and the discharge flow path that guides the flow in the circumferential direction, and rotates the impeller. It is composed of a curved flow path that bends in a direction parallel to the axial direction, the swirling flow flow velocity at the outlet end of the radial flow path is Vdif, and the circumferential cross-sectional area of the discharge flow path is maximized. When the circumferential flow velocity in the portion is Vd, the circumferential flow path of the discharge flow path is formed so as to have a relationship of 1.1 ≦ Vdif / Vd ≦ 2.0 .

The diffuser of the second centrifugal turbo machine of the present invention connects the radial flow path, the radial flow path and the discharge flow path that guides the flow in the circumferential direction, and is a flow that bends in a direction parallel to the rotation axis direction of the impeller. When it is composed of a curved flow path which is a road, the inner peripheral side radius of the curved flow path is R, and the axial flow path width of the radial flow path is h, 0.2 ≦ R / h ≦ 0. When there is a relationship of .5 and the swirling flow flow velocity at the outlet end of the radial flow path is Vdif and the circumferential flow velocity in the portion where the circumferential cross-sectional area of the discharge flow path is maximum is Vd, 1. The circumferential flow path of the discharge flow path is formed so as to have a relationship of 1 ≦ Vdif / Vd ≦ 2.0.

Centrifugal turbomachine third invention is provided with a Difuyu The second of the present invention.

Centrifugal turbomachine fourth invention has ingredients Bei the discharge flow path of the first present invention.

According to the present invention, it is possible to provide a diffuser, a discharge flow path, and a centrifugal turbomachine that can improve the efficiency of the centrifugal turbomachine without narrowing the operable flow rate range by a simple structure.

FIG. 5 is a cross-sectional view showing a flow path shape of a single-stage centrifugal compressor provided with a diffuser and a discharge flow path of the centrifugal compressor according to the first embodiment of the present invention. The schematic diagram which shows the discharge flow path which guides the flow in the circumferential direction downstream of a diffuser. (A) is a cross-sectional view of a diffuser portion showing the diffuser according to the first embodiment in comparison with the diffuser according to the conventional example, and (b) is a flow of the diffuser showing the diffuser according to the first embodiment compared with the diffuser of the conventional example. The whole correlation diagram of the road length and the cross-sectional area ratio, (c) is a diagram showing only the downstream side from the latter half of the radial flow path of the flow path length and the cross-sectional area ratio correlation diagram of the diffuser of (b). The graph which shows the breakdown of the pressure loss in a discharge flow path. A diffuser having a configuration of R / h = 1.18 according to a comparative example and a discharge flow path composed of a circumferential flow path having a relatively large cross-sectional area of the flow path of Vdif / Vd = 2.11 according to the comparative example are provided. The cross-sectional view which shows the flow path shape of a single-stage centrifugal compressor. It is a figure which shows the result of having simulated the loss coefficient at the time of adopting the diffuser and the discharge flow path which concerns on Embodiment 1 and Embodiment 2 and the loss coefficient of the discharge flow path of the comparative example with respect to the mass flow rate. The figure which shows the result of simulating the efficiency at the time of adopting the diffuser and the discharge flow path which concerns on Embodiment 1 and Embodiment 2 and the efficiency of the centrifugal compressor of the comparative example with respect to the mass flow rate. FIG. 5 is a cross-sectional view showing a flow path shape of a single-stage centrifugal compressor provided with a diffuser and a discharge flow path according to the second embodiment of the present invention.

The present invention relates to a centrifugal turbomachine such as a centrifugal compressor that compresses a working fluid, and particularly relates to a diffuser and a discharge flow path thereof.
Hereinafter, the embodiments for carrying out the present invention will be described in detail by exemplifying two embodiments. Although the two illustrated embodiments 1 and 2 will be described by taking a single-stage centrifugal compressor as an example, the present invention can be applied to a multi-stage case and other centrifugal fluid machines having a similar structure. ..

