JP5875429B2 - Multistage centrifugal blower - Google Patents

Description

  The present invention relates to a centrifugal blower that is one of fluid machines, and more particularly to a multistage centrifugal blower having a swirl flow generator.

An example of a centrifugal blower having a conventional swirling flow generator will be described below with reference to FIG.
FIG. 8 is a cross-sectional view of a conventional multistage centrifugal blower on a plane parallel to the impeller rotational axis.

  In FIG. 8, the multistage centrifugal blower having the conventional swirl flow generator is provided with a suction casing 1 on the upstream side of the fluid. On the downstream side of the suction casing 1, an inlet guide vane 2 (hereinafter referred to as IGV2) for controlling the capacity by applying an arbitrary swirling speed to the inflow flow into the first stage impeller 4a (hereinafter referred to as IGV2). Device). A plurality of the IGVs 2 are attached so as to surround the rotary shaft 3, and each vane is provided with a movable mechanism 8 so that the rotation angle can be arbitrarily changed. A plurality of impellers 4a to 4d for boosting the fluid from the casing 1 are mounted in the axial direction of the rotary shaft 3 on the downstream side of the IGV 2.

  Diffusers 5a to 5d for converting the dynamic pressure of the fluid flowing from the impeller outlet to the static pressure are arranged radially outward of the impellers 4a to 4d, and the next stage impeller inlet is disposed downstream of the diffusers 5a to 5d. Return channels 6a to 6c for guiding fluid to the outlet and a discharge casing 7 are provided. 9 indicates the impeller upstream flow path width.

Here, performance characteristics of the multistage centrifugal fan will be described with reference to FIGS.
FIG. 9 is a diagram showing speed triangles at each position from the IGV outlet to the first stage diffuser outlet, and FIG. 9A is a diagram showing the speed triangles when no pre-turn is given to the first stage impeller. FIG. 9B is a diagram showing a speed triangle when a pre-turn is given to the first stage impeller.

FIG. 10 is a diagram showing a full-stage pressure characteristic curve and a full-stage efficiency characteristic curve of a general multistage centrifugal fan with an IGV.
Note that the two-stage pressure characteristic curve has two characteristic curves with different IGV rotation angles, and the all-stage efficiency characteristic curve has 1 of the case where the swirl speed is not given to the first stage impeller inflow without changing the IGV mounting angle. Only the characteristic curves of the books are shown.

First, the performance characteristics when the turning speed is not given to the first stage impeller 4a (when IGV angle α _IGV = 0 °) will be described with reference to FIG.
In Fig.9 (a), the flow which passed the suction casing 1 and IGV2 flows in into the first stage impeller 4a in the state without a turning speed component. Inside the first stage impeller 4a, a pressure increase (theoretical head H _th ) corresponding to the impeller shaft power represented by (Equation 1) is given. Here, U represents the impeller circumferential speed, Cu represents the circumferential component of the impeller absolute speed, and g represents the gravitational acceleration. Subscript 1 represents the impeller inlet and 2 represents the impeller outlet.
H _th = (U ₂ · Cu ₂ -U ₁ · Cu ₁ ) / g (Formula 1)
When the pre-turn is not given, the second term U _{1 ·} Cu ₁ on the right side becomes 0. From (Equation 1), it is derived that the theoretical head basically increases linearly with a decrease in flow rate (decrease in impeller exit meridional surface velocity Cm ₂ ).

On the other hand, at each flow point, a value obtained by subtracting the total pressure loss of each part of the impeller, diffuser, and return channel from H _th is the actual first-stage pressure increase. Here, the magnitude of the total pressure loss at each flow point is the values of the incident angles i ₁ and i ₃ and the reduction ratios W ₁ / W ₂ and C ₃ / C in the impeller and diffuser shown in FIG. _It varies depending on a value such as ₄ .

The incident angle gradually increases as the flow rate decreases. At a flow rate where the incident angle is near 0, the total pressure loss is the smallest. As the incident angle deviates from 0, the difference between the blade inlet angle β _1b or β _3b and the inflow angle β ₁ , α ₃ increases, the loss increases, and finally stalls.

  On the other hand, the reduction ratio gradually increases as the flow rate decreases. If the reduction ratio is large, the loss is basically reduced by reducing the flow path wall friction. However, if the reduction ratio becomes excessive as the flow rate decreases, the thickness of the flow path wall boundary layer increases, resulting in an increase in loss and eventually a stall.

