JPS63205494A - Biaxial reversal centrifugal type fluid booster - Google Patents

Abstract

PURPOSE:To make such a high efficiency as approximating to an axial-flow system giveable, by using a vaned rotary diffuser (a second impeller) in place of a stator diffuser, and constituting this diffuser so as to be rotated at proper speed in the opposite direction to the convertional impeller (a first impeller). CONSTITUTION:A high speed fluid to be discharged out of the outer circumference of a first impeller 3 is led into a second impeller 3 rotating in the opposite direction to the impeller 3 at once or after being decelerated and boosted properly by a vortex chamber 5, or if necessary, a vaned fixed diffuser. And, a large relative speed is produced at an inlet of the impeller 6, while an absolute speed of the fluid at an outlet of the impeller 6 is sharply reduced by deceleratization due to a blade spread of the impeller 6 and that due to rotation of itself. With this constitution, dynamic pressure is efficiently converted into static pressure, and it is led into a volute casing 9 on the circumference of the impeller 6 at low speed, whereby a fluid friction loss inside this casing 9 is reduced, thus a large pressure rise of the fluid is securable at a higher efficiency as a whole.

Description

[Detailed description of the invention] (Industrial field) The invention makes it possible to improve the efficiency of centrifugal type (mixed flow type is a type of centrifugal type, so mixed flow type is also included) turbo blower or turbine pump. The present invention relates to a biaxial reversing centrifugal fluid booster.

(Prior art and its problems) The centrifugal type is inherently less efficient than the oil flow type, but
It is widely used in applications where the comparative rotation angle N8 is in a small range, ie, relatively high pressure and small capacity. For convenience of explanation, the feeder and compressor will be explained first.

The main reason why the efficiency of centrifugal type engines is said to be inferior to that of axial flow types is due to the low efficiency of the diffuser, which converts the high dynamic pressure provided by the blades into static pressure. Rather,
The efficiency of the blade wheel itself is not bad at all, but it depends on the design.9
The present inventor has experienced that it is possible to achieve 0chi or more.

Despite this, the overall efficiency is usually 65-75 for low N8.
%. Some large-capacity, high-Ns models have achieved an overall efficiency of 85 cm, but large-capacity axial flow models have achieved a total efficiency of 90 cm.
It's still not as good as ~95chi.

(Object of the present invention) The object of the present invention is to provide a biaxially reversing centrifugal type fluid booster that can achieve high efficiency close to that of an axial flow type, while eliminating the drawbacks of the conventional centrifugal type turbomachinery described above. The principle of the invention is to use a winged rotating diffuser (referred to as the second wheel) instead of a stationary diffuser, and to rotate this diffuser in the opposite direction to the conventional blade wheel (referred to as the first wheel). By rotating at a suitable speed, the short blades rather than a fixed diffuser efficiently convert most of the dynamic pressure imparted by the first wheel into static pressure, reducing the absolute velocity of the fluid at the exit of the second wheel. 1 in an outlet spiral casing or a casing with an appropriate space by reducing K appropriately.
11 The purpose is to reduce friction loss (and provide high efficiency close to that of the axial flow type even for low NsK).

Therefore, for high N s K, it also provides high efficiency comparable to that of the axial flow type.

(Composition of Akira Honshizuku) That is, according to the present invention, as a basic configuration.

The high-speed fluid discharged from the outer periphery of the first impeller is transferred to the second impeller, which rotates in the opposite direction to the first impeller, immediately or appropriately decelerated and pressurized by a vortex chamber or, if necessary, a fixed winged diffuser. The feed velocity of the fluid at the outlet of the second impeller is increased by creating a large relative velocity at the inlet of the second impeller, and by decelerating it due to the expansion of the blades of the second impeller and decelerating it due to its own rotation. By significantly reducing dynamic pressure, dynamic pressure is efficiently converted to static pressure, and by introducing it at low speed into a spiral casing around the outside of the second impeller or an appropriate casing with a wide space, fluid friction loss within the cage puffer can be minimized. , a biaxially inverted centrifugal fluid pressure booster characterized by being configured to obtain a large pressure increase of the fluid with high efficiency as a whole;
is obtained.

Here, just to be sure, I would like to clarify the comparison with a fixed diffuser for a conventional turbo blower or compressor.