<< Embodiment 1 >>
The diffuser and the discharge flow path of the centrifugal compressor according to the first embodiment of the present invention will be described with reference to FIGS. 1 and 3.
FIG. 1 is a cross-sectional view showing a flow path shape of a single-stage centrifugal compressor E provided with a diffuser 8 and a discharge flow path 6 of the centrifugal compressor according to the first embodiment of the present invention.
FIG. 2 is a schematic view showing a discharge flow path 6 that guides a flow in the circumferential direction downstream of the diffuser 8 of the first embodiment.

In the single-stage centrifugal compressor E of the first embodiment, an impeller 1 having blades 1a formed in a plurality of circumferences is rotatably provided around a rotation center C1. At the center of the impeller 1, a suction port 1s for sucking the working fluid 101 into the single-stage centrifugal compressor E (hereinafter referred to as the centrifugal compressor E) is provided. A diffuser 8 for converting the dynamic pressure of the compressed working fluid 101 into a static pressure is provided on the outer peripheral portion of the impeller 1. The diffuser 8 is configured in the order of the radial flow path 3, the curved flow path 4, and the cylindrical flow path 5 from the upstream. The diffuser 8 is formed so that the flow path is narrower than the upstream flow path.

As shown in FIG. 2, in the discharge flow path 6 downstream of the diffuser 8, a circumferential flow path 9 that guides the flow in the circumferential direction is formed so that the cross-sectional area expands from the upstream to the downstream in a circumferential shape. There is. A straight straight pipe flow path 10 is connected to the outlet end 9o of the circumferential flow path 9. The straight pipe flow path 10 is formed so that the cross-sectional area expands linearly from upstream to downstream. The outlet of the straight pipe flow path 10 is the compressor outlet end 10o.

<Operation of centrifugal compressor E>
In the centrifugal compressor E, as shown in FIG. 1, the impeller 1 rotates around the center of rotation C1, and the working fluid 101 is sucked from the outside of the compressor through the suction port 1s near the center (FIG. 1). Arrow α0). The sucked working fluid 101 is sent outward from the rotation center C1 by centrifugal force by the blade 1a of the rotating impeller 1 and is compressed.

As shown by the arrow α1 in FIG. 1, the compressed working fluid 101 is discharged as a swirling flow in the outer peripheral direction of the impeller 1. The discharged working fluid 101 passes through the flow path of the diffuser 8 having the radial flow path 3, the curved flow path 4, and the cylindrical flow path 5 to the discharge flow path 6 having the circumferential flow path 9 shown in FIG. It is discharged. The working fluid 101 sent to the discharge flow path 6 flows through the circumferential flow path 9 and the straight pipe flow path 10 (arrow α2 in FIG. 2), and is connected to the centrifugal compressor E (compressor outlet end 10o). It is discharged to the pipe h1 (arrow α3 in FIG. 2).

The diffuser 8 has the effect of decelerating the flow velocity of the working fluid 101 discharged in the outer peripheral direction of the impeller 1. Here, when efficiency is more important than the operable flow rate range of the centrifugal compressor E, a plurality of diffuser vanes 2 (see FIG. 1) are provided along the swirling flow direction of the discharged working fluid 101. The deceleration effect is increased by forming a flow path through which the working fluid 101 passes between the plurality of diffuser vanes 2.
On the contrary, when the wide range of the flow rate that can be operated is more important than the efficiency of the centrifugal compressor E, the diffuser vane 2 may not be provided.

In the first embodiment, the efficiency of the centrifugal compressor E can be improved without narrowing the operable flow rate range regardless of the presence or absence of the diffuser vane 2. For example, the first embodiment (the present invention) can be applied to a centrifugal compressor having a diffuser in which the trailing edge of a diffuser vane of a conventional example is bent and extended to a flow path, such as the centrifugal compressor described in Patent Document 1. .. Also in this case, the effect of improving the efficiency of the centrifugal compressor E can be obtained without narrowing the operable flow rate range. That is, a good effect can be obtained in terms of both the flow rate range and the efficiency.

In order to compare the features and effects of the first embodiment with a conventional centrifugal compressor, a diffuser vane 102 having a trailing edge on the inner diameter side from the radial flow path outlet end 103o of the comparative example (conventional technique) shown in FIG. 5 is used. A centrifugal compressor 9E composed of the provided diffuser 108 is used.