  From the above, at the first stage, the loss is minimum at the certain flow point (maximum efficiency point), and the efficiency is lowered at both the large flow rate side and the small flow rate side. Particularly on the small flow rate side, as described above, the loss increases as the flow rate decreases, and the gradient of the pressure characteristic curve gradually becomes flat, and finally a rising slope portion appears.

The multistage centrifugal blower has the same pressure characteristic curve and efficiency characteristic curve in the second and subsequent stages. The sum of the characteristics of each stage is the entire stage characteristic. Therefore, even in the entire stage characteristics, the maximum efficiency is obtained at a certain flow rate, the gradient of the pressure characteristic curve gradually becomes flat on the low flow rate side, and finally a rising slope portion appears. It is known that the upward slope portion of the pressure characteristic curve is a surging generation condition that causes self-excited vibration with respect to minute flow rate fluctuations and the entire blower cannot be operated by repeating forward flow and reverse flow. Therefore, this flow point becomes the operation limit of the multistage centrifugal fan when α _IGV = 0 ° (see the characteristic curve of α _IGV = 0 ° in FIG. 10).

Next, performance characteristics when a turning speed is given to the first stage impeller (when IGV angle α _IGV ≠ 0 °) will be described with reference to FIG. 9B.
In FIG. 9B, the flow that has passed through the suction casing and the IGV flows into the first stage impeller in a state where a turning speed component corresponding to the IGV angle is given. At this time, since the theoretical head of the first stage impeller is decreased as the second term U ₁ Cu ₁ ≠ 0 on the right side of the equation (1), the first stage pressure rise is decreased. Due to the decrease in the first-stage pressure increase, the inlet pressure of the second and subsequent stages also decreases, the inlet volume flow rate of the second and subsequent stages increases, and the pressure increase in these stages also decreases. Therefore, the whole-stage pressure rise is also reduced.

As shown in FIG. 9B, when α _IGV ≠ 0 °, the incident angle i ₁ of the first stage impeller is recovered and the reduction ratio W ₁ / W ₂ is also reduced. Accordingly, the stall flow point of the first stage impeller moves to the small flow rate side. At this time, the volume flow rate at the inlet of each element increases as described above in any impeller, diffuser, and return channel located downstream from the first stage impeller. Moving. As a result of the above effects, the flow rate point where the upward slope portion appears in the all-stage pressure characteristic curve moves to the small flow rate side as compared with the case of α _IGV = 0 °. Therefore, at this time, the surging occurrence point also moves to the small flow rate side.

When operating the actual product, the operating flow rate is changed while maintaining the specified pressure while changing the IGV angle. Therefore, the wider the minimum operating flow rate that satisfies the specified pressure, the wider the operating flow rate can be accommodated. The range from the minimum operating limit flow point to the specified flow point is the operating range of the blower (see the characteristic curve of α _IGV ≠ 0 ° in FIG. 10).

  In a multistage centrifugal blower, improvement of specification point efficiency and expansion of the operating range are generally technical issues. Usually, in turbomachinery, there is a trade-off between improving specification point efficiency and expanding the operating range. This is because, as described above, it is effective to increase the specification point incidence angle of each element to 0 ° or increase the specification point reduction ratio in order to improve the specification point efficiency, but at the same time, it will accelerate the stall. is there.

In order to expand the operating range while maintaining specification point efficiency, various operating range expansion devices have been proposed. The above IGV is one of them, and there is an example described in Patent Document 1 below, for example.
In addition, in the open type impeller, there is an example in which an annular bleed flow path is provided on the inducer portion shroud side and the impeller stall flow rate point is moved to the small flow rate side as in Patent Document 2 below. .
Furthermore, as in Patent Document 3 below, there is an example in which the diffuser blade is made variable so as to improve the specification point efficiency and the operation range.

JP 2002-5092 A JP-A-11-201094 JP 63-36098 A

  In a multistage centrifugal blower with IGV, the required specifications for the operating range on the small flow rate side are often strict (for example, a multistage centrifugal blower for sewage aeration). In that case, it is ideal to install an IGV for each impeller in multiple stages. However, this leads to a complicated structure and a significant increase in cost.

