The flow behavior inside a miniature centrifugal pump, under a periodic pulse flow rate, was studied by means of numerical simulation. For a given incoming periodic pulse flow with a sine wave, the performance of the centrifugal pump was investigated in the section with increasing flow and the section with decreasing flow, and the special points of the flow rate and the periodic flow were identified. Further, the energy gradient method and the Q-criterion were adopted to analyze the internal vertical structure and flow stability. It was found that the regions with large variations in velocity and total pressure were mainly located at the leading edge of the suction surface and the middle area of the pressure surface of the blades. Irregular pressure fluctuation frequency under the periodic pulse flow was shown; this was mainly concentrated in the low-frequency zones close to the impeller’s rotational frequency. In addition, for the same flow rate in the periodic pulse flow, the pressure frequency fluctuation for the increasing flow rate section was higher than that observed for the decreasing flow rate section. It was found that the most unstable sections appeared in the first half-period of the flow rate variation (large flow rate), according to the distributions of the Q criteria of the vortex and the energy gradient function K. In this section, motions of strong vortices led to large gradients of the mechanical energy.

The effect of inlet pre-swirl on the performance of a centrifugal pump is studied by numerical simulation. The governing equations are Navier–Stokes equations and the shear stress transport k–ω turbulence model. The numerical results show that the optimal operating point moves from the low flow region to the high flow region as the pre-swirl angle shifts from positive to negative. It is found by contours of Omega–Liutex that the positive pre-swirl angle is able to weaken the vortex on the blade suction and reduce the energy dissipation. On both the 0.5Q₀ and 1.2Q₀ operating conditions, the proportion of entropy production loss in the impeller and volute is about 60% and 30%, respectively. As the pre-swirl angle changes from negative to positive, the entropy production loss in the inlet and outlet pipes increases slowly, and the entropy production loss in the volute and impeller shows a decreasing trend and the peak area of entropy loss moves toward the outlet. Under the four pre-swirl angles, the main frequency is always the passing frequency of the blade. The pre-swirl angle affects the pressure fluctuation at the main frequency but has little effect at the secondary frequency. The change in velocity pulsation amplitude in the impeller in the positive pre-swirl angle is smaller than that in the negative pre-swirl angle. As a result, for the positive pre-swirl angle, the turbulent kinetic energy density in the impeller is low, and the energy loss is low, compared with negative pre-swirl. Under the low flow condition (0.5Q₀), the change in velocity pulsation amplitude in the inertial range of the energy spectrum under negative pre-swirl is greater than that under positive pre-swirl.

The performance of a vane diffuser plays an important role in transforming the high kinetic energy into pressure energy at the impeller outlet of a centrifugal compressor. In this study, the performance enhancement of a centrifugal compressor is achieved by designing a tandem cascade diffuser. An optimum value of the total bending angle of a tandem cascade diffuser is obtained through numerical simulation. The total bending angle is defined as the angle between the tangent line of the front blade arc line at the leading edge point and the tangent line of the rear blade arc line at the trailing edge point. The range of the total bending angle increases from zero to 21°. The simulation results show that the variation of the total bending angle has a great impact on the performances of the compressor stage. The best performance is achieved by the 10° model, by which the minimum total pressure loss coefficient and the maximum static pressure recovery coefficient are obtained. The mechanism of performance improvement by choosing a reasonable total bending angle is that the separated flow zone in the diffuser is constrained and the distribution of radial velocity at the inlet and outlet of the diffuser is more uniformed. With the 10° model, the separated zone almost completely disappears in the diffuser and a distribution of more uniform radial velocity along the meridional section is obtained. Compared with the stage of the prototype with a single row of vane diffuser, the stage with the newly designed tandem vane diffuser (10° model) achieved performance improvement, the efficiency increased about 6%, and the total pressure ratio increased about 3.5% at the flow coefficient of 0.15.