<Relationship between the radius R on the inner peripheral side of the curved flow path 4 of the diffuser 8 and the axial flow path width h of the radial flow path 3>
In the diffuser 8 shown in FIG. 1, R / h, which is a value obtained by dividing the inner peripheral side (curvature) radius R of the curved flow path 4 by the axial flow path width h of the radial flow path 3, is 0.2 or more and 0. It is desirable to configure it so that it is 5 or less (0.2 ≦ R / ≦ h0.5). Here, it is assumed that the maximum outer diameter D of the flow path of the diffuser 8 does not change. This is because if the maximum outer diameter D changes, the size of the centrifugal compressor E changes, making it difficult to evaluate the performance.

When R / h is larger than 0.5, the deceleration of Vdif and the shortening of the total length of the flow path of the diffuser 8 become insufficient, so that the pressure loss obtained by the synergistic effect of both the deceleration and the shortening of the total length of the flow path is reduced. The reduction will be small.

On the other hand, when R / h is as small as less than 0.2, the inner peripheral radius R of the curved flow path 4 is relatively small with respect to the axial flow path width h of the radial flow path 3. Therefore, when the flow direction is changed from the radial flow path 3 in the radial direction to the cylindrical flow path 5 in the axial direction in the curved flow path 4, the flow is easily separated on the inner peripheral side because R is too small, and the pressure is increased. There is concern that the loss will increase. Therefore, it is desirable to configure the diffuser 8 so that R / h is 0.2 or more and 0.5 or less (0.2 ≦ R / h ≦ 0.5).

The effect when 0.2 ≦ R / ≦ h0.5 will be described with reference to FIG.
FIG. 3A is a cross-sectional view of the diffuser 8 according to the first embodiment in comparison with the diffuser 108 (FIG. 5) of the comparative example. FIG. 3B is an overall correlation diagram of the flow path length and the cross-sectional area ratio of the diffusers 8 and 108 shown by comparing the diffuser 8 of the first embodiment with the diffuser 108 of the comparative example. FIG. 3C is a diagram showing only the downstream side from the latter half of the radial flow paths 3 and 103 in the correlation diagram of the flow path length and the cross-sectional area ratio of the diffusers 8 and 108 of FIG. 3 (b).

The cross-sectional view of the diffuser 8 composed of the radial flow path 3 and the curved flow path 4 shown in FIG. 3A is in the range of R / h = 1.18 (0.2 ≦ R / h ≦ 0.5). The diffuser 108 having the configuration according to the comparative example (outside) and the diffuser 8 according to the first embodiment having R / h = 0.36 satisfying the relationship of 0.2 ≦ R / h ≦ 0.5 are shown in comparison. There is. As described above, the maximum outer diameter D of the flow path of the first embodiment and the comparative example has the same dimensions.

The horizontal axis of FIG. 3B is (distance from the radial flow path inlet ends 3i and 103i in the central part of the flow path of the first embodiment and the comparative example) / (radial flow path in the central part of the flow path of the comparative example). The distance from the inlet end 103i to the curved flow path outlet end 104o), that is, (the length of the radial flow path 3 and the curved flow path 4 of the first embodiment) and (the radial flow path 103 and the curved flow path 104 of the comparative example). Length) and compare. The vertical axis of FIG. 3B is (the cross-sectional area of the flow path of the flow path distance of the first embodiment and the comparative example) / (the cross-sectional area of the flow paths of the outlet ends 4o and 104o of the curved flow path of the first embodiment and the comparative example). ) Is taken.

FIG. 3B shows the flow path distance from the radial flow path inlet end 3i of the first embodiment shown in FIG. 3A to the curved flow path outlet end 4o and the bending from the radial flow path inlet end 103i of the comparative example. The flow path distance to the flow path outlet end 104o is compared, and the flow path cross-sectional area at each flow path distance is compared.

