RU2606290C1 - Centripetal impeller pump for transfer of heated non-uniform by density liquids - Google Patents

Abstract

FIELD: pumps.

SUBSTANCE: invention relates to pump engineering. Pump includes inlet branch pipe, centripetal impeller, consisting of drive and covering discs, radial blades (4), bent along circle arc with radius (r_l) with center of curvature (O₂), displaced relative to center of rotation (O₁), and combined with it orthogonal coordinate system beginning radially along axis (Y) at distance (Y_o) and along axis (X) in direction of rotation from axis (Y) at distance (X_o), and conical bushing with decreasing diameter; supply, discharge and shaft, driven by induction motor. Wheel has inlet section, arranged along periphery on impeller (D_1rk) outer diameter, and outlet cross-section, located in central part on inner diameter (D_2rk=(0.3-0.7)D_1rk). Supply is located from outer side, opposite to wheel inlet cross-section, and made in form of helical channel (10) with constant width in circumferential direction, equal to wheel width at inlet, and with height counted from wheel outer diameter D₁ to helical channel wall inner surface (9), decreasing in proportion to angle of coverage (φ_cov), with height (h_sk), counted from initial section height (h_1sk) adjacent to inlet branch pipe and taken as coverage angle zero (φ_cov= 0°), corresponding to pump inlet area, to final height through one turn of current coverage angle (φ_cov=360°), equal to radial gap value (h_2sk=δ_r), providing free rotation of wheel. Discharge is located on outer side, opposite to wheel outlet section (2_pk), and made in form of annular diffuser, formed by outlet branch pipe wall inner surface with inner diameter and outer surface of wheel conical bushing.

EFFECT: invention is aimed at achievement of maximum efficiency due to formation of convection from periphery to center at action of rotary acceleration.

1 cl, 2 dwg

Description

It is known that in the fuel and energy industry there is a technical problem of pumping fuel oil, which is significantly complicated due to their increased viscosity (up to 1000 cSt), which leads to increased hydraulic losses in the flow part of the pumps, lower efficiency and increased energy consumption of pumping. To reduce the viscosity of fuel oils and increase the efficiency of pumping during transportation through pipelines, they are heated (up to about 60 ° C). Osediagonal or centrifugal types of pumps are usually used as pumps for pumping fuel oil. At the same time, there is a significant decrease in the efficiency of the pumps compared to their water characteristics obtained during model tests. Moreover, it turns out that such a decrease cannot be explained only by the influence of viscosity and an increase in the coefficient of friction, but also by some reason. For this purpose, a thermophysical analysis of the flow of the heated fluid in the flow part of the pumps was carried out. First, it is known (see [1] R. Feynman, R. Leighton, M. Sands. Feynman lectures on physics. 4. KINETICS, HEAT, SOUND, Mir Publishing House, Moscow, 1967) that in the aggregate of molecules heated liquid at a certain temperature there are always molecules with different speeds, with different temperatures, and therefore with different densities, the distribution of which can be approximately adopted corresponding to the normal distribution law or the like (see [1] p. 35 of Fig. 40.3). Secondly, the property of a liquid nonuniform in density is also known (see [2] L. G. Loitsyanskii. Mechanics of liquid and gas. Publishing House Science of the Main Edition of Physics and Mathematics. Moscow, 1973) to form rotation acceleration (ε = rω ² ) a centripetal flow directed from the periphery to the center (CONVECTION) (see [2] p. 103, 104, Fig. 20), described by the equation

Where

ρ is the density of the pumped liquid,

r is the current radius,

ω is the angular velocity of rotation.

, и, очевидно, во столько же раз увеличится интенсивность конвекции в поле воздействия ускорения вращения при таких условиях в центростремительном рабочем колесе насоса. Рассмотрение возможного на практике различного сочетания размеров и скорости вращения показывает реально достижимые параметры кратности ускорения вращения в пределах

,In addition, it can be shown that the usually observed convection of air or water near heated bodies in the field of influence of the acceleration of gravity (g = 9.81 m / s ² ) can be significantly increased in intensity in the field of influence of acceleration of rotation (ε = rω ² ). For example, for a prototype centripetal pump with very real dimensions (r = 100 mm) and rotation speed (n = 3000 rpm), we have rotation acceleration (ε = 9869.7 m / s ² ), which corresponds to the magnitude of the rotation acceleration factor at the level

