Numerical Simulation Of Hydraulic Characteristics Of Submerged Pump And Calculation Of Rotor Critical Speed

In the chemical industry,the steady,safe and efficient operation of the submerged pump plays a key role in the sustainable production.The pump frequently arises some problems,such as low hydraulic efficiency,fluctuation of service,bearing wear and even with broken axis in engineering.Mechanical vibration and hydraulic vibration are the main causes of the unstable operation of the submerged pump.The mechanical reasons include:the foundation plate is not leveling and alignment;the pipeline configuration is unreasonable which result in the stress deformation,the coaxial misalignment,the impeller and pump shaft accuracy,and the uneven quality distribution caused by casting;or it is the failure to achieve the rotation balance,etc.The hydraulic reasons include:the asymmetry of the inlet fluid of the pump,the excessive pressure pulsation of the inlet and outlet,and the excessive pressure difference between the front and rear cover plates of the impeller,etc.Taking the XDB700-27 submerged pump produced by a factory in China as an example,two schemes are redesigned on the basis of the original ones.Original programme:impeller without balance hole but pump body is provided with balance holes.Programme 1:impeller is provided with balance holes.Programme 2:impeller and pump body are all provided with balance holes.The external and axial force curves of the three different schemes are compared by numerical simulation,and the appropriate research plan is obtained.Starting with the hydrodynamic characteristics and the critical rotor speed,the pressure pulsation of the radial force and the outlet of the impeller under the different flow conditions under the appropriate scheme are analyzed.Finally,the critical speed of rotor with different axle diameter under the design flow is calculated.The results of the study are as follows:1.First of all,the external characteristic curves of the three schemes are obtained through steady numerical simulation.The head,efficiency and axis power curve are obtained.Under the designed flow,the head of the original scheme is 28.9m,the scheme 1 is 27.7m and the second is 28.3m.Under the same flow rate,the hydraulic efficiency of the original impeller is the largest,and the scheme 2 is lower,and the scheme 1 is smallest.Under the design flow,the scheme 1&2 are reduced by 5.8%and2.9%compared with the original scheme respectively.The axial power of the three schemes is not changed as a whole,and the change of balance hole position has no effect on shaft power.2.Under the three schemes,the axial force of the submerged pump decreases with the increase of the flow rate.Under the same flow,the axial force of the original scheme is the largest,and the scheme 2 is the smallest.Near the design flow point,the axial force of the original impeller is 10003.1N,the axial force of the scheme one is5812.1N,it decreases by 41.9%,the scheme two is 5591.8N,and it decreases by44.1%.Obviously,compared with the other,the scheme 1&2 can better balance the axial force of the impeller.Compared with the three kinds of research schemes,Under the condition that both the impeller and the guide bearing seat are openings,the axial force of the submerged pump is smaller and the efficiency is higher.Therefore,scheme2 is more suitable for the structural design of the pumps.3.Taking the constant calculation results as the initial value of the unsteady calculation,the unsteady calculation of 0.6Q_d,1.0Q_d and 1.4Q_d is carried out by Fluent.Then the Fourier transform is used to obtain the frequency domain of the radial force at each working condition.Under the designed flow,the peak radial force is the smallest and the obvious periodic variation with the time presentation.In one circle,the radial force has 6 wave peaks and 6 wave valleys and the angle of the adjacent wave peak is 60 degree.The radial force is larger in the off designed condition,and there is little periodic change.The reason is that radial force is less affected by the static and dynamic interference under the partial condition.The radial force pulsation is mainly 6 times of the frequency of the impeller rotation,that is to say,the main frequency of the radial force Fourier transform equals to the blade frequency.4.By monitoring the pressure fluctuation at the impeller outlet 2mm,we get the time domain chart and the frequency chart.The static pressure of each monitoring point shows an obvious periodicity and the monitoring pressure fluctuates on the250Kpa.The FFT transform frequency is an integer frequency of the shaft frequency,and the main frequency is 6 times of the shaft frequency.In other words,the pulsating main frequency is equal to the blade frequency.The main frequency amplitude at P₁&P₄ is higher than the other points,which is approximately equal.The maximum pulsation amplitude of P₂,P₃,P₅ and P₆ are approximately equal.This is because the points of P₁&P₄ are respectively located at the tongue of the volute and the partition tongue,and the static and dynamic interference are strong between the impeller and the tongue,which leads the amplitude of the main frequency and the maximum amplitude of this position higher.5.Using the critical speed of rotor in dry and wet environment calculated by ANSYS workbench,we analyzed the critical speed of rotor with different shaft diameters.When the shaft diameter is 90mm,the first two critical speeds of the rotor are approximately 1838r/min under the wet environment,and the critical speed of the third or fourth order is 3306r/min,so that the resonance phenomenon is avoided.Under the condition that the working strength of the rotor and the length of the rotor are constant,the axial diameter of 90mm is more suitable for engineering application.