The horizontal axis of FIG. 3C is (distance from the radial flow path inlet ends 3i and 103i of the flow path center portion of the first embodiment and the comparative example) / (radial flow path of the flow path center portion of the comparative example). The distances from the inlet end 103i to the curved flow path outlet end 104o) are taken to compare the flow path lengths of the radial flow paths 3, 103 and the curved flow paths 4, 104 of the first embodiment and the comparative example. The vertical axis of FIG. 3C is (the cross-sectional area of the flow path from the radial flow path inlet ends 3i and 103i of the first embodiment and the comparative example) / (the curved flow path outlet of the first embodiment and the comparative example). The cross-sectional areas of the ends 4o and 104o) are taken.

FIG. 3C is a diagram of the latter half of the flow path distance from the radial flow path inlet ends 3i and 103i (see FIG. 3A) of the comparative example and the first embodiment to the curved flow path outlet ends 4o and 104o. The comparison in 3 (b) is performed. That is, FIG. 3 (c) is an enlarged view of the latter half of the flow path of FIG. 3 (b).
The correlation diagram of the flow path length and the cross-sectional area ratio of the diffusers 8 and 108 of the first embodiment and the comparative example of FIGS. 3 (b) and 3 (c) is shown by comparing the first embodiment and the comparative example. ..

As shown in FIG. 3C, the cross-sectional area of the flow path at the radial flow path outlet end 3o of the first embodiment and the cross-sectional area of the flow path at the radial flow path outlet end 103o of the comparative example are about 0.92. Is about 0.84, which is about 9% larger in the first embodiment than in the comparative example. Therefore, in the first embodiment, the cross-sectional area enlargement ratio in the radial flow path 3 of the diffuser 8 is increased as compared with the comparative example.

Here, when the flow rate is constant and the density is the same, there is a relationship of A1 × V1 = A2 × V2. A1 and A2 are the cross-sectional areas of the flow path, and V1 and V2 are the flow velocities at the cross-sectional areas A1 and A2, respectively.
Therefore, the swirling flow velocity Vdif of the working fluid 101 when flowing into the curved flow path 4 is smaller in the first embodiment than in the prior art.

When Vdif is small, the friction loss expressed by the following equation (1), which is proportional to the square of the flow velocity, is also small, and the pressure loss due to the friction loss that occurs when the working fluid 101 passes through the curved flow path 4 is , The first embodiment is smaller than the conventional technique.
Pressure loss due to friction loss = 1/2 · (τ ₀ · ρ · v ² · d / Δl) (1)
Here, τ ₀ is the coefficient of friction, ρ is the inlet density, v is the flow velocity, d is the inner diameter of the flow path, and Δl is the length of the flow path.

The correlation diagram of the flow path length and the cross-sectional area ratio of the diffusers 8 and 108 of FIGS. 3 (b) and 3 (c) is shown by comparing the prior art and the first embodiment.
As shown in FIG. 3C, in the distance from the radial flow path inlet ends 3i and 103i to the curved flow path outlet ends 4o and 104o at the center of the flow path, the first embodiment is relative to the prior art. It can be seen that it is shortened by about 7%.

Therefore, in the first embodiment, the total length of the flow path of the diffuser 8 composed of the radial flow path 3 and the curved flow path 4 is shorter than that of the prior art, so that the working fluid 101 occurs when the working fluid 101 passes through the diffuser 8. As shown in Eq. (1), the pressure loss due to the friction loss is smaller in the first embodiment than in the comparative example in proportion to the total length of the flow path.

<Swirl flow flow velocity Vdif at the radial flow path outlet end 3o and the circumferential cross-sectional average flow velocity Vd of the portion 9o where the circumferential cross-sectional area of the discharge flow path 6 is maximum>
In the centrifugal compressor E, at the same time as the diffuser 8, the swirling flow velocity Vdif at the radial flow path outlet end 3o of the diffuser 8 is set, and the volume flow rate at the portion 9o where the circumferential cross section of the discharge flow path 6 is maximum is set in the circumferential direction. Discharge flow so that Vdif / Vd of the value divided by the circumferential cross-sectional average flow velocity (circumferential flow velocity) Vd obtained by dividing by the cross-sectional area (cross-sectional area of the portion 9o) is 1.1 or more and 2.0 or less. The circumferential flow path cross-sectional shape of the road 6 is formed.