  Accordingly, it is necessary to satisfy a sufficiently small flow rate operating range with as few as possible IGVs. For this reason, conventionally, the specification point incidence angles and reduction ratios of the impellers of all stages are set so that the small flow rate side characteristics are improved, and there is a problem that it is difficult to improve the specification point efficiency.

  An object of the present invention is to achieve both improvement in specification point efficiency and expansion of the operating range on the small flow rate side with a simple configuration as much as possible in consideration of the characteristics of each stage performance characteristic curve in a multistage centrifugal fan and the conditions when entering surging. An object of the present invention is to provide a multi-stage centrifugal blower capable of performing the above-described operation.

  In order to achieve the above object, the present invention provides a multistage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to an impeller immediately before an impeller of an arbitrary stage, and the Only the incident angle and the reduction ratio of the specification point of the impeller are set smaller than the impellers of the other stages.

  In order to achieve the above object, in the present invention, it is preferable that only the blade inlet angle of the stage impeller immediately after the swirl flow generator is set smaller than the blade inlet angle of the other stage impeller.

  In order to achieve the above object, the present invention is preferably set such that only the blade outlet angle of the impeller at the stage immediately after the swirl flow generator is smaller than the blade outlet angle of the impeller at the other stage.

In order to achieve the above object, the present invention is preferably configured such that the value of the rate of change WR of the ratio b ₂ / D ₂ between the impeller diameter D ₂ and the blade outlet flow path height b ₂ in two successive stages is Only when the stage immediately after the flow generator is included, the average WR of the other stages differs by 3% or more, and when the stage immediately after the swirl generator is upstream, the WR is maximum and swirl flow is generated. It is desirable to minimize the WR when the stage immediately after the apparatus is on the downstream side.

  According to the present invention, the characteristics of each stage performance characteristic curve in a multistage centrifugal fan and the conditions when entering surging are taken into consideration, and both improvement in specification point efficiency and expansion of the small flow side operating range are achieved with a simple configuration as much as possible. It is possible to provide a multistage centrifugal blower that can perform the above-described operation.

It is a figure which shows the impeller incident angle and the entrance / exit relative speed of the stage immediately after the pre-turn generator which concerns on Example 1 of this invention, and the other stage. It is a figure which shows the flow volume change of the static pressure rise of each stage when there is no turning speed in the impeller inflow flow of the stage immediately after the conventional turning flow generator, and when turning speed is provided. It is a figure which shows the all-stage pressure characteristic curve of the multistage centrifugal fan which concerns on Example 1 of this invention, and a all-stage efficiency characteristic curve. It is the figure which looked at the front edge vicinity of the impeller blade which concerns on Example 2 of this invention from the impeller rotating shaft extension line. It is the figure which looked at the rear edge vicinity of the impeller blade which concerns on Example 3 of this invention from the direction parallel to an impeller rotating shaft. It is the figure which drew the shape of the impeller meridian of two continuous steps according to Example 4 of the present invention. In a conventional multistage centrifugal blower, WR (continuous ratio b ₂ / D ₂ of impeller diameter D ₂ and blade outlet flow path height b ₂ ) when aiming to make the internal flow of all stages imp similar It is a figure showing the ratio of each WR stage value to the overall stage average value. It is sectional drawing which shows schematic structure of the conventional multistage centrifugal blower. It is a figure which shows the speed triangle when not giving a pre-turn with respect to a first stage impeller. It is a figure which shows the speed triangle at the time of giving pre-turn with respect to the first stage impeller. It is a whole-stage pressure characteristic curve of a general multistage centrifugal blower with a swirl flow generator, and a whole-stage efficiency characteristic curve.

  Hereinafter, an embodiment of the present invention will be described with reference to the drawings.

  The centrifugal blower of the present embodiment is a multi-stage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to an impeller immediately before an impeller at an arbitrary stage, and is basically a conventional example shown in FIG. Has similar components. However, as shown in FIG. 1, only the incident angle and reduction ratio of the stage impeller immediately after the IGV, which is the swirling flow generator, are set smaller than the other stage impellers. Different.

The effect of this configuration will be described with reference to FIG.
FIG. 2 shows a conventional multi-stage centrifugal blower with IGV when α _IGV = 0 ° with no turning speed in the impeller inflow immediately after _IGV , and when α _IGV ≠ 0 ° with turning speed _added. The flow rate change of the static pressure rise of a stage is shown. In this example, the number of stages is eight, and an IGV is provided immediately before the first stage.