, and, obviously, the intensity of convection in the field of rotation acceleration under such conditions in the centripetal impeller of the pump will increase by the same amount. Consideration of a possible combination of sizes and rotational speeds in practice shows the realistically achievable parameters of the multiplicity of rotation acceleration within

,

Thus, a plausible version suggests itself, explaining the occurrence of additional hydraulic losses in the sedimentary and centrifugal pumps when pumping heated fuel oil. Those. When the nominal diagonal flow in the sedimentary and radial in the centrifugal impellers of the nominal flow occurs, a counter convective flow arises, leading to additional hydraulic losses due to vortex formation, which can be eliminated in the centripetal impeller. It is advisable here to consider the process of CONVECTION in the centripetal impeller. As already mentioned, the observed convection of air or water near heated bodies in the field of action of the acceleration of free fall of bodies occurs as follows: the heated volume rises, and the same less heated volume falls in its place, i.e. displacement of a more heated volume by a less heated one occurs. And in the centripetal impeller, as indicated in [2], in the field of influence of acceleration of rotation, convection occurs in the direction from the periphery to the axis of rotation. Those. the axis of rotation is a special place for fluid concentration points. If, by inertia, the heated fluid slips through the axis of rotation, then it will necessarily return to a special point - the axis of rotation, i.e. there is an accumulation of the volume of fluid in the region of the axis of rotation. Consequently, the fluid pressure rises in the region of the axis of rotation, propagating along the outlet pipe until it exits the pump. Since the magnitude of the parameter of the multiplicity of rotation acceleration can reach thousands or more times with respect to the free fall acceleration of bodies, this process occurs very intensively, which additionally increases the pressure and, consequently, the pump efficiency. Based on the foregoing, a centripetal pump design is proposed as an effective means of achieving the technical result of reducing the energy consumption of the process of pumping heated heating oil. An analogue of a centripetal impeller can be taken to be an impeller of a centrifugal pump with radial blades (practically no different in appearance), but differing in the direction of curvature of the radial blades and having an opposite direction of rotation in the centripetal pump. Thus, as an invention, it is proposed: a centripetal pump for pumping a heated non-uniform density liquid (see Fig. 1 and Fig. 2), comprising:

- inlet pipe 0 with an inner diameter (d ₀ ) and the area of the inlet to the pump (F ₀ ),

- a centripetal impeller 1, consisting of a leading 2 and a cover 3 disks, radial blades 4, bent along an arc of a circle of radius (r _l ) with a bending center O ₂ offset from the center of rotation O ₁ and the origin of the orthogonal coordinate system combined with it, radially along the Y axis at a distance (Y _о ) and along the X axis in the direction of rotation from the Y axis at a distance (X _about ) and a conical sleeve 5 with a decreasing diameter from the initial size d _1в = (0.8-0.9) D _2рк to final size d _2v = (0.1-0.2) D _2rk .

Inlet 6, Outlet 7 and Shaft 8, which rotates the centripetal impeller 1 from the drive, which can be used as an induction electric motor, characterized in that the centripetal impeller 1 has an input section (1 pk) located on the periphery on the outer diameter D _1rk , and with the width of the centripetal impeller 1 at the inlet (A _1rk ), with the area of the formed cylindrical ring equal to (F _1rk = πD _1rk A _1rk ), and the output section (2rk) located in the central part on the inner diameter D _{2 pk} = (0.3-0.7) D _{1 pk} , with the width of the centripetal impeller 1 at the exit (A _{2 pk} ), with the area of the formed cylindrical ring equal to (F _{2 pk} = πD _{2 pk} A _{2 pk} ).

The peculiarity of this shaping radial blades 4, is that at any point M (see. Fig. 2) lying on the surface of the radial blade 4 can be graphically or analytically determine the setting angle of the blade (β _L) as the angle formed between the rays drawn from centers of rotation (O ₁ ) and bending (O ₂ ) at the desired point M;