Here, the portion 9o having the maximum circumferential cross-sectional area of the discharge flow path 6 is a flow path at the connection portion between the circumferential flow path 9 and the straight pipe flow path 10 in the discharge flow path 6 that guides the flow in the circumferential direction. A portion whose cross-sectional area is larger than the flow path cross-sectional area of the other circumferential flow path 9.

In order to reduce the pressure loss, the shape of the circumferential flow path 9 of the discharge flow path 6 that guides the flow in the circumferential direction is preferably a circular cross section.
Further, in the discharge flow path 6 shown in FIG. 2, a portion having a large flow path cross-sectional area from a portion having a small flow path cross-sectional area at the connection portion 9i between the circumferential flow path 9 and the straight pipe flow path 5 (see FIG. 1). The circumferential flow path portion (9i to 9o) up to 9o may be formed so that the flow path cross-sectional area gradually increases in proportion to the circumferential flow path length or in a correlation such as a quadratic function.

Further, the circumferential flow path length (9i to) from the portion where the flow path cross-sectional area at the connecting portion 9i between the circumferential flow path 9 and the straight pipe flow path 5 is small is the circumferential flow path 9 and the straight pipe flow. In the connection portion 9i with the road 5, for example, the circumferential flow is from 0 to 15% with respect to the circumferential flow path length (9i to 9o) between the portion where the flow path cross-sectional area is small and the portion 9o. The circumference of the discharge flow path 6 so that the area of 1/4 of the flow path cross-sectional area of the portion 9o having a large flow path cross-sectional area at the connection portion between the road 9 and the straight pipe flow path 10 has the same flow path cross-sectional area. A directional flow path shape may be formed.

FIG. 4 is a graph showing the breakdown of the pressure loss in the discharge flow path 6. The horizontal axis of FIG. 4 is the (mass flow rate) / (mass flow rate of the design point) in the discharge flow path 6, and the vertical axis of FIG. 4 is the pressure loss of the discharge flow path 6.
The loss in the discharge flow path 6 includes friction loss and deceleration loss. The friction loss shown by the alternate long and short dash line in FIG. 4 tends to increase as the mass flow rate in the discharge flow path 6 increases. On the other hand, the deceleration loss shown by the broken line in FIG. 4 tends to decrease as the mass flow rate in the discharge flow path 6 increases.

The loss in the discharge flow path 6 is represented by the sum of the friction loss and the deceleration loss. Therefore, when the sum is minimized, the pressure loss in the discharge flow path 6 is minimized. In other words, efficiency is maximized.
Using the above relationship, the value of Vdif / Vd is evaluated. It is evaluated as follows from the relationship between the friction loss and the deceleration loss in FIG.
When Vdif / Vd is as small as less than 1.1, the deceleration loss from Vdif to Vd is not sufficient, so that the deceleration loss decreases, but it occurs on the circumferential flow path surface of the discharge flow path 6 proportional to the square of Vd. The friction loss is increased, and the pressure loss generated in the discharge flow path 6 is increased.

On the other hand, when Vdif / Vd is larger than 2.0, the deceleration from Vdif to Vd between the radial flow path outlet end 3o and the circumferential flow path 9 of the discharge flow path 6 is too large. Although the friction loss is reduced, the deceleration loss due to the sudden decrease in the kinetic energy of the flow is increased, and as a result, the pressure loss is increased. Therefore, the cross-sectional shape of the circumferential flow path 9 of the discharge flow path 6 can be formed so that Vdif / Vd is 1.1 or more and 2.0 or less (1.1 ≦ Vdif / Vd ≦ 2.0). , Suitable for suppressing pressure loss.

