CC BY
105
UDC 621.454.2
Doi: 10.31772/2587-6066-2019-20-2-219-227
For citation: Zuev A. A., Nazarov V. P., Arngold A. A., Petrov I. M. [The method of the disk friction determining of low mass flow centrifugal pumps]. Siberian Journal of Science and Technology. 2019, Vol. 20, No. 2, P. 219-227. Doi: 10.31772/2587-6066-2019-20-2-219-227
Для цитирования: Зуев А. А., Назаров В. П., Арнгольд А. А., Петров И. М. Методика определения дискового трения малорасходных центробежных насосов // Сибирский журнал науки и технологий. 2019. Т. 20, № 2. С. 219-227. Doi: 10.31772/2587-6066-2019-20-2-219-227
THE METHOD OF THE DISK FRICTION DETERMINING OF LOW MASS FLOW CENTRIFUGAL PUMPS
A. A. Zuev1, V. P. Nazarov1, A. A. Arngold2, I. M. Petrov2
1 Reshetnev Siberian State University of Science and Technology 31, Krasnoyarsky Rabochy Av., Krasnoyarsk, 660037, Russian Federation
2JSC "Krasnoyarsk machine-building plant" 29, Krasnoyarsky Rabochy Av., Krasnoyarsk, 660123, Russian Federation E-mail: [email protected]
Low mass flow centrifugal pumps are currently widely used in the energy supply system of liquid rocket engines, the engines of correction, docks, consisting of on-Board power sources on-Board sources power supply system of fuel components in the in gas generator systems for inflating fuel tanks, and in temperature control systems of aircraft and spacecraft.
When designing low mass flow centrifugal pumps for aerospace purposes, methods for calculating and optimizing the flow rate are often used corresponding to the design methods of full-size centrifugal pumps, which limits the mode and design potential of pumps and affects their energy characteristics and reliability. Reliability requirements often lead to the need to reserve units and fuel-supply systems.
Despite the large amount of research works, the issues of reliable design of low mass flow centrifugal pumps with high energy and operational parameters for spacecraft and aircraft remains an urgent task.
The article analyses the operational parameters of low mass flow centrifugal pumps used in aircraft and spacecraft power systems. Taking into account working fluid used and the temperature range, it was found that a laminar rotational flow with Reynolds number characteristic Re = 103 + 3 -105 is realized in the lateral cavity between the impeller and the pump housing.
The determination of power losses on disk friction of the impeller technique is developed taking into account design features and the applied schemes. Equations for determining the disk friction coefficients are consistent with the dependencies obtained by other authors. The obtained equations for the laminar rotational flow made it possible to determine the dependences for the resistance moment and the disk friction power of the impeller determining of a low mass flow centrifugal pump.
Keywords: disk friction, power balance, low mass flow centrifugal pump, dynamic spatial boundary layer.
МЕТОДИКА ОПРЕДЕЛЕНИЯ ДИСКОВОГО ТРЕНИЯ МАЛОРАСХОДНЫХ ЦЕНТРОБЕЖНЫХ НАСОСОВ
А. А. Зуев1, В. П. Назаров1, А. А. Арнгольд2, И. М. Петров2
1 Сибирский государственный университет науки и технологий имени академика М. Ф. Решетнева Российская Федерация, 660037, просп. им. газ. «Красноярский рабочий», 31
2 АО «Красноярский машиностроительный завод» Российская Федерация, 660123, просп. им. газ. «Красноярский рабочий», 29
Е-таП: [email protected]
Малорасходные центробежные насосы в настоящее время находят широкое применение в системах топ-ливоподачи жидкостных ракетных двигателей малой тяги, двигателях коррекции и ориентации космических аппаратов, в составе бортовых источников мощности, газогенераторных системах надува топливных баков, системах терморегулирования летательных и космических аппаратов.
При проектировании малорасходных центробежных насосов аэрокосмического назначения зачастую используют методы расчета и оптимизации проточной части в большей степени соответствующие расчетным методикам полноразмерных центробежных насосов, что ограничивает режимный и конструктивный потенциал насосов и влияет на их энергетические параметры и надежность. Требования обеспечения надежности зачастую приводят к необходимости резервирования агрегатов и систем топливоподачи.