In FIG. 2, attention is first focused on α _IGV = 0 ° without a turning speed.
In this case, the gradient of the flow rate change due to the increase in static pressure at each stage maintains a lower right side on the large flow rate side, as in FIG. 10 described above. However, it gradually becomes gentle as the flow rate decreases, and is almost zero near the surging occurrence point.

Next, attention is paid to the gradient of the flow rate change of the static pressure increase at each stage when α _IGV ≠ 0 ° with the turning speed.
In this case, only the stage immediately after the IGV (first stage) shows a steep downward slope. The gradient at the first stage is suddenly gentle near the surging occurrence point and finally becomes almost zero at the surging occurrence point. On the other hand, except for the first stage, the gradient is already 0 at the maximum flow rate point measured at this IGV angle, and it can be seen that the gradient remains almost 0 even when the flow rate is reduced.

From the above, when α _IGV ≠ 0 °, the components other than the first stage impeller that is not directly given the swirl speed to the inflow flow are always stalled, and the pressure that is a condition for the stable operation of the blower There is no ability to maintain the downward slope of the characteristic. The slopes of the pressure characteristic curves of these stages are almost flat with respect to the flow rate change. However, since the first stage impeller which is given a turning speed and maintains a steep right-down slope, the pressure characteristic curve of all stages maintains a right-down slope, so that the entire blower can be operated without surging. When the flow rate is further reduced and the first stage impeller finally stalls, the slope of the pressure characteristic curve of all stages becomes flat, and it can be seen that surging occurs.

From this result, in the multistage centrifugal blower, if α _IGV ≠ 0 °, if only the impeller of the stage immediately after the IGV can move the stall generation flow point to the small air volume side, the stall of the other stages will occur on the large flow side. Even so, the operating range of the entire blower can be expanded.

  In the multistage centrifugal blower having the shape of each stage impeller shown in FIG. 1, since the incident angle and the reduction ratio of the stage impeller immediately after the IGV are set small, the surging flow rate of the entire blower is reduced to the small flow rate side. You can move. However, this reduces the efficiency of the specification point at the first stage.

  On the other hand, at the other stages, the impeller incident angle and reduction ratio are set larger than the first stage, and the efficiency of the specification point is improved. Since the overall efficiency of the multistage centrifugal fan is an average value of the efficiency of each stage, if the sum of the increased efficiency of other stages is larger than the decreased efficiency of the first stage, the overall efficiency of the specification point is improved. That is, in the multistage centrifugal blower having the shape of each stage impeller shown in FIG. 1, it is possible to achieve both the improvement of the entire stage efficiency at the specification point and the expansion of the operating range.

  The specification point incidence angle and reduction ratio of the first stage impeller are set according to the target operating range. On the other hand, the specification point incidence angle and reduction ratio values of the impellers of the other stages are at least larger than the initial stage values and the maximum efficiency (0 for the incidence angle. For the reduction ratio, the friction loss and It is an optimum value obtained from the relationship of deceleration loss, and it is a range smaller than (which can be derived by one-dimensional simple calculation).

FIG. 3 is a diagram showing a comparison of performance prediction results of the design example of the conventional multistage centrifugal fan and the design example of the multistage centrifugal fan described in the first embodiment.
In this embodiment, the impeller incident angle is changed between the stage immediately after the IGV and the other stages.

  In FIG. 3, in the design example described in the first embodiment, the specification point efficiency is improved as compared with the conventional design example. On the other hand, in the design example described in Example 1, the pressure increase on the small flow rate side is smaller than that in the conventional design example, but the pressure exceeding the specification is satisfied and there is no problem in operation. Furthermore, at this time, the surging generation flow point when the turning speed is given to the impeller inflow immediately after the IGV moves to the smaller flow rate side as compared with the conventional design example. Therefore, as a result, the operating range is equal to or greater than that of the conventional one.

  In this embodiment, the number of stages of the centrifugal fan is not limited as long as it is a plurality of stages. However, the larger the number of stages, the more advantageous the specification point efficiency can be increased. In addition, IGVs may be installed upstream of any stage of impellers, up to several stages, but it is simplest to install one IGV upstream of the first stage.

