JPS64561B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention provides a multi-stage centrifugal or mixed flow compressor,
It relates to multi-stage turbomachinery such as blowers.

Conventionally, this type of multi-stage turbomachine is usually configured such that the flow after leaving the impeller is guided to the next stage impeller through a diffuser, a U-turn, and a return passage. Among these channel elements, in the differential user, one is used that is composed only of a wall surface without guide vanes, or one that is composed of a large number of guide vanes or guide channels arranged on the circumference. ing.

Figure 1 is a cross-sectional view of a conventional two-stage centrifugal compressor.
The impeller 1 is driven by a shaft 10 and sucks gas through an inlet 7 and discharges it through an outlet of the impeller 1 having an outer diameter D ₁ . After that, the gas passes through the diffuser section 3 without guide vanes, decelerates, and
Enter the U-turn section 4 from the D ₂ part, and guide the guide vane 6.
It is sucked into the next stage impeller 1' through a return flow path 5 having a diameter. 2 is a casing; 11 and 11' are shaft labyrinth seals for preventing gas leakage; 12 is a liner ring for preventing leakage from the impeller; and 14 is a spacer for spacing the blades. In the example in Figure 1,
The diff user 3 has no guide vanes, and the area of the flow path formed by the wall surface of the casing increases as the diameter increases, so the gas is decelerated and compressed. 15 is a shaft nut, and 13 is a vortex chamber for collecting and sending out the final stage gas.

In this way, in the case where the differential user part 3 does not have guide vanes, a fairly long flow path is required to sufficiently decelerate the high-speed fluid flow exiting the impeller 1 and convert it into pressure energy. Impeller 1 outer diameter
A differential user portion 3 having an outer diameter D ₂ approximately 1.6 to 2 times larger than D ₁ is used. The differential user section 3 without guide vanes has the advantage of a simple structure and a wide range of applicable flow rates, but for the above reasons, the outer diameter of the casing 2 of the entire machine that houses the impeller 1 is large. Moreover, it is inefficient.

In contrast, differential users with guide vanes have been used to improve efficiency. This is an attempt to efficiently decelerate the flow and restore pressure by gradually increasing the flow path area between the guide vanes arranged on the circumference.

If this conventional example is shown in FIGS. 2 and 3,
Gas discharged from the impeller 1 is rectified and decelerated by a diff user section 3 having guide vanes 8 .
As shown in Part 3, a large number of guide vanes 8 are arranged on the circumference, a flow path is formed between the vanes, and the guide vanes 8 are usually curved as shown in the figure according to the swirling direction of the flow. It is normal to have this shape. In the case with such a guide vane 8, the effect of deceleration and pressure increase is greater than in the case without a blade, so the outer diameter D ₃ of the guide vane 8
Also, the outer diameter _D2 of the differential user section 3 can be made relatively small. The ratio is generally around 1.5 to 1.7, but the disadvantage is that the operating flow rate range is narrow.

Also, in terms of efficiency, the efficiency is higher near the design point compared to the case without blades, but the guide blades 8
are arranged so that they overlap on the circumference, the flow will be blocked at the smallest area between the blades, and if the blades are fixed, the direction of the flow and the angle of the blades will change. However, it had drawbacks such as being unsuitable for use over a wide flow range because it would not match outside of the design point.

Regarding the U-turn section 4 and the return passage 5, when the flow moves to the next stage after leaving the differential user section 3, the U-turn section 4 directs the flow from the outer circumference of the compressor toward the inner circumference. Make a U-turn. Since this flow is a swirling flow having a velocity component in the circumferential direction, a guide vane 6 is provided in the return flow path 5 in order to rectify the flow into a flow without swirling toward the axial center.

In the conventional type, the flow coming out of the differential user part 3 is a swirling flow with a large circumferential component remaining, so the guide vanes 6 that receive this flow have a fairly long curved part and flow in the radial direction near the inner periphery. The heading straight section will align the flow radially. Therefore, as the length of the guide vane 6 becomes longer, friction loss increases, and the length of the straight portion facing in the radial direction of the total length of the guide vane 6 becomes shorter as a result. The drawback was that the direction rectification effect could not be obtained.

SUMMARY OF THE INVENTION An object of the present invention is to provide a multi-stage turbomachine that has high efficiency, can be operated over a wide flow rate range, and can have a small outer diameter of the casing, while eliminating the above-mentioned drawbacks of the conventional ones. It is something to do.

In order to achieve this objective, the inventors conducted many experimental studies, and based on the knowledge obtained at that time, the present invention was made.