Поэтому, несмотря на большой объем проведенных научно-исследовательских работ, разработка методики проектирования малорасходных центробежных насосов с высокими энергетическими и эксплуатационными параметрами аппаратов остается актуальной задачей ракетного двигателестроения.
В статье проведен анализ режимных параметров малорасходных центробежных насосов, используемых в энергетических системах летательных и космических аппаратов. С учетом используемых рабочих тел и диапазона температурного режима выявлено, что в полости между рабочим колесом и корпусом насоса реализуется ламинарное вращательное течение с характерными числами Рейнольдса в диапазоне Яе = 103 + 3 -105.
С учетом конструктивных особенностей и применяемых схем разработана методика определения потерь мощности на дисковое трение рабочего колеса. Выражения для определения коэффициентов дискового трения согласуются с результатами, полученными другими авторами. Полученные выражения для ламинарного вращательного течения позволили определить математические зависимости для определения момента сопротивления и мощности дискового трения рабочего колеса малорасходного центробежного насоса.
Ключевые слова: дисковое трение, баланс мощности, малорасходный центробежный насос, динамический пространственный пограничный слой.
Introduction. A lot of research works [1-9], including low-flow centrifugal pumps [10-20], are devoted to methods of calculating, modelling, and designing centrifugal pumps for liquid rocket engines (LRE). In [1], an experimental characteristic of the rotational speed effect on the impeller speed is considered. In [4-7], the effect of the blade channel on the centrifugal pump performance is researched. The article [9] is devoted to design methods, as well as to the influence of geometrical and operational parameters on the velocity fields and performance distribution of low mass flow centrifugal pumps.
In research works by A. V. Bobkov [10-14], an analysis of the miniaturization of centrifugal-type superchargers on the kinematic parameters of the working fluid, which allow taking into account the factors of small size of the structure, has been carried out; the possibilities of increasing the efficiency of small-sized centrifugal electric pump units by reducing the rotor friction losses and the possibility of the pressure characteristics increasing were considered. In the works of V. V. Dvirny [15; 16] methods for improving supply units and the need to ensure a long life of low-flow blowers are considered. In the studies of M. V. Kraev and E. M. Kraeva [17-19], methods for calculating the energy parameters of low mass flow pumps units are given, the main operational factors are identified, areas of semi-open rational use and open-type impellers that provide high values of energy parameters low flow systems are defined.
However, despite the large amount of research, the method development for designing low mass flow centrifugal pumps with high energy and operational parameters for spacecraft and aircraft remains an urgent task in the field of rocket engine building.
Design and operating parameters characteristics. Low mass flow centrifugal pumps are currently widely used in the energy supply system of liquid rocket engines, the engines of correction, docks, consisting of on-Board power sources on-Board sources power supply system of fuel components in the in gas generator systems for inflating fuel tanks, and in temperature control systems of aircraft and spacecraft.
Low mass flow centrifugal pumps are characterized by the following parameters:
- working fluid consumption does not exceed V = 300 -10~6 M3/c;
- low discharge coefficient (the absolute velocity ratio of the meridional component to the circumferential component at the exit of the impeller)
— < 0.1; u2
- rotor speed of the pump from 3000 to 10000 rpm;
- impeller diameter does not exceed 0.05 m;
- speed ratio is in the range ns = 40 + 80 [20].
As a rule, an electric drive is used as a drive for low mass flow centrifugal pumps (including thermal control systems (TCS)). The electric drive uses brushless DC motors. The frequency of the drive shaft rotation is characterized by the rotation speed ra=314 ^1047 c-1. The required resource of low-flow pumps with ball-bearing is 40000-155000 hours of operation (from 4.5 to 18 years). To ensure the required resource, the design schemes of low-flow electric pump units (EPU) with supply elements redundancy are being developed [16].
The working fluids of EPU are various technical fluids: water-glycerine solvents, RM distillate oils, LZ-TK-2 coolant, etc. [19]. The temperature range of thermal control systems operability with LZ-TK-2 coolant is from -90 to +60 °C, and for immersed pumps which supply RM distillate oil is from +2 to 220 °C. Due to the wide range of working fluids used and operating temperatures, the kinematic viscosity of the fluids varies within v = 1-10"3 ^0.7 -10~6m2/s.
