A multi-impeller centrifugal water pump

Abstract

This invention relates to the field of centrifugal water pumps, and more particularly to a multi-impeller centrifugal water pump, comprising: a pump casing, a main shaft, and at least two axially parallel semi-open centrifugal impellers. The pump casing has an inlet and an outlet, and internally forms a fluid chamber and an independent volute-type flow channel corresponding to each impeller. These flow channels are connected in series via interstage flow channels, with the first stage connected to the inlet and the last stage connected to the outlet. The centrifugal impellers on the main shaft are fixed by a flat key. Each impeller includes a hub and backward-curved arc blades. The blades have a pressure surface, a suction surface, a leading edge, and a trailing edge. The leading edge connects to the outer circumference of the hub, and the trailing edge bends in the opposite direction of rotation. A flow guide baffle is provided between the impellers inside the pump casing, with annular grooves embedded in the edges and bolted in place. A mechanical seal assembly is installed in the central shaft hole. Adjacent centrifugal impellers rotate in opposite directions, forming a spiral flow channel between the blades with an inlet wider than the outlet. This invention improves fluid transport efficiency and stability, optimizes pressure distribution, and reduces flow losses.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of centrifugal water pumps, and particularly relates to a multi-impeller centrifugal water pump. BACKGROUND

[0002] In the design of conventional centrifugal water pumps, when facing the working condition requirements of high lift and large flow, engineers usually adopt two main design schemes: one is to achieve the performance requirements through the single-impeller multi-stage series connection, and the other is to increase the diameter of the impeller. However, both of the two design schemes have obvious technical limitations. As for the single-impeller multi-stage series connection structure, although the lift requirement can be met, the axial size of the pump body is significantly increased, which not only brings many inconveniences to the installation and space layout of the equipment, but also generates a large axial force in the running process, thereby increasing the vibration risk of the unit and affecting the stability and reliability of the equipment operation. On the other hand, the method of simply increasing the diameter of the impeller can improve the flow, but it will cause the flow velocity at the inlet of the impeller to be too high, which is easy to cause cavitation phenomenon. This cavitation effect not only reduces the operating efficiency of the water pump, but also causes erosion damage to the impeller material, thereby shortening the service life of the water pump.

[0003] In addition, in the conventional design of centrifugal pumps, the impeller blades often adopt a relatively simple straight plate type or a basic circular arc structure. This relatively rough blade profile design will lead to an undesirable flow state of the fluid in the blade flow passage, which is specifically manifested as follows: the fluid is easy to separate from the boundary layer when flowing through the blade surface, and complex secondary flow and vortex phenomena are formed in the flow passage. These undesirable flow characteristics will significantly increase the hydraulic loss, cause low energy conversion efficiency, and seriously restrict the performance improvement space of the water pump. Especially under high load working conditions, these flow problems will become more prominent, further reducing the operating performance and service life of the water pump. SUMMARY

[0004] Therefore, the present application provides a multi-impeller centrifugal water pump to overcome the problem that the fluid is easy to separate from the boundary layer and vortex when flowing in the blade flow passage in the prior art.

[0005] To achieve the above-mentioned purpose, the present application provides a multi-impeller centrifugal water pump, comprising:

[0006] A pump shell is provided with a water inlet and a water outlet, and a fluid chamber is arranged between the water inlet and the water outlet;

[0007] A main shaft is provided with a centrifugal impeller arranged in the axial direction, the centrifugal impeller comprises an impeller hub and a plurality of blades, and the centrifugal impeller is rotationally arranged in the fluid chamber and fixedly connected with the main shaft through a flat key;

[0008] The blade is a back-bent circular arc, each blade includes a pressure surface, a suction surface, a leading edge and a trailing edge; the leading edge of the blade is connected with the outer circle of the hub, and the trailing edge is bent towards the opposite direction of the rotation direction of the impeller;

[0009] Three front pressure convexes are arranged on the pressure surface, and one rear pressure convex is arranged on the suction surface;

[0010] The center line of each front pressure convex is a first curve segment, a second curve segment and a third curve segment respectively;

[0011] The first curve segment extends from the leading edge starting point A to the intermediate point B, and the profile fitting formula is:

[0012]

[0013] The second curve segment extends from the intermediate point B to the transition point C, and the profile fitting formula is:

[0014]

[0015] The third curve segment extends from the transition point C to the trailing edge terminal point D, and the profile fitting formula is:

[0016]

[0017] Wherein, is the first curve segment, is the second curve segment, is the third curve segment, is the hub diameter, is the impeller outer diameter, and θ is the polar angle, is the first curvature coefficient, is the second curvature coefficient, is the third curvature coefficient respectively, is the first radial correction amount, is the second radial correction amount, is the third radial correction amount;

[0018] The first curve segment, the second curve segment and the third curve segment are established in the polar coordinate system with the center of the impeller rotation as the origin;

[0019] The center line of the rear pressure convex is a fourth curve segment, and the profile fitting formula is:

[0020]

[0021] Wherein, is the fourth curve segment, is the fourth curvature coefficient, is the fourth radial correction amount.