and Inlet 6 is located on the outside opposite the inlet section (1 pk) of the centripetal impeller 1, is made in the form of a spiral channel 10 with a constant circumferential width (A _ck = const) equal to the width of the centripetal impeller 1 at the inlet (A _ck = A _1rk ), and the height measured from the outer diameter D _{1 of the} centripetal impeller 1 to the inner surface of the wall 9 of the spiral channel 10, decreasing in proportion to the angle of coverage (ϕ _ohv ) height (h _ck ), from the initial section height (h _1sk ) adjacent to the entrance Ohm branch pipe and taken at the beginning of the reference angle of coverage (ϕ _OHV = 0 °), the corresponding area of the entrance to the pump (h _1sk = F ₀ / A _SK ), to a final height after one revolution at an angle of coverage (ϕ _OHV = 360 °), equal to the value of the radial clearance (h _2sk = δ _r ). providing free rotation of the centripetal impeller 1;

and Branch 7, located on the outside opposite the outlet section (2 pk) of the centripetal impeller 1, is made in the form of an annular diffuser 11 formed by the inner surface of the wall of the outlet pipe 12 with an inner diameter (d ₂ ) and the outer surface of the conical sleeve 5 of the centripetal impeller 1 s from the pump outlet area (F _2).

In FIG. 1 shows a centripetal pump in the context with the designation of the main geometric dimensions of the centripetal impeller, inlet and outlet. In FIG. 2 shows a centripetal pump in plan with the designation of the main geometric dimensions of the inlet pipe, the radial blades of the centripetal impeller, the supply and direction of rotation of the impeller.

The need for special inlet profiling is related to the fact that the centripetal direction of the liquid acquires, as shown in [1], only in the centripetal impeller 1 in the field of rotation acceleration, and when the fluid flow in the inlet 6 is necessary to create an external effect on the fluid flow, which and is created by the inner surface 9 of the spiral channel 10. This effect occurs in the following order: in a pump filled with liquid before starting, after starting the blade 4 of a rotating centripet of the impeller 1 begin to interact with the stationary fluid, forming near the inlet edges of the radial blades 4 a tangled whirlwind swirl of countercurrents that entrains nearby fluid layers under the influence of turbulent viscosity in the direction of rotation and spreads its influence throughout the entire cross section of the spiral channel 10. But since the cross sections of the spiral channel 10 decrease in the same direction, creating a velocity and volumetric flow component directed to the center and, therefore, towards the end of the spiral channel 10, when its height and area become close to zero, the entire volume of liquid (V _cc ) filling the spiral channel 10 supply 6 (equal to the volume of the solid triangle with a height h _1sk and a base area F _1rk = πD _1rk A _1rk )

V _ck = h _1sk πD _1rk A _1rk / 2

will enter the centripetal impeller 1. It is assumed that such a process of formation of a tangled vortex sheet will continue after starting up when switching to the nominal operating mode. In a unit of time this gives the calculated value of the volumetric flow rate of the pump:

Q _p = n _with K _pr h _1sk πD _1rk A _1rk / 2

Where

n _s [r / s] - the number of shaft revolutions in 1 s,

To _pr ≤1 - slippage coefficient of the satellite stream of the vortex shroud countercurrents relative to the rotating centripetal impeller. It is noted that the slip coefficient due to the complexity of the process cannot be determined exactly. An estimate of its value can be made experimentally when testing a prototype centripetal pump according to the following scheme:

Q _p = n _with K _pr h _1sk πD _1rk A _1rk / 2 and Q _t = n _with h _1sk πD _1rk A _1rk / 2,

Where

Q _p - the volumetric flow rate measured by the flowmeter when testing a prototype centripetal pump,

Q _t - theoretical volumetric flow rate, calculated according to the geometric parameters adopted during design (h _1sk D _1rk A _1rk ) and rotation speed (n _s ),

where from

K _ol = Q _p / Q _t ≤1

When designing a prototype centripetal pump, the product K _pr n _s = 1 was approximately adopted, which gives for n = 2000-3000 rpm and n _s = 33.3-50.0 r / s the design value of the slip coefficient K _pr = 0 , 03-0,02≤0,1, which can be assumed approximately corresponding to reality, i.e. the accepted design speed of rotation of the satellite stream is more than ten times less than the speed of rotation of the impeller.