Setting a research problem. Known methods of designing centrifugal pumps do not provide conditions for the similarity of the processes implemented in the flow parts of the EPU and full-size turbopump units (TPU) of liquid-propellant rocket engines (LRE). In order to increase the reliability of the energy characteristics calculation, a refinement of the used calculation dependencies
and the development of EPU design methods for the considered standard sizes are required.
Further, the EPU power balance will be considered. The output power of a low mass flow centrifugal pump is defined as
NO = Nr e q - Nmech - Nd - Nleak - Nhydr,
where Nreq - required power, Nmech - mechanical power loss, NM - impeller disk friction power, Nleak - leakage power of the working fluid through the seals and Nhydr -
hydrodynamic power loss.
It is important to note that disk losses for low mass flow centrifugal pumps can reach 10% and depend on the type of working fluid and operating temperatures significantly.
Further, in the article, the methodology for determining the ratio of disk loss and power loss to disk friction and the EPU disk resistance moment will be considered.
Disk friction coefficient determination. The turbulent flow regime between the rotating impeller disk and the pump housing meets Reynolds criterion Re = 5 -105, for laminar mode Re<105. The Reynolds criterion for the lateral cavity between the impeller and the pump housing is defined as
raD2 Re =-2
4v
Taking into account the geometrical and regime parameters and working fluid physicomechanical characteristics for thermal control systems, the range for the Reynolds criterion will be defined as Re = 103 ^3-105, which corresponds to the laminar flow regime.
In the implementing laminar flow process the velocity distribution in the dynamic boundary layer is determined as
u = 1 -f! -1
U Is for laminar flow, the profile degree is m = 2 ^ 5.
Fig. 1 shows the distribution of laminar flow velocity profiles.
Fig. 2 shows a photograph of the laminar boundary layer. From fig. 1 and 2 it can be concluded that the presented velocity distribution function in the laminar boundary layer agrees well with experiment.
Depending on the degree of the velocity profile m there is a need to redefine the dependence of the friction induced shear stress near the wall surface in the boundary conditions of the laminar boundary layer x0. The equation for the law of friction of the gradient profile of the distribution of the velocity component in the boundary layer for laminar flow is written as
pU2
- = 0.293
/ ** \-0-5 ' U s ^
v
v у
When considering the pulse thickness loss equation for the gradient velocity distribution profile
mS
S =-,
(m + 1)(2m +1)
the friction induced shear stress for a rectilinear uniform flow is defined as
\-0.5
x0 = 0.293pUz
U
mS
v (m + 1)(2m +1)
x
0
u/U
0 0,1 0.2 0,3 0,4 0,5 0,6 0,7 0,Я 0,9 I
Fig. 1. Distribution of laminar flow velocity profiles
Рис. 1. Распределение профилей скорости ламинарного течения потока
Fig. 2. The velocity distribution profile of the laminar boundary layer
Рис. 2. Профиль распределения скорости ламинарного пограничного слоя
When considering the case of the working fluid flow in the channel between the rotating disk and the fixed wall, it is necessary to take into account the stream core angular velocity and the disk. If the circumferential component of the absolute flow velocity on the wall is
Uw =®CR,
then the circumferential friction stress on the wall is defined as
T0wa= 0.293PKR)2 ^8: N
If the circumferential component of the flow velocity on the disk is
Ud = (raD) R , then the circumferential friction stress on the disk will be equal to
xD:= 0.293p[(»d)R]2pH*-^^ ^
The considered case refers to the flow at which the distribution of the circumferential velocity component
corresponds to the equation R = ra = const.