Next, the effect of the centrifugal compressor E according to the first embodiment on the centrifugal compressor 9E according to the conventional example will be described with reference to FIGS. 5, 6, and 7.
FIG. 5 shows a diffuser having a configuration of R / h = 1.18 (outside the range of 0.2 ≦ R / h ≦ 0.5) according to the comparative example (conventional example) and Vdif / Vd = 2.11 according to the comparative example. Flow path shape of a single-stage centrifugal compressor 9E provided with a discharge flow path 106 composed of a circumferential flow path having a relatively large flow path cross-sectional area (outside the range of 1.1 ≦ Vdif / Vd ≦ 2.0). It is sectional drawing which shows. The single-stage centrifugal compressor 9E shown in FIG. 5 corresponds to the comparative example shown in FIGS. 6 and 7.

FIG. 6 shows the loss coefficient when the diffusers 8 (see FIG. 1) and 28 (see FIG. 8) and the discharge flow paths 6 and 26 according to the first embodiment and the second embodiment described later are adopted, and a comparative example (see FIG. 5). It is a figure which shows the result of simulating the loss coefficient of the discharge flow path 106 of) with respect to the mass flow rate.
FIG. 7 simulates the efficiency when the diffusers 8 and 28 and the discharge flow paths 6 and 26 according to the first embodiment and the second embodiment described later are adopted and the efficiency of the centrifugal compressor 9E of the comparative example with respect to the mass flow rate. It is a figure which shows the result.

Here, the diffuser 8 having the configuration of R / h = 0.36 satisfying the desirable 0.2 ≦ R / h ≦ 0.5 of the first embodiment and the Vdif satisfying the desirable 1.1 ≦ Vdif / Vd ≦ 2.0 A single-stage centrifugal compressor E (solid line in FIG. 6) provided with a discharge flow path 6 composed of a suitable circumferential flow path 9 having a flow path cross-sectional area of / Vd = 1.19 and a centrifugal compressor 9E of a comparative example. The explanation will be given focusing on the comparison with (the broken line in FIG. 6).

The loss coefficient of the discharge flow paths 6 and 106, which is the vertical axis of FIG. 6, is the total pressure difference from the curved flow path outlet ends 4o and 104o to the outlet end 6o of the discharge flow path 6 (see FIG. 2). It is a value divided by the dynamic pressure at 4o (see FIG. 1) and 104o (see FIG. 5).

In FIG. 6, the loss coefficients of the discharge channels 6 and 106 at the mass flow rate at the design point are 38% smaller in the first embodiment than in the comparative example, and the mass flow rate ratio is 80 to 120 with respect to the mass flow rate at the design point. The first embodiment is smaller than the comparative example in the whole range of%. This is because the Vdif / Vd of the first embodiment is changed from 2.11 of the comparative example (conventional technique) to a suitable 1.19 to suppress deceleration, and the kinetic energy in the circumferential flow path 9 of the discharge flow path 6 is suppressed. It is considered that this is because the deceleration loss due to the sudden decrease in the number of vehicles has decreased (see FIG. 4).

In FIG. 7, the efficiency of the design point at the mass flow rate is improved by 1.5% in the first embodiment (solid line in FIG. 7) as compared with the comparative example (broken line in FIG. 7). Further, the efficiency of the first embodiment is improved as compared with the comparative example in the entire range of the mass flow rate ratio of 80 to 120% from the design point. This is because the flow path length of the diffuser 8 is increased by changing the R / h from 1.18 in the comparative example to a suitable 0.36 (within the range of 0.2 ≦ R / h ≦ 0.5). In addition to shortening by about 7% by reducing R (see FIG. 3C), Vdif also decelerated by about 7% to reduce the pressure loss due to the friction loss of the diffuser 8, which is shown in FIG. It is considered that the loss coefficient of the discharge flow path 6 of the first embodiment is relatively smaller than that of the prior art by 38%.

According to the first embodiment, the discharge flow path 6 is provided with the circumferential flow path 9 having the flow path cross-sectional area of Vdiv / Vd = 1.19 satisfying 1.1 ≦ Vdif / Vd ≦ 2.0. Therefore, the loss coefficient of the discharge flow path 6 is 38% smaller than that of the comparative example at the design point, and is smaller than that of the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the mass flow rate at the design point.