First, we will discuss a comparison with a spiral casing, that is, a winged fixed diffuser, which is said to be more efficient than a winged diffuser.

The limit for efficient deceleration with a fixed winged diffuser is 1/3 of the diffuser exit speed/inlet speed ratio.
It is said to be between 1/4 and 1/4. For example, 240 at the entrance.
m/sec can only be decelerated to 80 to 60 m/see, after which most of the dynamic pressure is absorbed by the friction within the caking.

If this reduction ratio is increased, the length of the guide blades will become large, and the diffuser and casing will also become large, resulting in unfavorable results.

On the other hand, when a rotating diffuser is used, the diffuser blades form a path that spreads out in the same way as a fixed diffuser, which slows down the flow, and also rotates, which slows down the flow, and then reverses the flow. making it possible to introduce the impeller into a spiral casing around the outer periphery of the impeller or any suitable casing having a large space and low flow frictional resistance at a significantly lower absolute speed than any suitable conventional type; Moreover, as is known from Figure 2, the length of the n-th wheel, that is, the blade of the rotating diffuser, can be made shorter than the blade of the fixed diffuser, so the friction loss inside the wheel is also reduced, and the overall effect of these is to increase the length of the blade. This means that static pressure can be increased with high efficiency, and all of the above effects are due to the rotating diffuser.

Note that in the second impeller, the fact that the inlet and outlet absolute velocities in the circumferential direction are reversed indicates that a large pressure increase is achieved there. but,
In this type, the circumferential minute velocity of the absolute velocity at the exit of the second impeller may be in the opposite direction in the range exceeding the design flow rate, and in consideration of this, the width is rather wider than that of the spiral type casing. In some cases, it is better to use a sword with a casing that has a space so that it can flow easily in any direction.

Next, the present invention will be explained with reference to the drawings.

Fig. 1 is a cross-sectional view of the simplest embodiment of the radiation type blower according to the present invention, as seen from the side, and Fig. 2 is a main part of the embodiment of Fig. 1, as seen from the suction side. 3 and 4 are velocity diagrams at the inlet and outlet of the first impeller in the embodiment shown in FIG. It is a velocity diagram of an exit.

The reason why the configuration shown in FIG. 1 is the simplest is that the first and second impellers are directly fixed to the shaft of the motor. Since this type has an open suction port, it is not possible to connect piping to the suction port, but there are many applications as forced air blowers that are used with the suction port open. In FIG. 1, l is a motor for the first impeller, 2 is a first impeller wheel, 3 is a complete set of impellers, and 4 is an inducer for increasing the efficiency of the first impeller. This may be omitted in some cases. Reference numeral 5 denotes a vortex chamber provided between the first impeller 3 and the second impeller 6, which has a vortex chamber of lO in the radial direction.
The size is about 20 to 20 degrees, and is intended to reduce noise and equalize the flow, that is, to improve the efficiency of the second impeller 6. However, if the vortex chamber 5 is made too large, the second blade/wheel 6 will become larger and the loss of its rotating disk will become larger, which is disadvantageous. 7 is the motor for the second impeller 6, 8 is the second impeller shaft, and 9 is the spiral casing, but in some cases the circumferential speed of the absolute velocity of the airflow at the exit of the second impeller becomes almost zero. If it can be designed like this, it would be better to use a sword that is rotationally symmetrical or has a suitably wide space so that it can flow easily.

10 is a casing front cover that is opened when installing and removing the blade wheel.

The reversing second impeller is very effective in increasing the efficiency of centrifugal blowers, but on the other hand, it is necessary to devise a seal to reduce the amount of fluid on the discharge side leaking to the low pressure side. That is, in FIG. 4, the cover of the second impeller 6 is divided into two rings with a large diameter and a small diameter, with a circumference larger than the outer diameter of the second impeller as a boundary, and the small diameter part 11 is divided into two rings with a large diameter. It is preferable that the mouth ring be configured such that it can be attached to and detached from the first impeller, and that the mouth ring be provided so as to overlap the mouth ring of the first impeller. It is important to make this small-diameter portion 11 as thin as possible, light and well-balanced so that the dynamic balance does not change substantially when it is attached or removed. In addition, the mouth ring 12 of the first impeller and the mouth ring 1 of the cover 11 are
3 and maintain an appropriate gap so that it can perform its role as a seal well.