Furthermore, when α _IGV ≠ 0 °, it is very important in this embodiment that the gradient of the pressure characteristic curve at a stage other than immediately after the IGV becomes substantially flat with respect to the flow rate change. In a centrifugal compressor having a pressure ratio higher than that of a centrifugal fan, the slope of the pressure characteristic curve after stalling often increases sharply to the right. Therefore, the present invention is particularly effective in a multistage centrifugal fan.

Below, another Example of the multistage centrifugal blower in this invention is shown.
FIG. 4 is a view of the vicinity of the leading edge of the impeller blade as viewed from the impeller rotation axis extension line.

In FIG. 4, the centrifugal blower of the present embodiment is a multi-stage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to the impeller immediately before the impeller at an arbitrary stage, and the specification point incident angle and deceleration It has the same characteristics as the first embodiment with respect to the ratio, and as shown in FIG. 4, only the blade inlet angle of the impeller at the stage immediately after the swirl flow generator is from the blade inlet angle of the impeller at the other stage. Is also characterized by a small setting. In this case, the relative flow angle beta ₁ in the impeller inlet shown in Figure 9 is set to be equal in each stage.

By the structure described in this embodiment, the smaller the specification points incidence angle i ₁ of the impeller of the stage immediately after IGV, surging occurrence rate of the entire blower that moves can to small flow rate side. While that in the other stages, by setting the value close as possible to zero and the i ₁ of the impeller larger than the first stage, it is possible to improve all stages efficiency in design point.

In order to move the surging generation flow rate of the entire blower to the low flow rate side, in order to reduce i ₁ at the stage immediately after the IGV, as shown in FIG. 9, the impeller inlet relative flow angle β ₁ is increased, or The impeller blade inlet angle β _1b needs to be reduced.

Here, as shown in FIG. 9, β ₁ is determined by the impeller inlet peripheral speed U ₁ and the impeller inlet absolute speed C ₁ . When α _IGV = 0 °, β ₁ is determined by U ₁ and meridional surface direction component Cm _{1 of the} impeller inlet absolute velocity. U ₁ and Cm ₁ are expressed by the following (Expression 2) and (Expression 3) if the blockage due to the blade thickness at the leading edge of the impeller is ignored. Β ₁ is expressed by the following (formula 4). Here, R ₁ is the impeller blade leading edge radius, ω is the impeller rotational angular velocity, Q ₁ is the volume flow rate at the impeller inlet, and b ₁ is the flow channel height at the impeller inlet.
U ₁ = R ₁ · ω (Formula 2)
Cm ₁ = Q ₁ / (2 · π · R ₁ · b ₁ ) (Equation 3)
β ₁ = tan ⁻¹ (Cm ₁ / U ₁ ) (Formula 4)
From equation (2) to (Equation 4), in order to increase the beta ₁ to reduce the i _1, reduce the U _1, may be increased Cm _1.

In order to reduce U ₁ without changing the rotation speed, it is necessary to reduce the impeller leading edge radius R ₁ . However, since the rotational axis diameter can not be changed, in order to reduce the only R _1, loss narrowed extremely impeller upstream passage width 9 shown in FIG. 8 is increased. On the other hand, when the result in increased Cm _1, basically causes increased flow velocity at the impeller inlet, friction loss increases. Therefore, both when U ₁ is reduced and Cm ₁ is increased, the loss further increases in addition to the loss increase due to i ₁ reduction, and the efficiency of the stage immediately after the IGV is unnecessarily reduced.

For the above reasons, β _1b should be reduced in order to reduce i ₁ immediately after the IGV. Therefore, in this embodiment, i ₁ is adjusted by changing β _1b between the stage immediately after the IGV and the other stages.

Below, another Example of the multistage centrifugal blower in this invention is shown.
FIG. 5 is a view of the vicinity of the trailing edge of the impeller blades as seen from a direction parallel to the impeller rotation axis.

  In FIG. 5, the centrifugal blower of the present embodiment is a multi-stage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to the impeller immediately before an impeller at an arbitrary stage. This multi-stage centrifugal blower has the same characteristics as the first embodiment with respect to the specification point incident angle and the reduction ratio, and further, as shown in FIG. 5, the blade outlet of the impeller at the stage immediately after the swirl flow generator. Only the corner is characterized by being set smaller than the blade exit angle of the impeller at the other stage.