The thickness of the pulse loss in the circumferential direction with an arbitrary profile degree for the laminar flow wall can be defined as
S„„, = D1
1
Re5
4
R 5.
where
D1 = 0.04535
8m ( 47m 2 + 12m +1) (3m +1)2 (5m +1) 14m (m + 1)(7032m4 + 2602m3 + 413m2 + 32m +1) (3m + 1)(4m + 1)(5m + 1)(6m + 1)(7m + 1)(8m +1)
'2m +1 ^ „ +1
2 (3m + 1)(4m + 1)(5m +1) (m + 1)(47m +1) (3m + 1))4m + 1))5m + 1))6m + 1))7m + 1))8m +1) 2m (m + 1)(7032m4 + 2602m3 + 413m2 + 32m +1)
The thickness of the momentum loss in the circumferential direction with an arbitrary degree of profile for a disk with a laminar flow is
= 0.04535
f 4M 2-7LЛ 1 + H
1 J + L
0.2 R0'6 .
After transforming this equation, the thickness of the loss of momentum in the circumferential direction on the wall for laminar flow is obtained
8aD =
0.01256
I1+H)J
'3LJ + 4 L (K - 2 J )
_2 1 J + L
\ 0.2
R0.6.
x
Further, the thickness of the momentum loss in the circumferential direction on the disk is obtained as
4
-R5,
SaD = D 2-
1
1
ReD-
where
D2 -
0.0185
"2(3m+1)(4m+1)(5m+1) (m+1)(47m +1) (3m+1)(4m+1)(5m+1)(6m+1)(7m+1)(8m+1) 2m(m+1)(7032m4 + 2602m3 + 413m2 + 32m+1
(m+1)
2m+1
+1 I (47m+1)
8m(m+1) [ (m+1)(47m+1) 2m(47m2 + 12m+1 j (3m+1)(5 m+1) [ (4m+1) (3m+1) J*
x (7032m4 + 2602m3 + 413m2 + 32m+1)
(6 m+1)(7 m +1)(8 m+1) 6m(m+1)2(47m+1)(7032m4 + 2602m3 + 413m2 + 32m+1 (3m+1) (4 m+1)(5m+1)(6 m+1)(7 m+1) (8 m+1)
The velocity distribution profile m for practically important cases (which are realized with laminar flow in the cavity between the impeller and the pump) is summarized in tab. 1.
Taking into account the obtained equations for the impulse loss thickness in the circumferential direction, the tangential friction stresses on the wall for laminar flow are defined as
8
w = 0.293pv2Ref
T0a = 1 ,
(D1)2 r2 and the laminar disk is defined as
Qa '
0.293pv2ReD-c
(D2)2 R2
The velocity profile distribution m for practically important cases is summarized in tab. 2 for the frictional stresses of the laminar flow.
The friction coefficient equation for the wall and disk of the impeller with a laminar flow is expressed as
CA =
xR
p Re2 v2
Then the friction coefficient for the wall in the circumferential direction with a laminar flow is defined as _ 0.293
C fra _ i 2 '
(Dl)2Re5
and the friction coefficient for a disk in the circumferential direction with a laminar flow is
0.293
CD -С fra -
(D2)2 ReD_
Tab. 3 is for coefficient determining of the disk friction and the wall in the circumferential direction with a laminar flow depending on the degree of the velocity distribution profile. This table can be used for engineering calculations convenience m.
The loss coefficient of disk friction during laminar flow is defined as
C _ cw + CD
UM _ ° frafra .
For the working fluid flow case in the lateral cavity between the working disk and the centrifugal pump housing, the angular core velocity is determined as rac _ 0,5raD. Then the loss of disk friction coefficient laminar flow with laminar flow is defined as
f
см
1
(0.5ReD)
Л
0.293 0.293
(D1)2 (D2)2
Tab. 4 is for determining the laminar flow disk friction loss rate and it demonstrates the analysis of the obtained dependence and comparison with obtained results by other authors.
Fig. 3 presents the disk friction coefficient dependence for the laminar flow of the working fluid at Re < 105.