In some cases, the inside of the mouth ring 13 is configured into a labyrinth school. In addition, its outer surface is sealed by a labyrinth 14 supported by the front cover 10 [to reduce leakage of high-pressure fluid on the discharge side of the second impeller. A balance piston 16 of the second impeller is provided, and a labyrinth 17 seals between them.

The outer periphery of the balanced piston 6 is sealed by a labyrinth seal 1B supported by a casing 9.

FIG. 2 shows an example of a preferred blade wheel shape based on the idea of the present invention. That is, the first impeller has straight radial blades and an inducer, and is designed to generate large dynamic pressure with a small outer diameter. Since it is difficult to efficiently convert the large dynamic pressure into static pressure using a conventional fixed diffuser, the wing equipment is usually made with swept wings to increase the recoil as much as possible and reduce the dynamic pressure. However, in the end, only the efficiency described above is obtained. In the present invention, by inverting the second impeller with respect to the first impeller, the speed lf diagrams are shown below as 3.4, 5,
6 As explained in each figure, it is possible to obtain a large increase in static pressure by the second impeller and significantly reduce the dynamic pressure of the discharged fluid, and as described later, the rotational speed of the second impeller Since the amount is only about 1/2 to 1/1O of that of the first impeller, the disc friction loss does not become large, and after all, extremely high efficiency can be ensured for both the first and second impellers.

Figure 3 is a velocity diagram at the inlet of the first impeller, where ul is the circumferential velocity at the average diameter of the blade tip of the inducer suction port, eml is the flow velocity in the j-axis direction, and the relative velocity between the fluid and Tatsumi is W,
It is said that Since Dl is small, ul is small, and eml is originally small, so wl is also small, and therefore turbulence in the flow at the inlet and friction loss are small. This is the effect of an inducer.

Figure 4 is a velocity diagram at the exit of the second impeller, where sue is the circumferential speed,
The relative velocity W is the flow velocity in the half direction to Cm2, is Cmt T
ic equal. To simplify the explanation, let us assume that the number of blades is large enough to allow the fluid to cross accurately in the radial direction, and that the so-called wave coefficient is 1, then the absolute velocity of the discharged fluid is C! becomes. However, since the relative speed W is small and the blade length is short, the frictional resistance inside the blade wheel is very small.

In other words, it can be said that the pressure loss within the width-type blade wheel itself is small and the blade wheel self-efficiency is very high. Instead, the absolute velocity of the fluid at the outlet C! The problem is the efficiency of the diffuser.

FIG. 5 is a velocity diagram at the inlet of the second impeller. ul is the second m
1iL is the peripheral speed at the inlet, but due to various relationships, this is
It is preferable to set the speed to about 1/2 to 1/10 of the circumferential speed at the exit of the impeller. Csg is the local velocity of the fluid flowing into the second impeller,
Said speed C! When the minute velocity in the circumferential direction of is Cu.

Here, D is the exit diameter of the second wing wheel, D is the diameter of the second wing wheel, and when the vortex chamber 5 is enlarged and D is enlarged,
Cul is decelerated to a corresponding degree and becomes Cub, and static pressure increases corresponding to the difference in dynamic pressure. When the vortex chamber 5 is set to an appropriate size, the efficiency of converting the dynamic pressure into static pressure is high.
It also has the effect of reducing friction loss inside the second blade car,
Don't make it too big.

However, if a very high pressure rise is required, and as a result the relative speed between the first and second impellers greatly exceeds the speed of sound, a diffuser with short blades is provided around the outer periphery of the eddy electric 5. It is also possible to introduce it into the second impeller after appropriately decelerating it, but this is omitted from the diagram.

Fig. 6 is a velocity diagram at the exit of the second impeller. Compared to Fig. 5, the relative speed inside the impeller has been reduced by about half from w3 to w4, and this is due to the deceleration due to the widening of the passage inside the second impeller. It is something. On the other hand, the absolute circumferential velocity Cu4 is slightly reversed from that at the entrance due to the circumferential velocity u4.

It is essential that this reversed speed be kept low in order to reduce friction loss during introduction into and discharge from the casing on the outer periphery of the impeller. However, if the flow rate is thicker than the designed value, this Cu4 approaches zero, and if the flow rate becomes even larger, it will move in the opposite direction.

