KR20140053694A - Pump impeller - Google Patents

Abstract

The present invention relates to a pump impeller which can minimize the clogging phenomenon when slurry is transferred, so that the pump impeller is effectively driven even when the specific speed changes. The pump impeller includes a hub forming a center portion; and multiple blades formed on the hub to suck and discharge an object to be sucked from a mouth to an exit. The twisting angle (β) of the blade is determined by the following equation: β = tan^(-1)(ν_m/(u-( ν_(1u) + ν_(2u))/2)), where U denotes the velocity of the object in a circumferential direction, ν_m denotes average velocity of the object in an axial direction, ν_(1u) denotes the velocity of the object flowing in at the mouth, and ν_(2u) denotes a value of a tangential component of the velocity at the exit.

Description

[0001] PUMP IMPELLER [0002]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a pump impeller, and more particularly, to a pump impeller capable of minimizing clogging during slurry transportation and enabling efficient operation even when the speed is varied.

In order to solve the environmental pollution problem that has recently become serious, the water quality standard of wastewater discharge is gradually strengthened. In response to this, industry also invests in research and development to secure pollution prevention technology that can treat wastewater at low cost have.

Conventional technologies that are currently used to purify high-concentration plant wastewater and livestock wastewater are membrane separation technology, which improves the performance of major components including pumps as well as membrane performance at the time of initial development, and is now widely used in various forms.

One of the most important components in the high concentration wastewater treatment process is the recycle pump, and the power consumption of the circulation pump dominates the power consumption of the entire system. In addition, the nitrogen removal process in the wastewater treatment process requires a large amount and a large amount of wastewater to be recirculated for a long time compared with the conventional organic material process, so that the selection of an efficient recycling pump is very important It is important.

Although a 50% efficiency centrifugal pump is typically used in these fields, the problem of clogging of the pump due to the characteristics of the slurry pump, which is required to transport liquid containing solid particles, is frequent and the stable pumping action is difficult have.

Therefore, in the related art, it is necessary to improve the pump impeller which can improve the pumping efficiency in a wide operation range while minimizing the clogging phenomenon. For this improvement, the importance of the impeller design that reflects the characteristics of the slurry appropriately has been raised .

It is an object of the present invention to provide a pump impeller capable of minimizing clogging during slurry transportation and enabling efficient operation even when the velocity is varied.

According to an aspect of the present invention, there is provided a pump impeller comprising:

A central hub; A plurality of blades formed in the hub and configured to draw in the inhaled material at the inlet and exit to the outlet; , And the twist angle (beta) of the blade is determined by the following equation.

(Where u is the circumferential velocity, v _m is the axial average velocity, v _1u is the inlet velocity at the inlet, and v _2u is the tangential velocity component of the outlet velocity).

At this time, when flowing at a right angle to the inlet of the pump impeller, the u, v _m, v _1u and _2u v is calculated by the following equation, respectively.

(Wherein, Q is flow rate, D is the diameter of the impeller, D _O is the outer diameter of the impeller, Di is the inner diameter of the impeller, N is the number of revolutions of the impeller, H is the head, η is the efficiency.)

At this time, the twist angle beta is in the range of 14 to 30 degrees.

In this case, preferably, the warp angle? Of the blade ranges from 40 to 140, and the length ranges from 170 to 250 mm.

In the present invention, more preferably, the warping angle alpha of the blade is 110 deg., The twisting angle beta is 20 deg., And the length (l) is 240 mm.

According to another aspect of the present invention, there is provided a pump impeller,

A central hub; A plurality of blades formed in the hub and configured to draw in the inhaled material at the inlet and exit to the outlet; Wherein the twist angle beta of the blade is in a range of 14 to 30, the warp angle alpha is in a range of 40 to 140, and the length is in a range of 170 to 250 mm.

According to the present invention, the efficiency of the pump is reduced to 60% by setting the twist angle (β) of the blade implemented in the pump impeller to 14 to 30 °, the deflection angle (α) to 40 to 140 °, and the length (ㅣ) to 170 to 250 mm. Or more, especially about 87%. Therefore, the efficiency can be remarkably improved as compared with the conventional centrifugal pump of about 50%.