Table 1
Thickness of impulse loss on the wall and disk in the circumferential direction, for practically important cases in laminar flow
8
№ m ^aiv sa;
1 2 4 0.13612 4 1 R Re5 4 0.27349 4 1 R ReD-c
2 3 4 0.20609 4 1 R Re5 4 0.44887 4 1 R ReD-c
3 4 4 0.2708 4 1 R Re5 4 0.6382 ^f 1 R ReD-c
4 5 4 °.332l R 5 Ref 4 0.83828 f 1 R 5 ReD-c
Table 2
Tangential friction stresses on the wall and disk in the circumferential direction, for practically important cases in laminar flow
№ m Tw 0a TD 0a
1 2 8 0.794158pv2 Rec5 R2 8 0.560269pv2 ReD_c R2
2 3 8 0.645415pv2 Rec5 R2 8 0.437328pv2 ReD-5 R2
3 4 8 0.563045pv2 Re5 R2 8 0.366766pv2 ReD_c R2
4 5 8 0.508432pv2 Re5 R2 8 0.320017pv2 ReD-5 R2
Table 3
Friction coefficient on the wall and disk in the circumferential direction, for practically important cases with laminar flow
№ m С fia С D fra
1 2 0.794158 2 Re5 0.560269 2 ReD-5
2 3 0.645415 2 Re5 0.437328 2 ReD-5
3 4 0.563045 2 Re5 0.366766 2 ReD-5
4 5 0.508432 2 Re5 0.320017 2 ReD-5
Table 4
Disk friction coefficient, for practically important cases with laminar flow
№ m G. Shlikhting Cimbus, K. Smiden F. SHul'tc-Grunov С ^М
1 2 3.87 1 Re^ 3.14 R Re z ' R — = 0.02 (radius to gap ratio) z 2.67 Г Re2 1.787844 2 Re5
2 3 1.429221 2 Re5
3 4 1.227351 2 Re5
4 5 1.093553 2 Re5
If m = 2 and the gradient function of laminar flow agrees well with the dependence of G. Shlikhting [21], then the maximum deviation of the disk friction parameter does not exceed 7 %. If Re = 103 the difference is 5 %, and if Re = 105 the difference is 7 %. In general, all dependencies are in the region of the confidence span
defined by various authors and are in the parameter domain for disk friction coefficients from 0.113 to 0.027, depending on the Re criterion (fig. 3). It is important to note that when designing flow parts it is necessary to choose the turbulence transition which depends on the boundary conditions of the flows.
Fig. 4 shows the dependence on the disk friction coefficients, the friction coefficient for the wall in the circumferential direction and the friction coefficient for the disk in the circumferential direction of the laminar flow required for the flow parts design and for the power balance of low mass flow centrifugal pumps determining.
The radial element of the tangential friction stress is formed by the circumferential and consumable components taking into account the bottom lines slope and leaks through the sealing elements.
Radial stress of friction on the wall is
_w _ _w , _w
X0R _ %0Rp + 0Ra
and radial friction stress on the disk is
_D D D
X0R _ 0Rp x0Ra •
Radial stress of friction from the circumferential component is
T0Ra = ST0a '
where e - tangent of the bottom slope angle.
The radial friction stress will be determined depending on the flow rate associated with the amount of leakage through the sealing elements, and the dependence on the structure is similar to straight linear flow
To = 0.01256p V„2 f ^-^-"°'25
0 r fv (m + 1)(m + 2)
where Vp - flow rate.
The one side resistance moment of the low mass flow centrifugal pump working disk is defined as
R 2
MD = 2nj T0a r2dr = СМ PR5ro
5„2 D •
с
M
0.16
0,14
0,12
0,1
0,08
0,06
0.04
0,02
■ 4 ч • ч • ч • ч --G. Shlikhting
. ч . Ч * ч ♦ ч к V V .......Cimbus, К, Smideri -----m=l
4 • ЧЧ Xv. -m=2
— • - m=3 — ■ — m=4 - — F. SHul'tc-Grunov
SO ^ ^ - ^
^ - >- ^ "—_______ ~ ~ - - , _
lgRe
1000 10000 100000 Fig. 3. Coefficient of disk friction laminar flow Рис. 3. Коэффициент дискового трения ламинарного течения
С,
м
0,12 0,1 0,08 0,06 0,04 0,02
-Cm m=2
--- Cbody --Cdisk
N. ■ч. \
— — ^ ___
lgRe
1000 10000
Fig. 4. Coefficient of disk friction Рис. 4. Коэффициент дискового трения
100000
Then the impeller disk friction power on one side of the low mass flow centrifugal pump is defined as
NfrD = MD®D •
Conclusion. The analysis of low mass flow centrifugal pump operating parameters used in aircraft and spacecraft power systems was carried out. After analysing the working fluid characteristics and the temperature range, it was revealed that a laminar rotational flow with Reynolds
numbers Re = 103 + 3-105 is conducted in the cavity between the impeller and the pump.