If you want to respond to a wide range of flow rate changes,
It is best to place the casing in a suitably wide space so that it can flow easily in either direction, rather than in an i1% form. Generally speaking, the most important thing to improve efficiency is to use as high a rotation speed as possible for the first impeller and to minimize the exits of the first and second impellers. In order to reduce the diameter of the first impeller, it is best to make the bearing diameter of the first impeller smaller, and to make it more stable, the first impeller should have a double-supported structure. It is installed inside the nozzle on the suction side of the car, and the other drive side bearing is made hollow and passes through the shaft of the second impeller, and is further extended and thrown outside the axis of the second impeller! The configuration shown in FIG. 7 is preferable. In some cases, it is of course possible in practice to increase the rotational speed of the second impeller to the same degree as that of the first impeller to achieve a large pressure increase. In this case as well, the efficiency can be improved over the normal two-stage system. 7th
In the figure, parts common to those in FIG. 1 are denoted by the same reference numerals.

In this case, the lubricating oil in the first bearing on the reduced vehicle side is sent to the bearing via the lip 19 shown in FIG. 87, and is discharged via the lip 20. Therefore, a mechanical seal is used to seal the oil, a pressure tar caused by discharged fluid is used, or both seals are used in combination. 21 in FIG. 7 indicates the shaft seal portion, and 22 indicates the bearing portion.

In addition, in this case, the bearing on the impeller side of the second impeller is in the casing 9.
The other bearing is supported by a bearing housing 23 which is connected to the core housing nog by a 7-flange with an inlet. Also, the first due east is already-
The bearing 7 is supported by a bearing 24 connected to the shaft receiving housing 23 by a flange with an inlet. However, the bearing may be supported in any other suitable manner.

The first impeller is driven by a gear, belt, or direct connection to a motor at the end of the shaft 2, and the second impeller is driven by a pulley or gear 25 fixed to the shaft 8. This configuration requires a slightly higher cost for the lubrication system and its seals, but since the first impeller is dual-supported and the shaft length is not very long, it is suitable for achieving high speed rotation and high efficiency. The structure of the suction port becomes simple. In FIG. 7, the inducer at the inlet of the intermediate wheel is not separated as in FIG. 1, but is integrated, and the effect remains the same. The figure shows a sealed ball bearing as the bearing for the second impeller, but the bearing and its lubrication can be selected as appropriate.

One way to simplify the cooling of the bearing on the side of the first blade wheel and to prevent any leakage from entering the aircraft is to cool the bearing on the gN side of the blade wheel. out of the suction port casing, the shaft is extended through the hollow second impeller lll1lI8, and the second impeller shaft 8 is extended.
It is also practical to provide support by a bearing located separate from the holder, as shown in FIG. In this case, the shaft of the first impeller becomes quite long and needs to be made correspondingly thicker in order to stabilize the rotation of the shaft.As a result, the shaft of the second impeller also becomes thicker, and Although it is unavoidable that the bearing loss increases, the rotational speed of the second impeller is small and the loss itself is not inherently large, so the effect on the overall efficiency is small. Note that FIG. 8 shows an example of a mixed flow type impeller, which is a modification of the centrifugal type. If attention is paid to reducing the gap between the shroud 26 and the blades of each impeller, the mixed flow type impeller can not only provide high efficiency but also simplify the configuration of the second impeller, so the present invention This is preferable as a booster with a biaxial inversion structure. However, it is not suitable for very low flow rates. In Fig. 8, the driving means is a speed increasing gear for the first impeller, a belt transmission for the GZI wheel, and although the motors are separate, they are all belt driven or all gear driven.
The motors may be separate, or one motor may drive both blades, and the selection is made as appropriate. This also applies to other examples.

In general, electric motors of 55kW or more are made to order, and the larger they are, the more expensive they become.In fact, it would be cheaper to divide them into two units, and the starting current would be reduced by staggering the start of each motor. Small and easy. For example, 55k
A sword that has two smaller capacity than one VR, motors, starters, etc. are much cheaper, and one startup! This means that the flow will be small.