According to the present invention, the twist angle, the deflection angle, the length, and the like of the blade are realized as described above, thereby minimizing the clogging of the slurry during transportation and enabling efficient operation even at a low specific speed.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a view showing the shape of a pump impeller according to an embodiment of the present invention. FIG.
Fig. 2 is a perspective view showing a state in which the pump impeller of the present invention is installed on a guide pipe; Fig.
3 is a sectional view for designing the dimensions of a pump impeller according to the present invention.
FIG. 4 is a graph showing changes in twist angle and length of a blade according to a change in diameter of a pump impeller according to the present invention. FIG.
5 is a view illustrating a shape design process of a blade using a 3-D shape design program according to the present invention.
6 is a view of a grid used in the computational flow analysis according to the present invention.
7 is a view showing a shape change of a pump impeller according to a change in twist angle of a blade according to the present invention.
8 to 10 are graphs showing influences on the pump impeller efficiency, head, and shaft power according to the change of the twist angle of the blade according to the present invention.
11 to 14 are graphs showing influences on the shape, efficiency, heading, and axial force of the pump impeller according to the length of the blades according to the present invention.
FIG. 15 is a graph obtained by using a 3-D model in order to examine the influence of the twist angle and length of the blade on the efficiency, head, and shaft power of the pump according to the embodiment of the present invention.
FIGS. 16 to 19 are graphs showing influences on the shape, efficiency, heading, and axial power of the pump according to the change of the warp angle of the blade according to the present invention.

Hereinafter, preferred embodiments of the present invention will be described with reference to the accompanying drawings. In the following description, a detailed description of known functions and configurations incorporated herein will be omitted when it may make the subject matter of the present invention rather unclear.

2 is a perspective view of a pump impeller according to an embodiment of the present invention installed on a guide pipe. FIG. 3 is a cross-sectional view of a pump according to an embodiment of the present invention. Referring to FIG. 1, Sectional view for the dimensional design of the impeller.

1 to 3, a pump impeller 100 according to an embodiment of the present invention is inserted into a guide pipe 10 and is rotated by a motor (not shown) And suck in the inhaled material to be discharged through the outlet (30). The pump impeller 100 of the present invention includes a hub 110 corresponding to a central portion and a plurality of blades 120 formed on the hub 110.

The hub 110 is not limited in form, but is preferably implemented in a substantially cylindrical shape. The shaft 111 connected between the hub 110 and the motor transmits the driving force of the motor to the hub 110 to rotate the pump impeller 100. The diameter d of the shaft 111 considers the value of the stability factor of 1.72 in consideration of the square key groove. The material of the shaft 111 is, for example, SM30C-3, which is about 34 mm when determining the diameter (d) of the shaft (111) by using this value because the allowable torsional stress value is 205 kgf / The diameter dh of the hub 110 is determined to be 1.8 times or more of the diameter of the shaft 111 (dh ≥ 1.8d), and in this embodiment, It is determined to be 80 mm.

In this embodiment, the blade 120 may have three or more identical blades 120 so that the slurry can be easily pumped. As shown in FIG. At this time, it is preferable that the three blades 120 are formed to be balanced at an angle of 120 degrees with respect to each other.

In order to maximize the efficiency of the pump impeller 100 having such a structure, not only the inlet shape according to the viscosity of the slurry but also the blade shape of the blade 120 can be adjusted so that the wastewater sucked at the time of designing the pump impeller 100 passes through the tip of the blade 120, It is very important to set the profile of the server 120.

Thus, in the present embodiment, the outer diameter D ₀ of the pump impeller 100 is first determined from the following formula (1) from the number of revolutions N and the head H at the design point.

In the equation (1), Ku is a value determined from the specific speed (Ns), and in the present embodiment, Ku = 3.8 when Ns = 4195 is used. The outer diameter (D ₀₎ of a pump impeller 100 is calculated therefrom using the formula (1) is 392㎜, and finally decided 400㎜. In the case of the inner diameter Di of the pump impeller 100, it is determined by v in the following equation (2) which is determined according to the specific speed Ns. In this embodiment (Ns = 4195) So that the inner diameter Di of the pump impeller 100 becomes 0.2D ₀ and is determined to be 80 mm.

Next, the length l of the blade 120 is determined. When the number (z) of blades 120 is set to 3 by using the experimental table presented by Stepanoff, the reference value (1 / t, t: pitch) is firstly calculated from the vane spacing . In this embodiment, Ns = 4195 and v = 0.2. Therefore, the table has a vane spacing value of the following equation (2) from the table. Finally, the length l of the blade 120 is calculated by using the pitch t of the blade 120 (t =? D ₀ / z = 420 mm), and 240 mm is determined as follows. However, the vane spcing is designed to have a value 1.3 times greater in terms of stiffness as it goes to the inside of the blade 120, and is calculated to increase linearly to 1.3 while moving from the outer diameter to the inner diameter of the blade 120.