The method for determining the power losses due to disk friction of the impeller with design features and the applied schemes was developed. The equations for disk friction coefficient determining are consistent with the results of other authors. The obtained equations for the laminar rotational flow established mathematical dependencies to determine the resistance moment and power of the impeller disk friction of a low mass flow centrifugal pump.
References
1. Valentini, D., Pace, G., Pasini, A., Torre, L., Ha-davandi, R., & d' Agostino, L. (2017). Fluid-induced ro-tordynamic forces on a whirling centrifugal pump. European Journal of Mechanics - B/Fluids, 61, Р. 336-345. Doi: 10.1016/j.euromechflu.2016.09.004.
2. Liu, M., Tan, L., & Cao, S. (2019). Theoretical model of energy performance prediction and BEP determination for centrifugal pump as turbine. Energy. Doi: 10.1016/j.energy.2019.01.162.
3. Sengpanich, K., Bohez, E. L. J., Thongkruer, P., & Sakulphan, K. (2019). New mode to operate centrifugal pump as impulse turbine. Renewable Energy. Doi: 10.1016/j.renene.2019.03.116.
4. Ding, H., Li, Z., Gong, X., & Li, M. (2018). The influence of blade outlet angle on the performance of centrifugal pump with high specific speed. Vacuum. Doi: 10.1016/j.vacuum.2018.10.049.
5. Skrzypacz, J., & Bieganowski, M. (2018). The influence of micro grooves on the parameters of the centrifugal pump impeller. International Journal of Mechanical Sciences, 144, Р. 827-835. Doi: 10.1016/j.ijmecsci.2017.01.039.
6. Zhang, N., Liu, X., Gao, B., Wang, X., & Xia, B. (2019). Effects of modifying the blade trailing edge profile on unsteady pressure pulsations and flow structures in a centrifugal pump. International Journal of Heat and Fluid Flow, 75, Р. 227-238. Doi: 10.1016/j.ijheatfluidflow.2019.01.009.
7. Chen, H., He, J., & Liu, C. (2017). Design and experiment of the centrifugal pump impellers with twisted inlet vice blades. Journal of Hydrodynamics, Ser. B, 29(6), Р. 1085-1088. Doi: 10.1016/s1001-6058(16)60822-3.
8. Yu, R., & Liu, J. (2018). Failure analysis of centrifugal pump impeller. Engineering Failure Analysis, 92, Р. 343-349. Doi: 10.1016/j.engfailanal.2018.06.003.
9. Salehi, S., Raisee, M., J. Cervantes, M., & Nour-bakhsh, A. (2018). On the flow field and performance of a centrifugal pump under operational and geometrical uncertainties. Applied Mathematical Modelling, 61, Р. 540-560. Doi: 10.1016/j.apm.2018.05.008.
10. Bobkov A. V. [The influence of the factor malorazmernyj to the design of the supercharger centrifugal type]. Uspekhi sovremennoy nauki. 2017. Vol. 4. No. 2. P. 127-130 (In Russ.).
11. Bobkov A. V. [A geometric criterion of optimizing the design of blade machines]. Mezhdunarodnyy zhurnal prikladnykh i fundamentalnykh issledovaniy. 2013. No. 6. P. 49 (In Russ.).
12. Bobkov A. V. [Energy expediency of increasing the number of stages in a small centrifugal electric pump unit]. Sovremennyye problemy nauki i obrazovaniya. 2012. No. 4. P. 84 (In Russ.).
13. Bobkov A. V., Tsvetkov E. O. [Improving the pressure qualities of the centrifugal pump thermal control system]. Mezhdunarodnyy zhurnal prikladnykh i fundamentalnykh issledovaniy. 2012. No. 10. P. 110 (In Russ.).