In Figure 9, the bearing on the suction side of the first impeller is either installed separately outside the suction casing as in Figure 8, or the other bearing is installed inside the shaft of the second impeller, as shown in Figure 8. This is an example of the device of the present invention which prevents the shaft of the first impeller from becoming long and also prevents the bearing of the second impeller from becoming excessively large. By arranging the bearing for the first impeller in this way, the shaft of the first impeller can be made very short, which greatly increases the primary critical rotational speed, and also prevents the shaft seal from failing. Even if lubricating oil leaks, there is no risk of it getting into the machine. Instead, the lubrication system for the bearing of the first blade wheel provided in the second blade axle becomes a little more complicated. This bearing is called an in-shaft bearing for short, and together with the bearing of the second impeller located on its outer periphery, it is called a double bearing.The outline of its lubrication system is shown in Figure 9, and the details are as follows. will be explained with reference to FIG.

In FIG. 9, the two bearings supported by the second wheel are each supported by bearing housings 23 and 27, which are fixed in sequence to the casing 9 by seven flanges with locks. The oil is supplied from the oil inlet 29 provided in the lubricating oil housing 28, which is fixed to the tip of the bearing housing/g 27 by a flange with an inlet, through the elongated hole 30 drilled in the axis of the shaft 8 of the second impeller. The discharged oil is then supplied to the impeller side bearing of the second impeller through a number of pores drilled in the inner periphery of the cavity around the end of the bearing. The oil is collected and discharged without leaking to the outside by the oil collecting mechanism provided in the housing/g.

Note that %32 is a lubricating oil shaft seal provided at the end of the shaft 8, which may be selected from a mechanical seal, a threaded seal, or the like as appropriate.

FIG. 10 is a detailed view of the double bearing section, and details of the thrust bearing and the lubricating oil discharge system will be explained using this figure.

In FIG. 10, 33 is a thrust ring fixed to the shaft of the first impeller by a threaded portion 35 adjacent to the journal 34, and receives the thrust of the first impeller by a thrust bearing 36 on the end face of the in-shaft bearing 31. and stabilize the axial position.

The lubricating oil passes between the thrust ring 33 and the thrust bearing 36, is shaken off by the sharply finished thrust ring 17&, gathers in the oil collection groove 3BK provided on the inner surface of the second impeller hub 37, and flows around it. The oil is supplied to the impeller side bearing 40 of the second impeller through multiple oil passage holes 39 provided in the . The seal on the impeller side of the lubricating oil of the bearing 40 has to be balanced against the pressure of the airflow leaking between the balance piston 15 and the labyrinth 17, and special consideration must be given to it.

In other words, U! Airflow is the second gm balance piston 16
Although the outside air is released through a number of ventilation holes 41 drilled on the circumference of the inner corner of the thrust ring 3, some pressure still remains inside the paranos piston 16.
3 is formed in a cylindrical shape on the m wheel side, and its thrust bearing has a constriction on the side and sharp edges to facilitate oil drainage. To prevent leakage, a part of the airflow flows into the labyrinth of the bearing nozzle 23 through the ventilation hole 4B so that it does not mix with the main flow of lubricating oil.
The air is collected in a space 44 provided on the side of the impeller, and is discharged from an outlet 45 provided at the bottom of the space 44. Since there is a possibility that a small amount of oil may be mixed into this airflow, a buffer is provided to separate the oil.

The space 44 is constructed by extending a thin-walled sleeve 46 from the balance piston 16 and attaching a cover 47 having a labyrinth surrounding the outer periphery to the bearing housing 23.

On the other hand, the oil that has lubricated the bearing 40 is collected in a space 48 provided on the other side of the bearing housing 23, and is discharged from an oil drain port 49 provided at the bottom thereof.

Since the shaft inner ring bearing 31 rotates in the opposite direction to the journal 34, its relative speed is large and the resistance and heat generation are correspondingly large (this is why it is better to slightly increase the clearance of the bearing to alleviate this). This is because the shaft stability is maintained even if the clearance is increased slightly due to the large relative speed.This configuration allows the shaft length of the first impeller to be the shortest, and the central part of the shaft is optically separated. It can be made thicker and the high-speed stability of the shaft can be improved, making it suitable for cases where extremely high speeds are required.