In other words, in the past, the outer diameter of the differential user part
Focusing only on the boosting effect of the differential user, it was thought that a larger D ₂ would be preferable. For example, FIG. 4 is a graph showing the experimental results of the relationship between the outer diameter ratio D ₂ /D ₁ and the pressure recovery rate. In either case, the outer diameter ratio
The larger D ₂ /D ₁ is, the greater the pressure recovery rate and the greater the pressure rise. Conventionally, we focused only on the effect in the differential user part, and looked at the outer diameter ratio.
D ₂ /D ₁ was set relatively large.

However, according to the inventors' research, if the multistage compressor improves the deceleration effect in the differential user,
In fact, it was confirmed that when the outer diameter ratio D ₂ /D ₁ is large, it actually causes loss as a whole and reduces efficiency. The present invention is configured to make the outer diameter ratio D ₂ /D ₁ as small as possible, contrary to the conventional practice, and furthermore, by reducing the outer diameter ratio D ₂ /D ₁ , the problem that will be newly generated will be solved. This was done by solving the problem.

According to the inventors' research, if the length of the diff user is made long for pressure recovery as in the conventional system (as a result, the outer diameter ratio D ₂ /D ₁ is also increased), the flow path of the diff user part 3 will return. It was confirmed that the flow path of the flow path 5 also became longer, and the friction loss in those flow paths increased, and on the contrary, the efficiency decreased.

However, simply reducing D ₂ /D ₁ causes new problems.

That is, if the guide vanes 8 of the differential user section 3 have a conventional structure that curves as the free vortex rotates as shown in FIG. 3, the fluid exiting the guide vanes 8 has a considerably large circumferential velocity component. The inlet of the guide vane 6 of the return passage 5 is also formed at a relatively shallow angle with respect to the circumferential direction to receive the flow. On the other hand, this flow must be diverted in the radial direction within the guide vane 6, but the straight radial portion of the entire length of the guide vane 6 cannot be made long. Therefore, the direction must be changed over a short distance, but this is difficult, and the fluid with a swirling component (circumferential component) flows into the impeller 1' of the next stage, and the required rotation is carried out in the next stage. It becomes difficult to generate pressure.

In order to solve this problem, the inventors conducted further research and found that the camber line of the guide vane has been changed from the conventional method where the angle with respect to the circumferential direction gradually becomes shallower as the radius increases, as shown in Figure 3. This problem was solved by creating a warp in the opposite direction so that it approaches the radial direction as the radius increases. That is, as a result, the guide vanes 6 of the return flow path 5
The direction of the flow entering the inlet of is closer to the radial direction,
Although the overall length of the guide vane 6 is short, the length of the part near the radial direction is on the contrary longer than that of the conventional guide vane, and a flow with almost no swirl component is sent to the suction port of the next stage impeller 1'. This made it possible to prevent damage to the compression characteristics of the next stage.

The above-mentioned FIG. 4 compares the performance only in the differential user section 3, but the present invention aims to improve the performance comprehensively, including flow paths other than the differential user section 3, and FIG. The experimental results are shown below. Figure 5 shows the impeller 1 of the first stage of the compressor.
The flow path from the inlet to the inlet of the next stage, that is, the impeller 1, the diffuser section 3, the U-turn section 4, and the return flow path 5, is shown to represent the overall performance per stage of the compressor. The vertical axis of the figure is the pressure, and the horizontal axis is the dimensionless coefficient that represents the flow rate (pressure coefficient and flow coefficient, respectively).
The three curves in the figure have ratios of the outer diameter D ₂ of the differential user part 3 and the outer diameter D ₁ of the impeller 1 of 1.2, 1.4,
1.6 each case is shown.

According to this experiment, unlike the results shown in Fig. 4 observed at the entrance and exit of the differential user section 3, it was found that rather than making the length of the differential user section 3 sufficiently long to recover the pressure, U-turns were made at a point with a small diameter. It has been found that it is more effective to connect it to section 4 and pass it through return channel 5. The reason for this is that the longer the differential user section 3 becomes, the longer the return flow path 5 becomes, which actually increases friction loss.

According to experiments, D ₂ /D ₁ is considered to be too short for the differential user section 3 compared to the conventional method.
It has been found that the best pressure recovery can be obtained in the entire range from the diffuser section 3 to the return flow path 5 when the pressure is close to 1.2 to 1.45. That is, in FIG. 5, as the value of D ₂ /D ₁ increases, the efficiency decreases significantly on the large flow rate side.