By the structure described in this example, changing the impeller exit velocity triangles in FIG outlet relative velocity W ₂ of the impeller stages immediately after IGV becomes larger, resulting IGV impeller stage immediately after Since the reduction ratio W ₁ / W ₂ becomes small, the surging flow rate of the entire blower can be moved to the small flow rate side. While that in the other stages, W _{1 /} W ₂ of the By larger than the first stage, it is possible to improve all stages efficiency in design point.

Multistage The centrifugal blower, basically by performing the design was the same all the impeller outer diameter D ₂ of each stage can minimize the overall diameter of the blower. This is when to satisfy the specifications pressure at all stages and thus by changing the D ₂ for each stage, than the impeller outer diameter in the case of the D ₂ and all stages the same, smaller outer diameter and a large stage impeller This is because a step occurs. The outer diameter of the entire blower, would be governed by the size of the step having the maximum D ₂ of? Therefore, it is effective to design all stages of D ₂ as the same, which is synonymous with the respective stage pressure increase in all stages the same.

In order to change W ₁ / W ₂ between the stage immediately after the IGV and the impeller of the other stages under the condition that D _{2 of} all stages is the same, as shown in FIG. 9, the impeller blade outlet angle β It is necessary to change _{2b or} to change the blade outlet flow path height b ₂ (the meridional surface direction component Cm _{2 of the} impeller outlet absolute speed). This embodiment corresponds to a shape in which β _2b is changed between the stage immediately after the IGV and the impeller at the other stages.

Below, another Example of the multistage centrifugal blower in this invention is shown.
FIG. 6 is a diagram in which the shape of the impeller meridian of two successive stages is overlaid in the multistage centrifugal blower of the present embodiment.
FIG. 6A shows a case where two successive stages do not include the stage immediately after the IGV.
On the other hand, FIG. 6B shows a case where two consecutive stages include a stage immediately after the IGV.

In FIG. 6, the centrifugal blower of the present embodiment is a multistage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to the impeller immediately before the impeller at an arbitrary stage. This multistage centrifugal blower has the same characteristics as the first embodiment with respect to the specification point incident angle and the reduction ratio. Furthermore, as shown in FIG. 6, the change rate WR of the ratio b ₂ / D ₂ of the impeller diameter D ₂ and the blade outlet flow path height b ₂ in two successive stages is the stage immediately after the swirl flow generator. Only when the WR is 3% or more different from the average value of the WR of the other stages and the stage immediately after the swirl flow generator is upstream, the stage immediately after the swirl flow generator is downstream It is characterized by minimum WR when it is on the side.

By adopting the configuration described in the present embodiment, the reduction ratio W ₁ / W ₂ of the impeller immediately after the IGV is reduced, and the surging flow rate of the entire blower can be moved to the small flow rate side. While that in the other stages, W _{1 /} W ₂ of the By larger than the first stage, it is possible to improve all stages efficiency in design point.

  In the multistage centrifugal blower, the working fluid is gradually increased in pressure at each stage, so that the fluid density at each stage inlet gradually increases and the volume flow rate at each stage inlet gradually decreases. At this time, from the first stage to the last stage, if the flow path height is gradually reduced by the volume flow ratio between two successive stage inlets and the blade angle distributions are matched, the internal flow of all stages of the impeller is similar. Can achieve the same performance in all stages.

In Example 3, the multistage centrifugal blower basically the same all the impeller outer diameter D ₂ of each stage, said making the design of each stage pressure increase in all stages the same is valid. Therefore, the density ratio between the inlets of two consecutive stages (the first stage to the second stage, the second stage to the third stage, ... (N-1) stage to the N stage, where N is the total number of stages), that is, the volume flow rate ratio. Are equal in all of these. Accordingly, when the flow inside the impellers of all stages is similar, the ratio b ₂ / D ₂ of the impeller diameter D ₂ and the blade outlet flow path height b ₂ shown in the following (formula 5) is 2 The rate of change WR between the two stages is also substantially equal, less than 1 in all of these.
_{WR = / (b 2 / D} 2 at k-th stage) _{((b 2} / _D 2 at k-1) th stage) <1 (Equation 5)
In actual products, there are problems such as manufacturing accuracy, and the internal flow of impellers in all stages cannot be made completely similar.