FIG. 11 shows that the shaft of the first impeller 3 fixed in a cantilever manner passes through the shaft of the hollow second blade N6, and the bearing on the impeller side is inserted into the hollow shaft of the second impeller 6 on the impeller side. A simple and preferable configuration is shown in which the second bearing is fixedly provided at a position where the shaft of the second wing Jg3 passes through the second Y (wheel 6) and extends out.

4 parts 2 among the symbols written in Figure 311
All the others except 3m, 40g, 44a, and 47m indicate the same elements as already described. In addition, the shaft bearing 3
In the examples shown in FIGS. 9 and 1O, the oil lubricating the cylinder 1 is further led to the bearing of the Vc second impeller 6, but in the example shown in FIG.
Stop guiding to the impeller side bearing 40a of the I wheel 6,
The liquid is discharged from the upper and lower discharge ports 45 through a space 44a formed by the bracket 23a that supports the bearing and the recessed portion of the cover 47a. The reason why the discharge ports 45 are provided at the top and bottom is to release part of the gas leaking from the balance piston from the lubricating oil from the top, and to discharge oil from the bottom.

Lubricating oil is supplied from the oil supply hole 29 provided in the bearing housing that supports the other bearing 40 of the second impeller 6, and the lubricating oil is supplied to the in-shaft bearing 31 from the second impeller 6 of the hollow shaft 8 of the second impeller 6. It enters along the shaft 2 from the gap between the end of the shaft and the shaft 2 of the first impeller 3, and also enters the bearing of the housing 24 in the opposite direction to be cleaned and discharged respectively. .

A coupling is attached to the end of the shaft 2 of the R-th wheel 3, and the drive shaft 1
The drive shaft is connected to a and the speed of the drive shaft is increased by an appropriate speed increasing device. However, instead of a coupling, a pulley or a gear may be attached to increase the bell hm speed or gear speed. In this case, the belt should be pulled with equal force from both sides,
Driven by dual motors. Although the second impeller is shown as being driven by a pulley 25, this may also be a gear.

Note that the No. 1 impeller in FIG. 11 does not have a side plate and only has a mouth ring, which serves both to reinforce against the centrifugal force of the inlet blade and to prevent fluid leakage.

Further, although there is no vortex chamber on the outer periphery of the second impeller, and the second impeller rotates within a wide casing, the second B wheel 6 itself may be regarded as a type of improved vortex chamber.

(Effects of the Invention) With the above configuration, the feeder machine and the compact machine of the present invention have a total efficiency of 65 to 75 liters compared to the conventional machine, respectively.
It can be improved to 85 to 90 inches, so relatively speaking it is 2
Since power savings of 0 to 30% can be expected, the price increase due to the biaxial reversing type can be amortized within a year due to the power savings. For example, 3013
Taking a kW high-pressure blower for sewage treatment as an example, the annual electricity cost is 36 yen per 1 kW.
When operating for 0 days, 3t10X20X24X36 (1=
51,840,000 yen, and if we were to be able to save 20 yen, the savings would be about i,ooo million yen per year, while the price increase would be a fraction of that, and within - years. Of course, it is possible to depreciate the cost, and it may be possible to abolish the conventional blower and replace it with the blower of the present invention. Especially the 8-stage turbo blower that is directly connected to the motor.
Because of its low speed and the bending loss of the flow, the efficiency is only about 65%, so if you replace it with one with an efficiency of 85%, you will get an efficiency increase of about 30%, which means a power consumption of about 30%. This is a huge benefit as it saves money. Additionally, this type of blower requires less floor space, so a larger capacity blower can be placed on the same floor space.

Although the blower of the present invention has been mainly described above, the present invention is not limited thereto, and although there are some differences in sealing, it can be configured similarly as a turbine pump and achieves the same increase in efficiency. Benefits can be obtained.