Figure 6 shows this trend in order to explain it in an easy-to-understand manner.
D ₂ /D ₁ is plotted on the horizontal axis, and the flow rate in Figure 5 is plotted on the vertical axis.
This is a comparison of three pressure values at , . In other words, in the case of D ₂ /D ₁ = 1.2, the pressure is high at any flow rate, so let this be considered 1 and other
It is calculated by taking the ratio of the pressure of D ₂ /D ₁ . Looking at this, as the flow rate increases →→,
It is clear that the larger the ratio of D ₂ /D ₁ is, the lower the pressure rise per stroke becomes. In Figure 6, if we allow a decrease in pressure on the vertical axis of up to 2%,
The optimal range of D ₂ /D ₁ can be considered to be 1.2 to 1.45. If the dimension D ₂ is set within this range, the efficiency will be improved compared to the conventional compressor casing 2, and the outer diameter of the compressor casing 2 will be reduced, resulting in a significant reduction in weight.

Furthermore, the problem that the provision of guide vanes narrows the applicable flow range compared to a differential user without guide vanes can be solved by widening the spacing between the guide vanes, for example by making them wider than the blade length. It is.

That is, in the present invention, an airfoil-shaped guide vane is provided in the diff user, and the warp direction of the camber line of this vane is selected and attached in such a manner that the flow is directed in the radial direction, so that after leaving the vane, Since a considerable pressure increase is obtained and the flow direction is turned in the radial direction, if the fluid is immediately guided to the U-turn section and passed through the return flow path without increasing the diameter of the diffuser, the return flow can be reduced. Since the guide vanes can be designed to be short, a highly efficient flow path with low loss can be designed, the size of the casing can be made small, and the flow can be sufficiently aligned in the radial direction at the inner periphery. This produces the effect that the characteristics of the stage impeller are not impaired.

Further, it is preferable that the guide vanes have an airfoil-like cross-sectional shape rather than a plate-like shape.

The reasons for this are: (1) In turbomachinery, the operating flow rate generally varies over a wide range, and when the flow rate changes, the inlet angle of the flow entering the blade changes. In this case, a collision of flows occurs at the front of the blade, and the flow is disturbed at the blade surface, but if an airfoil is used to smooth the flow, it becomes insensitive to the inflow angle, and the inflow angle becomes It has the advantage of flowing smoothly along the wing surface even if the airflow changes.

(2) For the reasons mentioned above, the flow flows smoothly along the blade surface, so losses are minimized. If the guide vanes are thin plate-shaped, flow separation is likely to occur and losses will increase.

(3) Since the flow along the wall is smooth, the so-called turning angle between the inlet and outlet directions of the blades can be made large, which increases the flow deceleration effect.
That's how big the boost is.

(4) The flow between the wings is a flow sandwiched between the dorsal surface and the ventral surface of the wing. In this case, the flow rate through the dorsal surface is high, and the flow rate through the ventral surface hardly changes. As a result, the pressure on the back side where the flow rate is faster is lower, and the flow passing through the ventral side has a higher pressure, so the flow tends to move toward the lower pressure side in the direction of the dotted line. For this reason, in general, a flow with a high velocity may try to separate from the back surface due to inertia and create a vortex at the end of the back surface, but since the flow shown by the dotted line pushes this back, it is difficult to form a vortex. In other words, it has the effect of suppressing peeling.

By creating an airfoil shape like this, the flow flows smoothly due to the effect of lift (a phenomenon in which a significant pressure difference occurs between the ventral surface and the back surface of the wing), and it is possible to provide a large turning angle. Become. As a result, a large deceleration effect can be obtained in a short section,
This effectively increases the pressure accordingly.
Moreover, there is an advantage that the above-mentioned effects of the blades do not change even if the flow rate changes over a wide range. This point provides a fundamentally different effect from simply using a plate-shaped guide vane.

The present invention provides a multi-stage turbomachine having an impeller, a differential user section, a U-turn section, and a return passage.
Guide vanes are arranged at equal intervals on the circumference in the differential user part, and the cross-sectional shape of the guide vanes is warped so that the direction of the tangent to the camber line approaches the radial direction as the radius increases. A multi-stage turbomachine, characterized in that the outer diameter D ₂ of the outlet portion of the differential user portion is D ₂ =(1.2 to 1.45) D ₁ with respect to the outer diameter D ₁ of the impeller. It is.