FIG. 7 is a diagram showing the ratio of the WR of each stage with respect to the average value of all stages WR when aiming to make the flow inside the impellers of all stages similar in a conventional multistage centrifugal blower.
In FIG. 7, as can be seen from this example, the WR of each stage is shifted by about 3% at maximum with respect to the WR average value. On the other hand, in this embodiment, only when the value of WR includes the stage immediately after the swirl flow generator, the average value of WR of the other stages differs by 3% or more and the stage immediately after the IGV is on the upstream side. The WR is the maximum in all stages, and the WR when the stage immediately after the IGV is downstream is the minimum in all stages.

This means that the blade outlet channel area with respect to the volume flow rate is narrower in the stage immediately after the IGV than in the other stages. If this 9, stage impeller Deguchiko meridional direction velocity component Cm ₂ immediately after IGV is greater than the impeller of the other stages. Therefore, the impeller outlet relative velocity W ₂ is reduced in stages after IGV, resulting IGV speed reduction ratio W _{1 /} W ₂ of the impeller stages immediately after, it becomes smaller than the other stages.

In order to change W ₁ / W ₂ between the stage immediately after the IGV and the impeller of the other stages under the condition that D _{2 of} all stages is the same, as shown in FIG. 9, the impeller blade outlet angle β It alters the _2b, vane outlet channel width b _{2 (impeller} outlet absolute velocity of the meridional direction component Cm ₂₎ needs to be changed. In the present embodiment these, corresponding vane outlet width b ₂ in a shape is relatively changed by the impeller of the stage and the other stage after IGV.

  As described above, according to the present invention, in a multistage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to an impeller immediately before an impeller at an arbitrary stage, the specifications of the multistage centrifugal blower are as simple as possible. It is possible to provide means that achieves both improved point efficiency and expansion of the small flow rate side operating range.

DESCRIPTION OF SYMBOLS 1 ... Suction casing, 2 ... Swirling flow generator, 3 ... Impeller rotating shaft, 4a, 4b, 4c, 4d ... Impeller, 5a, 5b, 5c, 5d ... Diffuser, 6a, 6b, 6c, 6d ... Return channel , 7 ... Discharge casing, 8 ... Movable mechanism of swirl flow generator, 9 ... Impeller upstream flow path width.
<Explanation about symbols>
C: Absolute velocity, Cm: Absolute velocity meridional component, Cu: Absolute velocity circumferential component, D ... Diameter, DRi ... Impeller reduction ratio, DRi ... Diffuser reduction ratio, Hth ... Theoretical head, Hth ... Theoretical head , Q: Volume flow rate, R: Radius, U: Impeller peripheral speed, W: Relative speed, WR: Ratio of change between impeller diameter and blade outlet flow path height between two consecutive stages, b ... impeller flow path height, g ... gravitational acceleration, i ... incident angle, alpha ... absolute flow angle, alpha _IGV ... swirling flow generator rotation angle, beta ... relative flow angle, beta _b ... wheel vane angle, omega ... impeller Rotational angular velocity.
<Explanation about subscripts for symbols>
DESCRIPTION OF SYMBOLS 1 ... Impeller entrance, 2 ... Impeller exit, 3 ... Diffuser entrance, 4 ... Diffuser exit.

Claims (4)

In a multistage centrifugal blower having at least one swirling flow generating device that imparts a swirling flow to an impeller immediately before an impeller at an arbitrary stage,
A multistage centrifugal blower characterized in that only the specification point incident angle and / or the reduction ratio of the stage impeller immediately after the swirl flow generator are set smaller than those of the other stage impellers. In the multistage centrifugal blower according to claim 1,
Only the blade inlet angle of the stage impeller immediately after the swirl flow generator is set smaller than the blade inlet angle of the other stage impeller. In the multistage centrifugal blower according to claim 1,
Only the blade outlet angle of the stage impeller immediately after the swirl flow generator is set smaller than the blade outlet angle of the other stage impeller. In the multistage centrifugal blower according to claim 1,
Only when the change rate WR of the ratio b ₂ / D ₂ of the impeller diameter D ₂ and the blade outlet flow path height b ₂ in two successive stages includes the stage immediately after the swirl flow generator. The WR is 3% or more different from the average value of the WR of the stage, and the WR is the maximum when the stage immediately after the swirling flow generator is on the upstream side, and the WR when the stage immediately after the swirling flow generator is on the downstream side. Multistage centrifugal blower characterized by being minimized.

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