14. Bobkov A. V., Tsvetkov E. O. [Features of balance of power losses in electric pump units of spacecraft thermal control systems]. Izvestiya Samarskogo nauch-nogo tsentra Rossiyskoy akademii nauk. 2011. Vol. 13. No. 1-2. С. 290-292 (In Russ.).
15. Dvirnyy V. V., Krushenko G. G., Golovanova V. V., Dvirnyy G. V., Petyayeva N. N., Kirianova K. A. [Improvement of units for heat transportation in spacecraft]. Issledovaniya naukograda. 2016. No. 3-4 (18). P. 12-16 (In Russ.).
16. Dvirnyy V. V., Dvirnyy G. V., Khnykin A. V., Golovanova V. V., Krushenko G. G. [Providing long lasting resource of economical compressors]. Issledovaniya naukograda. 2014. No. 3 (9). P. 12-20 (In Russ.).
17. Krayeva E. M. [Energy parameters of high-speed low-flow pumps]. Vestnik Moskovskogo aviatsionnogo instituta. 2011. Vol. 18. No. 3. P. 104 (In Russ.).
18. Krayeva E. M. [Calculation of energy parameters of high-speed centrifugal pumps of low speed]. Izvestiya vysshikh uchebnykh zavedeniy. Aviatsionnaya tekhnika. 2010. No. 1. P. 48-50 (In Russ.).
19. Krayeva E. M. [To calculation of operating parameters of centrifugal pumps of low speed]. Vestnik SibCAU. 2009. No. 2 (23). P. 168-170 (In Russ.).
20. Bobkov A. V., Katalazhnova N. I., Kachalov A. A. [TCalculation of field lines ratio specific speed centrifugal blower spacecraft]. Uspekhi sovremennogo estest-voznaniya. 2004. No. 4. P. 50-51 (In Russ.).
21. Shlikhting G. Teoriya pogranichnogo sloya [The theory of the boundary layer]. Moscow, Science Publ., 1974. 712 p.
Библиографические ссылки
1. Fluid-induced rotordynamic forces on a whirling centrifugal pump / D. Valentini, G. Pace,. Pasini A. [et al.] // European Journal of Mechanics - B/Fluids. 2017. Vol. 61. P. 336-345. Doi: 10.1016/j.euromechflu.2016.09.004.
2. Liu M., Tan L., Cao S. Theoretical model of energy performance prediction and BEP determination for centrifugal pump as turbine // Energy. 2019. Vol. 172. P. 712-732. Doi:10.1016/j.energy.2019.01.162.
3. New mode to operate centrifugal pump as impulse turbine / K. Sengpanich, E. L. J. Bohez, P. Thongkruer, K. Sakulphan // Renewable Energy. 2019. Vol. 140. P. 983-993. Doi: 10.1016/j.renene.2019.03.116.
4. The influence of blade outlet angle on the performance of centrifugal pump with high specific speed / H. Ding, Z. Li, X. Gong, M. Li // Vacuum. 2019. Vol. 159. P. 239-246. Doi: 10.1016/j.vacuum.2018.10.049.
5. Skrzypacz J., Bieganowski M. The influence of micro grooves on the parameters of the centrifugal pump impeller // International Journal of Mechanical Sciences. 2018. Vol. 144. P. 827-835. Doi: 10.1016/j.ijmecsci.2017.01.039.
6. Effects of modifying the blade trailing edge profile on unsteady pressure pulsations and flow structures in a centrifugal pump / N. Zhang, X. Liu, B. Gao [et al.] // International Journal of Heat and Fluid Flow. 2019. Vol. 75. P. 227-238. Doi:10.1016/j.ijheatfluidflow.2019.01.009.
7. Chen H., He J., Liu C. Design and experiment of the centrifugal pump impellers with twisted inlet vice blades // Journal of Hydrodynamics. Ser. B. 2017. No. 29(6). P. 1085-1088. Doi: 10.1016/s1001-6058(16)60822-3.
8. Yu R., Liu J. Failure analysis of centrifugal pump impeller // Engineering Failure Analysis. 2018. Vol. 92. P. 343-349. Doi: 10.1016/j.engfailanal.2018.06.003.