Embodiments of the present invention will be described using the drawings. 7th
8 and 8 show a two-stage centrifugal compressor.

The guide vanes 9 of the differential user section 3 are manufactured in the shape of an airfoil, and the radius is large in the tangential direction of the camber line 16 so as to turn the swirling direction of the flow after leaving the impeller 1 in the radial direction. The curvature is applied so that it approaches the radial direction.

FIG. 9 shows the shape and positional relationship of the guide vanes 9 configured as shown in FIGS. 7 and 8 in more detail. In this figure, the guide vane 9 has a chord length L (the length obtained by connecting the tip and rear ends of the airfoil with a straight line), and this chord is attached so as to form an angle 8 with the radial direction. on the other hand,
In FIG. 9, the velocity vector of the flow at the diameter D ₁ of the impeller 1 is shown, where u indicates the outer peripheral speed of the impeller 1 and c indicates the absolute velocity of the gas discharged from the outer circumference of the impeller 1. . The angle that c makes with the radial direction is designated α, and δ is configured to be attached to the eye several degrees smaller than α.

With this configuration, when the gas flow that exits the impeller 1 flows along the blade surface, it is influenced by the lifting force, and the gas flow is directed to the radial side rather than the direction of the swirling flow when there are no blades. will be transferred to
It has been confirmed that, together with an increase in the differential user area, there is a very large deceleration and boosting effect. The guide vanes 9 are clearly depicted in FIGS. 8 and 9 in such a way that the tangential direction of the camber line 16 approaches the radial direction as the radius increases, but this effect causes the flow to swirl as described above. The purpose is to divert the flow in the radial direction, and any blade shape that has this effect falls within the scope of the present invention, regardless of the curvature of the camber wire 16.

In Fig. 9, the velocity vector of the gas flow immediately after leaving the guide vane 9 is indicated by v ₁ , and the velocity vector of the flow in this part without the conventional vane is indicated by v _2. . FIG. 10 shows these two flows plotted on rectangular coordinates in the radial direction and the circumferential direction. In this figure, v ₁ and v ₂ indicate the case where the flow rates are exactly the same since the velocity components v ₁ '' and v ₂ '' in the radial direction are both equal v _M . That is, in a flow field with a diameter D ₃ located at the rear end of the guide vane 9, the velocity v ₁ that is turned in a direction closer to the radial direction by the guide vane 9 is higher than the velocity without the vane.
It is clear that the speed is slower than v ₂ . In other words, the pressure increases as the flow rate decreases. Furthermore, what is important in Fig. 10 is the swirling component of the flow, that is, the velocity component in the circumferential direction.
Comparing v ₁ ′ and v ₂ ′, it can be seen that v ₂ still has a large swirling component.

Next, the case where these flows enter the return channel 5 through the U-turn portion 4 will be explained with reference to FIGS. 11a and 11b.

The return flow path 5 is usually provided with a guide vane 17, and the role of this guide vane 17 is to direct the swirling flow having a circumferential velocity component flowing from the front-stage diff user section 3 toward the axial center. The goal is to rectify the flow without swirling. This effect will be explained with reference to FIG. 11. The outer peripheral portion of the guide vane 17 is bent in the direction of the flow so as to smoothly receive the flow. The angle that the tip makes with the radial direction is γ
Then, the larger the circumferential component of the inflowing swirling flow, the larger the angle γ becomes. An example of a case where this γ is large as in the conventional one is the guide vane 1 shown in Fig. 11.
It is 7'.

In general, in the case of a conventional turbomachine with a differential bladeless user, this γ depends on the ratio of D ₂ /D ₁ .
The value of is approximately 55 to 65 degrees, which is difficult to turn the direction of flow toward the center over a short distance;
In the case of the guide vane 17', as shown in the figure, the vane inevitably has a large curved section, and the straight section 18 toward the center is short. This shows that it is difficult for the gas to be sucked into the next stage to flow without any swirling component; in reality, it will be sucked into the impeller 1' with a considerable swirling component remaining. . Both theoretically and experimentally, in centrifugal or mixed-flow turbomachinery, when the impeller sucks in a flow that has a swirl in the direction of rotation, the head decreases and the required pressure cannot be achieved. well known. In order to avoid this, it is a good idea to complete the turning as short as possible and make the straight portion 19 long, like the return guide vane 17 shown in FIG. 11.