In addition, it has R / h = 0.36 that satisfies 0.2 ≦ R / h ≦ 0.5. Therefore, when the maximum outer diameter D of the flow path of the diffuser 8 is the same as that of the comparative example, R is small and the flow path length of the diffuser 8 is shortened by about 7% (see FIG. 3C). In addition, Vdif is also decelerated by about 7% to reduce the pressure loss associated with the friction loss, and the efficiency is improved as compared with the comparative example.

The curved flow path 4 provided on the downstream side of the diffuser vane 2 is formed so that the R / h is 0.2 or more and 0.5 or less. Therefore, without increasing the maximum outer diameter D of the diffuser 8, the cross-sectional area expansion ratio in the radial flow path 3 is increased, and the total length of the flow path of the diffuser 8 composed of the radial flow path 3 and the curved flow path 4 is increased. Can be shortened.
Therefore, the pressure loss of the diffuser 8 can be reduced while the deceleration effect of the diffuser 8 is appropriately increased by increasing the cross-sectional area expansion ratio.

Based on the above, a simple structure that satisfies 0.2 ≦ R / h ≦ 0.5 or / and 1.1 ≦ Vdif / Vd ≦ 2.0 can improve efficiency without narrowing the operable flow rate range. You can get a type turbo machine.
<< Embodiment 2 >>

FIG. 8 is a cross-sectional view showing the flow path shape of the single-stage centrifugal compressor 2E provided with the diffuser 28 and the discharge flow path 26 according to the second embodiment of the present invention.
The diffuser 28 and the discharge flow path 26 of the centrifugal compressor 2E according to the second embodiment of the present invention will be described below with reference to FIG.

The centrifugal compressor 2E according to the second embodiment has the same basic configuration as that of the first embodiment, but the radius R on the inner peripheral side of the curved flow path 24 in the diffuser 28 is the axial flow path width of the radial flow path 23. The only difference is that R / h, which is the value divided by h, is composed of a diffuser 28 larger than 0.5, which does not satisfy 0.2 ≦ R / h ≦ 0.5.
Since the other configurations are the same, similar components are designated by the same reference numerals, and detailed description thereof will be omitted.

Regarding the effect of the centrifugal compressor 2E according to the second embodiment (see FIG. 8) on the centrifugal compressor 9E (see FIG. 5) according to the comparative example, FIG. 5 of the comparative example and the result of simulating the loss coefficient with respect to the mass flow rate. 6 and FIG. 7 showing the results of simulating the efficiency of the centrifugal compressors 2E and 9E with respect to the mass flow rate will be described below.

The optimum circumference of the flow path cross-sectional area of the diffuser 28 and Vdif / Vd = 1.27 having the configuration of R / h = 1.18 (outside the range of 0.2 ≦ R / h ≦ 0.5) according to the second embodiment. A comparison between a single-stage centrifugal compressor 2E provided with a discharge flow path 6 composed of a directional flow path 9 and a comparative example will be described.

In FIG. 6, the loss coefficient of the discharge flow path at the mass flow rate of the design point is 22% smaller in the second embodiment (dashed line in FIG. 6) than in the comparative example (broken line in FIG. 6), and is at the design point. The second embodiment is smaller than the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the mass. This is because the Vdif / Vd of the second embodiment is changed from 2.11 of the comparative example to 1.27 satisfying the suitable 1.1 ≦ Vdif / Vd ≦ 2.0, so that the circumferential direction of the discharge flow path 6 is changed. It is probable that the deceleration loss due to the sudden decrease in kinetic energy in the flow path 9 (see FIG. 2) became smaller.

In FIG. 7, the efficiency at the mass flow rate of the design point is improved by 0.9% in the second embodiment (dashed line in FIG. 7) as compared with the comparative example (broken line in FIG. 7) with respect to the mass of the design point. The efficiency is improved in the second embodiment as compared with the comparative example in the entire range of the mass flow rate ratio of 80 to 120%.