[Brief explanation of the drawing]

Fig. 1 is a cross-sectional view of the simplest embodiment of the axial flow type blower according to the present invention, seen from the side, and Fig. 2 is a main part of the embodiment shown in Fig. 1, seen from the suction side. The sectional view, FIG. 3 and Coffin 4 are speed diagrams of the inlet and outlet of the second impeller of the embodiment shown in FIG. 1, and FIGS. 5 and 6 are the inlet of the second impeller of the embodiment of FIG. 1. FIG. 7 is a side sectional view of the -7 drive side bearing of the embodiment shown in FIG. 1 extended and installed outside the axis of the second impeller, and FIG. A side sectional view of an example of a mixed-flow type impeller, which is a modification of the centrifugal type, and Fig. 9 shows a lubrication system when a -7 bearing is provided in the shaft of the second impeller, which is a modification of the embodiment shown in Fig. 8. Figure 1O is a side sectional view showing the details of the lubrication system in Figure 9, Figure 11 is a cross-section when the shaft of the first impeller is provided through the second impeller. It is a diagram. In the figure

Claims (7)

[Claims] (1) High-speed fluid discharged from the outer periphery of the first impeller is immediately transferred to the second impeller, which rotates in the opposite direction to the first impeller, or to a vortex chamber or, if necessary, to a fixed winged diffuser. Therefore, it is introduced after being appropriately decelerated and pressurized, and a large relative speed is created at the inlet of the second impeller, and the second impeller is decelerated by the spread of the blades of the second impeller and by its own rotation. Dynamic pressure is efficiently converted into static pressure by significantly reducing the absolute velocity of the fluid at the car outlet, and the fluid is introduced at low speed into the spiral casing around the outer periphery of the second impeller or an appropriate casing having a wide space. A biaxially inverted centrifugal fluid pressurizing device characterized in that it is configured to reduce fluid friction loss within the casing and to obtain a large pressure increase of the fluid as a whole with high efficiency. (2) the shaft of the first impeller extends to its suction side, and the shaft is either the shaft of the prime mover or is supported by a bearing so as to be rotated by the prime mover; The axis of the second wheel extends in the opposite direction to the axis of the first wheel, and the axis is either the axis of the prime mover itself or is supported by a bearing so as to be rotated by the prime mover. A biaxially inverted centrifugal fluid pressure booster according to claim (1). (3) The shaft of the second impeller is formed hollow, and the end of the first impeller axle extending toward the suction side of the first impeller is supported by a bearing and a shaft seal from the peripheral wall of the suction nozzle by a rib. The first blade axle extends from the back of the first impeller, and is supported by the ribs, and has passages for supplying and discharging lubricating oil and, if necessary, passages for sealing fluid in the ribs. The two-shaft reversible centrifugal fluid booster according to claim 1, characterized in that the device extends through the shaft and is supported by a bearing provided outside the hollow shaft. (4) The shaft end of the first impeller on the impeller side is hollow, and a bearing is provided therein to hold one shaft end of the first impeller, and the lubricating oil for this bearing is supplied to the center of the second impeller. The discharged oil is further guided to the impeller side bearing of the second impeller, and the other bearing of the first impeller is installed outside the suction casing provided on the impeller suction side. A biaxially inverted centrifugal fluid pressurization device according to claim (1). (5) The shaft of the second impeller is hollow, and the cantilever shaft of the first impeller is passed through concentrically, and the bearing on the impeller side of the first impeller is installed within the shaft at the end of the hollow shaft on the impeller side. , the other bearing is fixedly provided at a location where the shaft of the first impeller is out of the hollow shaft, and the lubricating oil for the first impeller bearing is provided between the bearing housing of the second impeller and the first impeller. The inner shaft bearing is supplied from the gap between the hollow shaft end and the shaft of the first impeller, and in the opposite direction, the other fixed bearing housing and the first impeller are connected to each other. The biaxially reversible centrifugal fluid pressure boosting device according to claim 1, characterized in that the device is configured to enter along the axis of the impeller and discharge as appropriate. (6) The shaft of the first impeller is formed hollow, and the shaft extending to the suction side of the first impeller is supported by a bearing provided outside the suction casing, and the shaft extending to the other side is the shaft of the second impeller. The biaxial reversible centrifugal fluid pressurizing device according to claim (1), wherein the two-shaft reversing centrifugal fluid pressurizing device is supported by bearings provided outside the hollow shaft and passing through the hollow shaft. (7) The side plate of the first impeller is divided into two rings with a large diameter and a small diameter ring with a circumference larger than the outer diameter of the first impeller as the border, and the small diameter part can be attached to and removed from the large diameter part at will. A biaxially inverted centrifugal fluid pressurizing device according to claim 1, characterized in that a mouth ring is provided in the small diameter portion.