The value to be determined next is the twist angle (?) Of the blade 120. The twist angle beta is very important in designing the pump impeller 100 of the present invention. The twist angle beta of the blade 120 is influenced by the shape of the blade. In this embodiment, it is determined from the following equation (4).

In the above equations, the axial average velocity vm can be 2.56 m / sec when the flow Q and the inner and outer diameters D _o and Di of the pump impeller 100 are used. The circumferential velocity u depends on the diameter D of the position to be calculated and can be calculated using the number of rotations N and the outer diameter D of the impeller 100. [ Further, assuming that the flow rate introduced from the inlet 20 is introduced at a right angle, _v1u = 0, and the tangential component value _v2u of the flow velocity of the outlet 30 can be calculated from Equation (7) Efficiency η = 90%).

FIG. 4 is a graph showing changes in blade twist angle and blade length according to the diameter change of the pump impeller according to the present invention.

4, the twist angle β of the blade 120 in the case of the diameter D of the pump impeller 100 obtained from equations (4) to (7) has a value of 12.5 in the outer diameter D ₀ As the diameter D decreases, the twist angle beta increases and finally reaches 62.9 in the inner diameter Di. Conversely, the diameter D of the impeller 100 and the length l of the blade 120 are substantially proportional. As can be seen from these results, since the length l and the twist angle? Of the blade 12 are changed according to the outer diameter, it is easy to implement the shape of the three-dimensional blade 120, which reflects all of these values, Is not.

In the present embodiment, the effect of the twist angle beta of the blade 12, the length l of the blade 120, and the warp angle alpha of the blade 120 on the performance of the pump through the three- And the shape is finally determined from this. Hereinafter, this will be described in more detail.

1) Three-dimensional computational flow analysis range

After determining the basic dimensions of the pump impeller 100 according to the present embodiment, the detailed design of the shape is performed by using a 3-D shape design program (for example, ANSYS BladeGen program) and a 3-D flow analysis program (for example, ANSYS CFX program) . The pump according to the present embodiment is an axial flow pump, and the number of the blades 120 can be designed to 2 to 4. However, in this embodiment, considering the direct coupling in terms of the durability of the 12-pole motor and the number of rotations of the blades N = 580 rpm 1 and FIG. 3, the pump impeller 100 having three blades 120 is determined in consideration of the values and the discharge flow values.

The design of the pump impeller 100 according to the present embodiment starts with defining the basic shape dimensions that satisfy the required basic condition flow rate, heading, efficiency, and the like.

5 is a view illustrating a shape design process of a blade using a 3-D shape design program according to the present invention.

Referring to FIG. 5, a dedicated 3-D shape design program (for example, ANSYS BladeGen program) according to the present invention calculates the length l of the blade 120 and the twist angle beta on the basis of the hub and shroud, Dimensional pump impeller 100 when it is appropriately input according to the diameter of the pump impeller 100. 3-D solid modeling of the pump impeller 100 having the shape of the blade 120 is performed prior to the 3-D flow analysis by changing the variable values to be analyzed by using the three-dimensional shape designing program.

As a main parameter to be determined in the three-dimensional configuration and the 3-D calculation of the pump impeller 100 in this embodiment, the length l of the blade 120 constituting the three-dimensional shape of the blade 120, (?) and the deflection angle (?) are important design parameters which have an absolute influence on the performance of the pump. These design variables are values that can be determined only by the three-dimensional flow analysis, unlike the main design variables (D ₀ , Di, z, i) that can be determined through simple calculation as described above.

2) Computational flow analysis method

FIG. 6 shows a lattice used in the computational flow analysis according to the present invention.

As shown in FIG. 6, a tetrahedral grid is mainly used to reduce the time required for forming a grid (more than 130 grid is required for 3-D model) rather than the complexity of the shape. On the wall, Layer prism lattice layer. Then, only one of the three blades 120 is analyzed in consideration of the calculation time, and the total number of elements is about 800,000. Numerical analysis was performed using the ANSYS CFX-12 program, a commercial analysis program, and the working fluid was set as water considering the verification process after the prototype production. The continuity equation and the momentum equation except the energy equation were used as a whole, and the turbulence model was analyzed by adopting the ke model. As a boundary condition, a pressure of 1 bar was used on the inlet 20 side and a flow rate of 18.5 m 3 / min on the outlet side 30, and the analysis was performed at 580 rpm. The analysis is performed using a reference frame rotation (RFR) method in which a rotational coordinate system is set for the pump impeller 100 and the inlet 20 and the outlet 30 are rotated by a counter- Wall) as the boundary conditions.