Knowing the blade installation angle (β _l ) and the estimated volumetric flow rate (Q _p ) allows us to calculate and construct velocity triangles at the inlet according to the known dependencies, calculate the theoretical head (N _t ), dynamic (N _d ) in the intermediate sections and at the exit of the centripetal impeller ) and static (N _st ) pressure components:

H _t = U ₂ C _2i / g N _dyn = C _and ² / _2g , N _st = C _and (2U ₂ -C _and ) / 2g

Attention is drawn to the value of the swirl (C _and ) of the fluid flow, the value of which is calculated from the expression

C _and = UC _z / tg (β _l ),

where U is the current value of the peripheral speed U = πDn / 60,

D is the current value of the diameter of the impeller,

n is the rotational speed rpm

C _z - current meridional (discharge) speed (C _z = Q _p / F _pk ),

F _pk - the current area of the centripetal impeller (F _pk = πDA _pk ),

And _pk is the current width of the centripetal impeller,

β _l - the current value of the angle of the blade.

The current distribution of the swirl (C _and ) of the fluid flow along the radius in the centripetal impeller directly affects the acceleration of rotation of the fluid flow (ε _w = rω ² _w ), which is determined through the angular velocity of rotation of the fluid flow from the relation

ω _W = C _and / r.

Therefore, when calculating and designing a centripetal impeller, it is recommended to take the radius distribution of the fluid flow (C _and ) in accordance with the law of “free circulation” (rС _and = const).

In accordance with this law, when moving along the radius of the impeller towards the center, the swirl of the fluid flow (C _and ) naturally increases, and consequently, the acceleration of rotation (ε _W ) also increases.

The question may arise that with this arrangement, the inlet and fluid inlet organization centripetal impeller on the outer diameter D _1rk anti-cavitation quality of the pump will be at a low level, but it is not, and under certain conditions, on the contrary, it is possible to obtain improvement. This is known (see [3] “HIGH-TURNING BLADE PUMPS” edited by Dr. B.V. Ovsyannikov and Dr. V.F. Chebaevsky, Moscow, Mechanical Engineering, 1975) using screw pumps, having the best anti-cavitation qualities among the known vane pumps and those closest in features to the fluid flow at the entrance to the centripetal impeller (by the nature of the fluid flow in the region of the input edges of the blades, with the formation of a tear-off cavitation cavity). A generalization of numerous experimental data showed that the limiting minimum values of the cavitation coefficient (λ _cr ) and the cavitation reserve (Δh _cr ) are achieved at lower values of the flow coefficient:

Where

With _1z - axial (flow) speed at the entrance of the fluid flow into the impeller,

U ₁ - peripheral speed at the entrance to the outer diameter of the impeller.

. Указанных (оптимальных) пределов изменения коэффициента расхода для осевого шнекового рабочего колеса можно достигнуть только при увеличении площади на входе за счет увеличения наружного диаметра D₁, а это сопровождается также увеличением окружной скорости (U₁), что приводит к возрастанию относительной скорости (W₁)Typically, for screw impellers, a flow rate is assigned within

. The indicated (optimal) limits of variation of the flow coefficient for the axial screw impeller can be achieved only by increasing the inlet area by increasing the outer diameter D ₁ , and this is also accompanied by an increase in peripheral speed (U ₁ ), which leads to an increase in the relative speed (W ₁ )

W ₁ ² = U ₁ ² + C _1z ²

at the entrance of the fluid flow into the impeller and to increase the cavitation reserve (Δh _cr ), as follows from its expression through the indicated parameters (forla TIGM - see [4] AA Lomakin. Centrifugal and axial pumps. Mechanical engineering. Moscow. 1966. Leningrad, p. 195):

where C _1z is the average absolute speed at the entrance of the fluid flow into the impeller,

W ₁ - average relative speed at the entrance of the fluid flow into the impeller,

на уровне λ_кр≤0,025 имеет линейный характер и определяется по формуле (см. [3], стр. 195):n = λ _cr - cavitation coefficient (critical value), usually achieved for the most successful screw wheels with

at the level λ _cr ≤0.025 is linear and is determined by the formula (see [3], p. 195):

, а следовательно, опуститься (благодаря линейной зависимости коэффициента кавитации от коэффициента расхода) в область еще более низких его значений:Such a contradictory ratio of parameters is characteristic of axial impellers, and for a centripetal impeller this contradiction is eliminated, because decrease of the flow velocity at the inlet of the centripetal impeller is possible to provide, without increasing the outer diameter D ₁ and a corresponding increase of the relative velocity W _1, and only by increasing the width of the centripetal impeller inlet (A _1rk), and thereby achieve even lower values flow rate

and, therefore, to descend (due to the linear dependence of the cavitation coefficient on the flow coefficient) to the region of its even lower values:

λ _cr ≤0.022, as required.