The difference between the second embodiment and the comparative example is that in the second embodiment, the circumferential flow path in which the flow path cross-sectional area of Vdif / Vd = 1.27 within the range of 1.1 ≦ Vdif / Vd ≦ 2.0 is suitable. While the discharge flow path 6 composed of 9 is adopted, in the comparative example, the flow path cross-sectional area of Vdif / Vd = 2.11 outside the range of 1.1 ≦ Vdif / Vd ≦ 2.0 is adopted. The only point is that the discharge flow path 106 (see FIG. 5), which is composed of a relatively large circumferential flow path, is adopted. Therefore, it is considered that the efficiency improvement of 0.9% is due to the loss coefficient of the discharge flow path 26 of the second embodiment shown in FIG. 6 being relatively smaller by 22% than that of the comparative example.

According to the second embodiment, since Vdif / Vd = 1.27 within the range of 1.1 ≦ Vdif / Vd ≦ 2.0, the loss coefficient of the discharge flow path 6 is improved by 22% as compared with the comparative example. The efficiency is improved as compared with the comparative example in the entire range of the mass flow rate ratio of 80 to 120% with respect to the mass of the design point.

Therefore, a single-stage centrifugal compressor (centrifugal turbomachinery) 2E capable of improving efficiency can be obtained without narrowing the operable flow rate range by a simple structure satisfying 1.1 ≦ Vdif / Vd ≦ 2.0. ..

It should be noted that the above-described first and second embodiments show an example of the present invention, and various specific forms and modified forms are possible within the scope of the claims.

As an example of utilization of the present invention, it can be applied to a centrifugal turbo machine represented by a centrifugal blower (blower) or a centrifugal compressor. In the multi-stage centrifugal turbo machine, the present invention may be applied to the final stage.

1 Impeller 2 Diffuser vanes 3, 23 Radial flow path 3i Radial flow path inlet end 3o Radial flow path outlet end 4 Curved flow path 4o Curved flow path outlet end 5 Cylindrical flow path 6, 26 Discharge flow path 8, 28 Diffuser 9 Circumferential flow path 9o Part where the circumferential cross-sectional area of the discharge flow path is maximum 101 Working fluid E, 2E Centrifugal compressor (centrifugal turbomachinery)
h Axial flow path width of the radial flow path R Radius of curvature on the inner circumference side of the curved flow path (inner circumference side radius)
Vd Circumferential cross-sectional average flow velocity obtained by one-dimensional calculation obtained by dividing the volumetric flow rate in the portion where the circumferential cross-sectional area of the discharge flow path is maximum (the portion where the circumferential cross-sectional area of the discharge flow path is maximum Circumferential cross-sectional average flow velocity in
Vdif Swirling flow velocity at the outlet end of the radial flow path

Claims (4)

It is composed of a radial flow path, a curved flow path that connects the radial flow path and a discharge flow path that guides the flow in the circumferential direction, and a flow path that bends in a direction parallel to the rotation axis direction of the impeller.
When the swirling flow velocity at the outlet end of the radial flow path is Vdif and the circumferential flow velocity at the portion where the circumferential cross-sectional area of the discharge flow path is maximum is Vd.
1.1 ≤ Vdif / Vd ≤ 2.0
A discharge flow path of a centrifugal turbomachine, characterized in that a circumferential flow path of the discharge flow path is formed so as to have a relationship of. It is composed of a radial flow path, a curved flow path that connects the radial flow path and a discharge flow path that guides the flow in the circumferential direction, and a flow path that bends in a direction parallel to the rotation axis direction of the impeller.
When the radius on the inner peripheral side of the curved flow path is R and the axial flow path width of the radial flow path is h,
0.2 ≤ R / h ≤ 0.5
There is a relationship,
When the swirling flow velocity at the outlet end of the radial flow path is Vdif and the circumferential flow velocity at the portion where the circumferential cross-sectional area of the discharge flow path is maximum is Vd.
1.1 ≤ Vdif / Vd ≤ 2.0
A diffuser of a centrifugal turbomachine, characterized in that a circumferential flow path of the discharge flow path is formed so as to have a relationship of the above. A centrifugal turbomachine characterized by comprising the diffuser according to claim 2. A centrifugal turbomachine provided with the discharge flow path according to claim 1.

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