The shaft front-positioned tubular pump device has been widely used in practical engineering, but the shaft rear-positioned pump device is rarely used due to its low efficiency. To investigate the effect of the shaft position on the performance of a tubular pump device, the optimized shaft front-positioned and shaft rear-positioned pump devices were compared and studied. Both tubular pump devices adopt a TJ04-ZL-06 pump model. Three dimensional steady and unsteady numerical simulations combined with model tests were used to compare the difference of two pump devices. Meanwhile, three groups of pressure monitoring points were set at different positions of the pump device to collect pressure information and pressure pulsation was analyzed. The results show that, the highest efficiency of the shaft front and rear positioned pump device are 81.78% and 80.26%, respectively. The hydraulic performance of the two inlet passages is excellent, and the hydraulic loss is close to each other. Therefore, the hydraulic performance of the pump device depends mainly on the hydraulic performance of the outlet passage. The shaft is set in the outlet passage, which will increase the hydraulic loss and reduce efficiency. Under design conditions, the pressure pulsation amplitude at the impeller inlet is the largest, and the pressure pulsation amplitude increases from the hub to the shroud. The pressure pulsation amplitude of the shaft rear-positioned pump device is larger than that of the shaft front-positioned pump device. The pressure pulsation at the impeller inlet and outlet is greatly affected by the number of blades, and the main frequency is three times the RF. This study can provide practical and effective guidance for the design and optimization of the shaft front-positioned and rear-positioned tubular pump devices, which has theoretical value and application value.

The unstable flow of a shaft tubular pump device (STPD) leads to energy loss, thereby reducing its efficiency. The aim of this study is to investigate the distribution pattern of energy loss in STPDs. This paper reveals that the two components with the highest proportion of energy loss are the impeller and the outlet passage. Furthermore, turbulent entropy production is the primary cause of energy loss. Due to the wall effect, the energy loss in the impeller mainly occurs near the hub and shroud. Additionally, the presence of a tip leakage vortex near the shroud further contributes to the energy loss in the region near the shroud. This results in the energy loss proportion exceeding 40% in the region with a volume fraction of 14% near the shroud. In the outlet passage, the energy loss mainly occurs in the front region, with a volume fraction of 30%, and the energy loss in this part accounts for more than 65%. Finally, this study reveals the locations of the vortex in the STPD under different flow-rate conditions, and when the distribution of energy loss is visualized, it is found that the energy loss occurs high in the vortex regions.

In order to improve the pump device efficiency of the frequent operating condition of the extra-low head pumping station, the energy performance of the front-positioned shaft tubular pump device at two guide vane inlet angles has been researched. Based on the function of the guide vane in the pump device, the guide vane blades are divided into three parts: the inlet section, the middle section, and the outlet section. Combining numerical simulation and model tests, the energy performance of the pump device with the inlet section angle adjusted to 0° and −12° were studied and compared, respectively. The research results indicate that the inlet section angle of the guide vane has a significant effect on the energy performance of the pump device. When the guide vane inlet section is adjusted clockwise, the pump device efficiency of the optimal operating point—while the efficiency of the pump device at a low head and large discharge that deviate from the optimal operating point—will be improved. The farther the working condition deviates from the optimal operating point, the greater the influence. Within the scope of the working conditions studied in this paper, the pump device efficiency of the optimal operating point is reduced by about 2%, and the pump device efficiency in the low head and high flow conditions is increased by 5% at the maximum. Adjusting the inlet section angle of the guide vane, the flow pattern in the guide vane will be improved, and the hydraulic loss of the guide vane will be decreased, thus the pump device efficiency is increased. The numerical calculation results of the energy performance agree with the model test results; the maximum error of the pump device efficiency is less than 7%. Adjusting the angle of the inlet section of the guide vane has great significance to the hydraulic design and engineering application of the extra-low head pump device.

In engineering, the highest operating head of the pumping station is usually controlled to be slightly lower than the lowest saddle bottom head of the axial‐flow pump. However, in the practical operation, it is found that the highest operating head of the pumping station is obviously lower than the saddle bottom head of the pump device, which leads to the reduction of the operating range of the pumping station. To investigate the difference of lowest saddle bottom head between axial flow pump and axial flow pump device and apply it correctly, the energy performance tests of the TJ04‐ZL‐06 hydraulic model and its corresponding pump device were carried out to obtain the external curves, and numerical simulation was carried out to analyze and compare the internal flow field and pressure distribution. The results show that when the flow rate decreases, the first saddle‐shaped region of the axial‐flow pump and the saddle‐shaped region of the pump device are caused by the decrease of the lift coefficient due to the increase of the attack angle between flow and blade. When the flow rate is less than 0.32Q_d, the influence range of backflow in the inlet pipe is large, which leads to the high‐pressure zone near the wall of the inlet pressure measurement section during the pump performance test, and hence the second saddle‐shaped region of the axial‐flow pump is essentially a measurement illusion. It is suggested that the inlet pressure measurement section should be set at least 4D_p away from the inlet flange of the impeller when testing the performance of the axial‐flow pump under the condition of small flow rate, and the first saddle bottom head of the axial‐flow pump or the saddle bottom head of the corresponding pump device can be considered as the control value of the highest head of the pumping station.

For pumping station projects in plain areas, shaft tubular pump devices are often used due to the low head. In actual operation, the common working range of the pump device is generally 0.8 Q_bep~1.2 Q_bep. Therefore, it is of great significance to study the hydraulic performance and pressure pulsation characteristics of the pump device in the working range. In this study, the hydraulic performance of a shaft pump device was tested by a model test, and the internal flow and pressure pulsation characteristics were analyzed by numerical simulation. The results obtained from the model test and numerical simulation were in general agreement, indicating that the numerical results were reliable. The results show that the inlet passage has a good flow pattern in the working range, which may offer a favorable flow state for the impeller. When Q = 0.8 Q_bep, the flow in the impeller and guide vane was chaotic, the guide vane had a poor adjustment function on flow direction, and the flow in the outlet passage presented in a spiral motion. When Q = 1.0 and 1.2 Q_bep, the flow in the impeller and guide vane was ordered, and the spiral flow in the outlet passage improved. In the working range, the pressure pulsation was similar. The main frequency at the impeller inlet and outlet was consistent with the blade passing frequency. For the same flow rate condition, the amplitude rose from hub to shroud and declined from impeller inlet to outlet. In addition, the amplitude decreased with an increasing flow rate.

This study aimed to reveal the energy loss characteristics of each part of the shaft tubular pump device (STPD) under stall conditions. Numerical simulations were conducted by using the SST k-ω turbulence model with curvature correction, and the reliability of the simulation results was verified by a model test. The entropy production method was used to evaluate and visualize the energy loss. The results show that turbulent entropy production (TEP) is the main source of energy loss in each component of the STPD, and the TEP increases significantly with the deterioration of stall. The energy loss in the impeller is mainly concentrated near the shroud and hub, while in the guide vanes it is mainly concentrated near the shroud and suction surface of the blades. In addition, with the deterioration of stall, the energy loss in the inlet of the impeller and guide vanes increases significantly. Influenced by the backflow from the impeller, there is a significant amount of energy loss at the outlet segment of the inlet passage, and the location of the high-energy-loss region is consistent with the backflow region. Affected by the flow separation vortex at the tail of the guide vanes, the energy loss in the outlet passage is mainly concentrated at the inlet segment.