It is natural to assume that when profiling a centripetal impeller, other known features of profiling of auger impellers will be taken into account, such as: thickness of inlet edges, dressing (backing) of the input edges of the blades on the suction side and other profiling features known from literature that are necessary to achieve increased anti-cavitation qualities. The difference is only in the choice of the density of the blades, the value of which is recommended to be chosen, as is done when designing centrifugal radial impellers, too, for reasons of the adequacy of the effect of radial blades on the swirl of the fluid flow (C _and ):

τ _opt = b / t _cp = (l, 6-1.8),

where b is the length of the chord (see Fig. 2),

t _cf - the average circumferential pitch of the blades (t _cp = πD _cf / Z _l ),

D _cp = (D _{1 pk} + D _{2 pk} ) / 2,

Z _l - the number of blades.

. Безусловно, для этого потребуются дальнейшие исследования центростремительных рабочих колес насоса. Представляет интерес для насосостроения также исследование его антикавитационных качеств в области сверхнизких значений коэффициентов расхода

, что достижимо сделать именно для центростремительных рабочих колес.Thus, we can conclude: the design of a centripetal vane pump is proposed as the most effective among various types of vane pumps to achieve a positive effect of an additional increase in pressure and efficiency in the presence of convection from the periphery to the center of the centripetal impeller when pumping a heated liquid of nonuniform density, taking into account their thermophysical properties of the flow in the field of action of the acceleration of rotation, and, in particular, for a possible increase in its intensity with increased values of the parameter of the multiplicity of acceleration of rotation in the range of values:

. Of course, this will require further studies of the centripetal impellers of the pump. It is of interest to pump engineering to also study its anticavitation qualities in the field of ultra-low values of flow coefficients

, which is achievable precisely for centripetal impellers.

Claims (1)

A centripetal vane pump for pumping a heated liquid of non-uniform density, containing an inlet pipe with an inner diameter (d ₀ ) and an inlet area of the pump (F ₀ ), a centripetal impeller consisting of a drive and cover discs, radial blades bent along an arc of a circle of radius (r _l ) with the bending center O ₂ offset from the center of rotation O ₁ and the origin of the orthogonal coordinate system aligned with it radially along the Y axis by a distance Y _о and along the X axis in the direction of rotation from the Y axis at a distance X _about , and a conical sleeve with a decreasing in length (L _W ) diameter from the initial size (d _1B = 0.8-0.9) D _2rk to the final size (d _2B = 0.1-0.2) D _2rk , inlet, outlet and shaft, which rotates the centripetal impeller from the drive, which can be used asynchronous electric motor, characterized in that the centripetal impeller has an input section (1 pk) located on the periphery on the outer diameter D _{1 pk} , with width centripetal impeller inlet (A _1rk) to form a cylindrical tso area (F _1rk) and outlet section (2rk) disposed in the central portion on the inner diameter D _2rk = (0,3-0,7) D _1rk, and a width of the centripetal impeller at the outlet (A _2rk) forms a cylindrical a ring with an area (F _{2 pk} ), and the supply is located on the outside, opposite the inlet section (1 pk) of the centripetal impeller, and is made in the form of a spiral channel with a constant circumferential width (A _ck = const) equal to the width of the centripetal impeller at the entrance (A _ck = A _{1 pk} ), and the height measured from the outside the diameter of the centripetal impeller D ₁ to the inner surface of the spiral channel wall, decreasing in proportion to the angle of coverage (ϕ _ohv ), with a height (h _ck ) counted from the initial section height (h _1sk ) adjacent to the inlet pipe and taken as the reference angle (φ _OHV = 0 °), the corresponding input area of the pump (h _1SK = F ₀ / a _sa) to a final height in one current detection angle turnover (φ _OHV = 360 °), equal to the radial clearance (h _2sk = δ _r), which provides free rotation to the working centripetal ENA and outlet located on the outer side, opposite the outlet section (2rk) centripetal impeller is formed as an annular diffuser, formed by the inner surface of outlet wall with an inner diameter (d ₂₎ and the outer surface of the conical sleeve centripetal impeller exit area of pump (F ₂ ).

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