9. On the flow field and performance of a centrifugal pump under operational and geometrical uncertainties / S. Salehi, M. Raisee, M. J. Cervantes, A. Nourbakhsh // Applied Mathematical Modelling. 2018. Vol. 61. P. 540-560. Doi: 10.1016/j.apm.2018.05.008.
10. Бобков А. В. Влияние фактора малоразмерно-сти на конструкцию нагнетателя центробежного типа // Успехи современной науки. 2017. Т. 4, № 2. С. 127-130.
11. Бобков А. В. Геометрический критерий оптимизации конструкции лопаточной машины // Международный журнал прикладных и фундаментальных исследований. 2013. № 6. С. 49.
12. Бобков А. В. Энергетическая целесообразность увеличения числа ступеней в малоразмерном центробежном электронасосном агрегате // Современные проблемы науки и образования. 2012. № 4. С. 84.
13. Бобков А. В., Цветков Е. О. Повышение напорных качеств центробежного насоса системы терморегулирования // Международный журнал прикладных и фундаментальных исследований. 2012. № 10. С. 110.
14. Бобков А. В., Цветков Е. О. Особенности баланса потерь мощности в электронасосных агрегатах систем терморегулирования космических аппаратов // Известия Самарского науч. центра РАН. 2011. Т. 13, № 1-2. С. 290-292.
15. Совершенствование агрегатов для транспортировки тепла в космических аппаратах / В. В. Двир-ный, Г. Г. Крушенко, В. В. Голованова [и др.] // Исследования наукограда. 2016. № 3-4 (18). С. 12-16.
16. Обеспечение длительного ресурса малорасходных нагнетателей / В. В. Двирный, Г. В. Двирный, А. В. Хныкин [и др.] // Исследования наукограда. 2014. № 3 (9). С. 12-20.
17. Краева Е. М. Энергетические параметры высокооборотных насосов малого расхода // Вестник Московского авиац. ин-та. 2011. Т. 18, № 3. С. 104.
18. Краева Е. М. Расчет энергетических параметров высокооборотных центробежных насосов малой быстроходности // Известия вузов. Авиационная техника. 2010. № 1. С. 48-50.
19. Краева Е. М. К расчету эксплуатационных параметров центробежных насосов малой быстроходности // Вестник СибГАУ. 2009. № 2 (23). С. 168-170.
20. Бобков А. В., Каталажнова Н. И., Качалов А. А. Расчет полей линий уровня коэффициента быстроходности центробежного нагнетателя космического аппарата // Успехи современного естествознания. 2004. № 4. С. 50-51.
21. Шлихтинг Г. Теория пограничного слоя. М. : Наука, 1974. 712 с.
© Zuev A. A., Nazarov V. P., Arngold A. A.,
Petrov I. M., 2019
Zuev Alexander Alexandrovich - Cand. Sc., associate Professor, Department of Aircraft Engines; Reshetnev Siberian State University of Science and Technologies. E-mail: [email protected].
Nazarov Vladimir Pavlovich - Cand. Sc., Professor, Department of Aircraft Engines, Reshetnev Siberian State University of Science and Technologies. E-mail: [email protected].
Arngold Anna Anatolievna - Department of Special Connectors and Instruments, Krasnoyarsk Machine-Building Plant. E-mail: [email protected].
Petrov Ivan Mikhaylovich - Deputy chief designer for engines, propulsion systems and power plants, Krasnoyarsk Machine-Building Plant. E-mail: [email protected].
Зуев Александр Александрович - кандидат технических наук, доцент, доцент кафедры двигателей летательных аппаратов; Сибирский государственный университет науки и технологий имени академика М. Ф. Решетне-ва. Е-mail: [email protected].
Назаров Владимир Павлович - кандидат технических наук, профессор, заведующий кафедрой двигателей летательных аппаратов; Сибирский государственный университет науки и технологий имени академика М. Ф. Решетнева. Е-mail: [email protected].
Арнгольд Анна Анатольевна - начальник бюро спецсоединителей, приборов и пультов аппаратуры; АО «Красноярский машиностроительный завод». E-mail: [email protected].
Петров Иван Михайлович - заместитель главного конструктора по двигателям, двигательным установкам и энергоустановкам; АО «Красноярский машиностроительный завод». E-mail: [email protected].
