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Power seals and shutdown seals for pumps: The power seals and shutdown seals for auxiliary impellers can overcome some of the shortcomings of packing and mechanical seals; they feature a simple structure, reliable sealing performance, and no leakage even when a small amount of fluid is present. They are now widely used in the metallurgical mining and petrochemical industries. Dynamic seal of the auxiliary impeller: The dynamic seal of the auxiliary impeller (also known as centrifugal seal, hydrodynamic seal, etc.) can be divided into blade seal (Figure 1) and auxiliary impeller seal. As the pump rotates, the impeller generates liquid at pressure P, which flows toward the outlet while also pressing against the packing chamber, causing the fluid to leak outward. However, due to the action of the secondary vanes and secondary impeller, it generates a centrifugal force P1′ or P2′. Its direction is opposite to the direction in which the impeller generates pressure P. Thus, the leaked liquid can be pushed back. That is, an “isobaric seal” or “negative pressure seal” is created within the sealed chamber, ensuring that the pump remains leak-free during operation. Thanks to the parking seal device, the pump remains sealed even when it is stopped. 1. Calculation of the secondary vane seal: Several open radial ribs are provided on the plane of the impeller rear cover plate; these constitute the secondary vanes. It remains at a very small gap from the pump casing, and the liquid rotates at an angular velocity similar to that of the impeller, rather than at half that angular velocity as in the case when there are no auxiliary vanes. This reduces the pressure exerted by the liquid at the packing chamber. According to the formula: KF = EF – EK = {–} (1), there is a certain gap between the secondary blades and the pump casing. In the gap, the angular velocity of the liquid is less than the angular velocity ω of the impeller, but greater than ω/2. Stepanov believes that this angular velocity can be approximated as: ω′ = ω(1/t)/2. Here, ω′ is the angular velocity of the liquid in the gap between the rear cover of the working wheel and the pump casing; ω is the angular velocity of the working wheel; S is the distance between the pump casing and the secondary vanes on the rear cover of the impeller; and t is the average height of the secondary vanes. From this, an expression for the liquid pressure EK in front of the packing chamber can be derived: HBr = H – 1/285((n/1000)²){D²² – DR²/(DR² – Db²)}. Here, HBr is the head pressure after pressure reduction by the secondary vanes (in meters of water column); H₂ = H – V³²/(2g), where V₃ = KV₃√2Gh, and V₃ represents the average flow velocity inside the volute. During calculations, it can be assumed in advance that there is an isobaric seal, i.e., HBr = 0 ； In the case of negative pressure sealing, by using a negative value for HBr in equation (2), the outer diameter DR of the secondary blade can be calculated. If the calculated DR is greater than the D2 value, then the secondary impeller sealing structure needs to be considered. During calculation, t can be selected in advance. It is generally set between 0.5 and 1 centimeter. s-t represents the gap between the secondary vane and the pump casing, and its value must be ensured by the machining precision. The smaller the gap, the greater the balancing capability, but the requirements for machining and assembly are higher. When the precision of the parts related to this clearance is at grade 4–6, s-t is generally set at 0.03–0.3 cm (a smaller value is used for small pumps). Furthermore, the DR value obtained from this calculation tends to be high. The correct value should also be determined after being adjusted through experiments. 2. Calculation of the dynamic seal for the auxiliary impeller: The dynamic seal for the auxiliary impeller offers specific advantages in the petrochemical industry and in the electroplating industry for transporting special fluids. Due to the needs of our work on developing rubber-lined pumps, we have conducted some experimental research in this area. Certain results have been achieved in alternative packing seals and mechanical seals. A secondary impeller dynamic seal is well utilized in rubber-lined pumps. Subsequently, mechanical seals were also replaced in F-type corrosion-resistant pumps. Practice has shown that dynamic sealing has great prospects for development. When the pump is in operation, it is assumed that the liquid ω_liquid in the clearance δZ of the secondary impeller cavity rotates at ψω (Figure 4); due to the presence of this clearance, the angular velocity of the liquid ω_liquid is clearly less than the angular velocity W of the working wheel. This ratio is denoted by ψ. Then, the pressure difference at any radius r_on the outer circumference of the auxiliary impeller is: P2_on = γ/2g (U22_on – U2_liquid) = γ/g (ψ2ω2(r22_on – r2_liquid))/2. (3) When the gap Δ between the guide vanes and the auxiliary impeller is large, it can be assumed that P2 = P_high. When the gap Δ is small, it can be assumed that P2 = P_high × ω²/8g (r22 – r12). Assuming that the radius of the liquid on the low-pressure side is r = r_liquid, the maximum pressure difference generated by the impeller is ΔPmax = Cγ/8gω² (D22 – D12). If expressed in terms of head, then HP = ΔPmax/γ = C/8g (nπ/30)² (D22 – D12). By substituting g = 980 cm/s², this expression can be simplified to: HP = C/71.6 (n/1000)² (D22 – D12). Here, C is the backpressure coefficient, which is determined by the blade height h and the gap δZ. An exact solution has not yet been able to be obtained analytically; it can only be given experimentally. Based on the results of various experiments, and for the sake of simplifying calculations, the following values can be adopted during design: when δZ is greater than 3 millimeters, take C=0.75–0.8; when δZ is less than 3 millimeters, take C=0.85–0.9. On the smooth surface of the auxiliary impeller, a pressure of Hs is also generated, directed in the opposite direction to the pressure exerted by the auxiliary impeller. Hs = Cs × Hts (6) Where Hs is the pressure increase on the smooth surface of the auxiliary impeller (in meters); Cs is the smooth-surface coefficient, typically taken as Cs = 0.1; Hts is the theoretical pressure increase on the smooth surface of the auxiliary impeller (in meters), with Hts = 1/71.6 × (n/1000)² × (D22² – D12²). The total pressure increase capability of the dynamic seal for the entire auxiliary impeller is given by: H = HBr – HP – Hs. Both the sealing of the auxiliary blades and the auxiliary impeller itself require additional power, which is referred to as extra power. This power is primarily consumed by the frictional losses between the secondary impeller and the liquid. We believe that this power value will not exceed the power consumption caused by the leakage that occurs when a sealing ring is installed on the rear wheel disc. Furthermore, it is also a power constant value that is different from that consumed in the drainage through the flow path unloading holes, and which increases as the wear gap value increases. This power is proportional to the square of the impeller’s outer diameter, and proportional to the average width of the blades. To achieve the same balancing effect, it is often appropriate to reduce the outer diameter of the blade while increasing its width. Additionally, the additional power is also related to the boosting capability. When the negative pressure value is high, the vaporization and decomposition surface rises, the outer diameter of the auxiliary impeller increases, and the power consumption rises as well. To minimize power loss as much as possible, the secondary pressure value should not be set too high.