3) Results of 3-D computerized flow analysis

FIG. 7 is a view showing a change in shape of a pump impeller according to a change in twist angle of a blade according to the present invention, and FIGS. 8 to 10 are diagrams showing changes in blade twist angle according to the present invention, (H) and the axial force (P).

7, the outer diameter (D ₀ = 400 mm), the inner diameter (Di = 80 mm), the length (l = 240 mm), the warp angle (? = 90 mm) of the blade 120 in the pump impeller 100, 20 degrees, and 30 degrees, respectively, while the twist angle beta is fixed to be constant. As can be seen from FIG. 7, when the twist angle β of the blade 120 is changed, the axial length la of the impeller 100 increases along with the twist of the blade 120. When the flow rate (Q) is given as 18.5 m 3 / min, the analysis is performed up to 30 ° by increasing the twist angle (β) from 14 ° to 2 °. In addition, even when the twist angle? Is changed, the shape viewed from the front is the same, but it can be seen that the rotational direction area increases with the axial length la of the blade 120 as the twist angle? Increases.

8 to 10 are graphs showing influences of the efficiency, the head, and the shaft power of the pump impeller according to the blade twist angle?, Respectively. Referring to FIGS. 8 to 10, the twist angle (β) increases continuously with the twist angle (β) in the case of head (H), but in the case of efficiency, the twist angle (β) increases only up to 20 °, . This is because even though the length l of the blade 120 is constant but the rotational direction area increases along with the axial length la of the blade 120 as the twist angle beta increases, This is because the axial force P increases.

9 and 10, as the twist angle β of the blade 120 increases, the axial force increases as compared with the increase of the head. Therefore, in the case of efficiency, the twist angle β is 20 °, As shown in FIG. This efficiency (?) Is calculated as shown in the following equation (8).

γ is the specific weight, Q is the flow rate, H is the head, and P is the shaft power.

It can be seen that the resultant value corresponds to the target value H > 0.5m and P < 7kW because the heading angle (H) is 1.34m and the axial force (P) is 4.4kW at twist angle? .

11 to 14 are graphs showing influences on the shape, efficiency, heading, and axial force of the pump impeller according to the length of the blades according to the present invention.

11, when the length l of the blade 120 increases with respect to the shape of the pump impeller 100, the width of the blade 120 increases along with the axial length la. 12 to 14 are the results of calculating the efficiency, heading, and axial force of the pump while varying the length l of the blade 120 from 170 mm to 250 mm by 10 mm. As can be seen from the figure, the increase in the length l of the blade 120 shows a tendency that the efficiency, the head, and the shaft force both increase slightly.

In case of efficiency, it has the maximum value at 240 mm, and it shows a tendency to decrease again. From this result, it can be seen that the length l of the blade 120, which is one of the design variables, is relatively insensitive to the twist angle?.

15 is a graph using a 3-D model to examine the effect of the blade on the efficiency, head, and axial force of the pump according to the twist angle and length of the blade according to the embodiment of the present invention.

15, the length l of the blade 120, which is a design variable, is relatively less sensitive in terms of the efficiency? Than the twist angle?, But the head H and the shaft force P), it is important in the region where the twist angle beta is large. It can be seen that the axial force increases rapidly with respect to the length (I) of all the blades 120, so that the efficiency value has a maximum value when the twist angle β is 20 °.

FIGS. 16 to 19 are graphs showing influences on the shape, efficiency, heading, and axial force of the pump according to changes in the deflection angle of the blades according to the present invention.

16, when the warp angle? Of the blade 120 increases, the shape of the blade 120 is bent by the rotation direction of the blade 120 (counterclockwise in the drawing). The width of the blade 120 along with the axial length la has an increasing shape. 17 to 19, the efficiency, the head, and the shaft power of the pump 120 are increased while the warp angle? Of the blade 120 is increased from 40 to 10 degrees, , The efficiency, the head, and the axial force tend to increase slightly. And, when the angle of deflection (α) of both efficiency, head, and axial force has a maximum at 110 °, it shows a tendency to decrease. From this result, it can be seen that the warp angle? Of the blade 120, which is one of the design variables, is relatively more sensitive to the performance of the pump than the length (1) (about 11% within the analysis range).