In this embodiment, the swirling component of the flow flowing into the return flow path 5 is reduced by bending the guide vanes 9 of the differential user section 3 in the opposite direction, so that the angle γ can be made small. In the case of the guide vane 17, γ is about 45 to 55 degrees, so the straight part 19 of the return guide vane 17 can be made longer than the straight part 18 of the conventional example such as 17', and almost A flow without swirling components can be sent to the second stage impeller 1'. As a result, the head of the second-stage impeller 1' can be secured as designed, and the length of the return passage 5 can also be shortened, so that losses due to friction are also reduced.

FIG. 11a shows a velocity vector representing a pattern in which the flow exiting the differential user section 3 is bent at the U-turn section 4 and enters the return channel 5, and the velocity vector shown by the solid line is the embodiment of the present invention. This is the direction of flow realized by the configuration, and the vector indicated by the dotted line indicates the direction of flow when deceleration is performed using the conventional method. As shown in this figure, in the U-turn section 4, the incoming flow and the outgoing flow form the same angle β with respect to the radial direction.
The entrance angle of the guide vane 6 of the return flow path 5 is set in accordance with this angle β. If this angle β is small, the guide vanes 6 can easily direct the flow in the radial direction. In Fig. 11b, the blade 17 shows the case where β is small, and 17' shows the case where β is large,
In the conventional flow where a large swirling component remains, β
, the guide vanes 17' must be greatly curved in order to direct the flow in the radial direction, which reduces the length of the straight section 18 of the vane that guides in the radial direction. As the length of the flow path increases, it becomes difficult to sufficiently rectify the flow in the radial direction.

On the other hand, according to the configuration of the embodiment of the present invention, the flow is already oriented in the radial direction at the inlet portion of the guide vane 17 than in the conventional configuration, so that the length of the flow path at the curved portion can be shortened. , since the straight portion 19 can be made sufficiently long, the flow guiding effect is high and the flow can flow straight into the impeller 1' of the next stage, without degrading the performance of the next stage.

In other words, in turbomachinery, if the flow at the inlet of the impeller has a swirling component in the direction of rotation, the head decreases and the required pressure cannot be achieved, so this is effective in preventing this. That's why.

Although a centrifugal compressor has been described above, the present invention can also be applied to a mixed flow type compressor, a blower, and the like.

According to the present invention, the outer diameter of the differential user part can be reduced to make the overall size of the turbomachine smaller and lighter, and the efficiency can be improved compared to conventional ones.Moreover, the multistage turbomachine can be applied to a wide flow rate range. can be provided, and can have extremely great practical effects.

[Brief explanation of the drawing]

FIG. 1 is a vertical cross-sectional view of a conventional example, FIG. 2 is a vertical cross-sectional view of another conventional example, and FIG. 3 is a cross-sectional view taken along the line A-A.
Figure 4 is a graph showing the relationship between outer diameter ratio and pressure recovery rate, Figure 5 is a graph showing the relationship between flow coefficient and pressure coefficient, and Figure 6 is a graph showing the relationship between outer diameter ratio and pressure coefficient ratio. Graph, FIG. 7 is a longitudinal sectional view of the embodiment of the present invention, FIG. 8 is a cross-sectional view taken along the line B-B, FIG. 9 is an explanatory diagram of the differential user guide vane, and FIG. 10 is a vector at the outlet of the differential user guide vane. Comparison diagram, 11th
Figure a is a comparison diagram of the flow vectors before and after the U-turn section, the left side is the conventional one, the right side is the embodiment of the present invention, and Figure 11b is a comparison diagram of the guide vanes of the return flow path, the left side is the conventional one, and the right side is the inventive embodiment. An example is shown. 1, 1'... impeller, 2... casing, 3...
... Difference user part, 4 ... U-turn part, 5 ... Return flow path, 6 ... Guide vane, 7 ... Suction port, 8, 9
...Guide vane, 10...Shaft, 11, 11'...Labyrinth seal, 12...Liner ring, 13...
...volute chamber, 14... spacer, 15... shaft nut, 16... camber wire, 17... guide vane, 1
8,19...straight line part.

Claims (1)

[Scope of Claims] 1. In a multi-stage turbomachine having an impeller, a differential user section, a U-turn section, and a return flow path, guide vanes are arranged at equal intervals on the circumference in the differential user section,
The cross-sectional shape of the guide vane is an airfoil shape that is curved so that the direction of the tangent to the camber line approaches the radial direction as the radius increases, and the outer diameter D of the outlet portion of the differential user portion ₂ is a multi-stage turbomachine characterized in that, with respect to the outer diameter _D1 of the impeller, _D2 =(1.2 to 1.45) _D1 .