As described above, in the case of the pump impeller 100 according to the present embodiment, in order to have higher efficiency than the centrifugal pump of about 50% in the past, the twist angle β of the blade 120 is in the range of 14 ° to 30 ° , The warping angle? Is in the range of 40 to 140, and the length (?) Is in the range of 170 to 250 mm. It can be seen that the twist angle β of the blade 120 is in the range of 17 ° to 26 ° and the warp angle α is in the range of 50 ° to 140 ° in order to increase the efficiency to 80% Particularly, in the pump impeller 100 according to the present embodiment, the blade 120 has the maximum efficiency at 110 °, 20 °, and 240 mm, respectively, when the warp angle α, twist angle β, . However, in the case of a pump for conveying water, since the deflection angle [alpha] of the blade 120 largely adopts a value smaller than 90 [deg.] Due to a shaft power problem, , 70 °. The analytical performance of these prototypes is shown in Table 1 below.

Details Item value Flow rate (Q) 18.5 m ³ / min Rotation speed (V) 580 rpm Blade outer diameter (D ₀ ) 400 rpm The blade inner diameter (D _i ) 80 mm Blade twist angle (?) 20 ^o Blade length (l) 240 mm Total Efficency
(Target: 35%) α = 110 ° 86.8% α = 90 ° 85.0% α = 70 ° 81.9% Head
(Target: 0.5m) α = 110 ° 1.44 m α = 90 ° 1.34 m α = 70 ° 1.24 m Shaft power
(Target: 7kW) α = 110 ° 4.6 kW α = 90 ° 4.4 kW α = 70 ° 4.2 kW

While the invention has been shown and described with reference to certain preferred embodiments thereof, it is to be understood that the invention is not limited to the disclosed embodiments. Those skilled in the art will appreciate that various modifications, additions and substitutions are possible, without departing from the scope of the appended claims, The genius will be so self-evident. Accordingly, the true scope of the present invention should be determined by the technical idea of the appended claims.

In order to solve the environmental pollution problem, purification technology of high concentration plant wastewater and livestock wastewater is being developed, and recently there is a demand for efficiency of the pump which is one of the most important parts in the high concentration wastewater treatment process. For this purpose, the design of the impeller for the pump is an important factor.

In this respect, the pump impeller according to the present invention exhibits a much higher efficiency than the conventional centrifugal pump in terms of efficiency, and thus can be very usefully used in environmental industry fields such as wastewater treatment and purification technology.

10: guide tube 20: impeller inlet
30: impeller outlet 100: pump impeller
110: hub 120: blade
?: deflection angle of the blade?: deflection angle of the blade

Claims (7)

A central hub;
A plurality of blades formed in the hub, the blades configured to draw in the inhaled material at the inlet by rotation and exit to the outlet; Lt; / RTI &gt;
And the twist angle (beta) of the blade is determined by the following equation.

(Where u is the circumferential velocity, v _m is the axial average velocity, v _1u is the inlet velocity at the inlet, and v _2u is the tangential velocity component of the outlet velocity). The method according to claim 1,
When entering at a right angle to the inlet of the pump impeller, the u, v _m, v _1u and _2u v are the pump impeller, characterized in that, calculated by the following equation, respectively.

(Wherein, Q is flow rate, D is the diameter of the impeller, D _O is the outer diameter of the impeller, Di is the inner diameter of the impeller, N is the number of revolutions of the impeller, H is the head, η is the efficiency.) The method according to claim 1,
Characterized in that the twist angle (beta) is in the range of 14 ° to 30 °. The method of claim 3,
Wherein a deflection angle? Of the blade is in a range of 40 to 140, and a length is in a range of 170 to 250 mm. 5. The method of claim 4,
Wherein the warping angle alpha of the blade is 110 deg., The twisting angle beta is 20 deg., And the length l is 240 mm. A central hub;
A plurality of blades formed in the hub, the blades configured to draw in the inhaled material at the inlet by rotation and exit to the outlet; Lt; / RTI &gt;
Wherein the blade twist angle beta is in the range of 14 to 30 and the warp angle alpha is in the range of 40 to 140 and the length is in the range of 170 to 250 mm. The method according to claim 6,
Wherein the twist angle (beta) of the blade is calculated by the following equation.

(Where u is the circumferential velocity, v _m is the axial average velocity, v _1u is the inlet velocity at the inlet, and v _2u is the tangential velocity component of the outlet velocity).

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