[0022] Further, the pump housing inside corresponds to the centrifugal impeller is provided with independent volute flow channel, the volute flow channel beginning end and the water inlet are communicated, volute flow channel end and the water outlet are communicated.

[0023] Further, the projection of the blade in the axial direction of the impeller accounts for 55% to 65% of the outer diameter of the impeller hub.

[0024] Further, the blade thickness is 3mm to 4mm.

[0025] Further, the spacing between adjacent blades gradually increases from inside to outside along the radial direction of the impeller, the spacing at the leading edge is 0.3 times the diameter of the impeller inlet, and the spacing at the trailing edge is 0.5 times the diameter of the impeller outlet.

[0026] Further, the front and rear pressure protrusions have the same width of 2.5mm.

[0027] Further, the number of blades is 6, and the equivalent diameter of the flow channel formed between adjacent blades gradually increases in the flow direction.

[0028] Further, the bending angle of the trailing edge is 25°.

[0029] Further, a spiral flow channel is formed between adjacent blades, and the inlet width of the spiral flow channel is greater than the outlet width.

[0030] Further, a round corner transition is provided at the outlet end of the blade.

[0031] Compared with the prior art, the beneficial effects of the present application are that: by using segmented curve line design for the pressure surface and suction surface of the blade, combined with the bending radius of the trailing edge, the fluid is smoothly transitioned in the flow channel, effectively suppressing flow separation and vortex generation. The design of the rear-bent circular arc blade and the gradually expanding flow channel reduces the flow resistance and improves the hydraulic efficiency, and the pump has lower energy loss than the traditional straight blade or simply curved blade pump.

[0032] Further, the blade thickness is 3mm to 4mm, which ensures that the blade structure strength is sufficient to withstand fluid pressure and centrifugal force, while avoiding increasing flow resistance due to excessive thickness or causing blade deformation due to insufficient thickness.

[0033] Further, the blade profile is defined by a mathematical formula, and the curvature coefficient and correction amount allow accurate control of the blade shape, ensuring consistency and repeatability of manufacturing.

[0034] Further, the first curve segment is a quadratic parabolic curve, and the curvature uniformly changes with the angle, ensuring smooth and stable transition of the fluid from axial to radial, and this segment is mainly used to optimize the inlet impact and reduce flow loss.

[0035] Further, the second curve segment is an exponential growth curve, which provides faster radial expansion in the middle region of the impeller, effectively converting kinetic energy into pressure energy, and this segment is the main area of energy conversion, and its design directly affects the head and efficiency of the water pump.

[0036] Further, the third curve segment is a logarithmic decay curve, which gradually slows down the growth rate, making the profile smoothly approach the outer diameter of the impeller This segment ensures smooth flow of fluid after leaving the trailing edge of the blade, reduces vortex and separation phenomena, and reduces outlet loss.

[0037] Further, the fourth curve segment is a cubic root curve, which has a trend of first fast and then slow change in curvature with increasing polar angle θ, which can form a specific pressure gradient distribution on the suction surface, thereby suppressing boundary layer separation and enhancing the adhesion ability of fluid on the suction surface of the blade., in cooperation with the front pressure convex on the pressure surface, further optimizes the velocity field and pressure field distribution in the blade flow passage, and improves the overall hydraulic performance of the water pump.

[0038] Further, through the combination design of the above four curve segments, the pressure surface and suction surface of the blade form a three-dimensional profile with a specific curvature variation law. Among them, the three curve segments of the front pressure convex adopt the combination form of quadratic parabola, exponential curve and logarithmic curve from the inlet to the outlet of the impeller, which can be differentiated and regulated according to the flow characteristics of fluid at different radial positions: in the first curve segment area near the leading edge, the gentle change of the quadratic function guides the fluid to enter the flow passage smoothly, avoiding impact loss at the inlet; in the middle of the second curve segment area, the rapid growth characteristics of the exponential function accelerate energy conversion and improve pressure energy acquisition efficiency; in the third curve segment area near the trailing edge, the decay characteristics of the logarithmic function realize the stable transition of the flow state, and prepare for the fluid to enter the volute flow passage.

[0039] Further, the fourth curve segment of the pressure convex adopts a power function form, and its curvature variation is between the exponential growth of the second curve segment and the logarithmic decay of the third curve segment, forming a pressure gradient distribution on the suction surface that matches the pressure surface, further suppressing boundary layer separation and promoting the coordinated flow of fluid on both sides of the blade.

[0040] Further, this multi-curve combined blade profile design, combined with the parameterized adjustment of the curvature coefficients and radial correction amounts, can accurately match the fluid dynamics characteristics under different flow and head conditions, making the velocity distribution and pressure distribution on the blade surface more uniform, thereby effectively reducing hydraulic loss and improving the overall performance of the water pump.

[0041] Further, the spacing between adjacent blades gradually increases from inside to outside along the impeller radial direction, the spacing at the leading edge is 0.3 times the diameter of the impeller inlet, and the spacing at the trailing edge is 0.5 times the diameter of the impeller outlet. In this way, the flow variation law of the fluid at different positions of the impeller radial direction can be dynamically adapted. At the impeller inlet area, the smaller spacing can effectively constrain the fluid entering the flow passage, avoid vortex phenomenon caused by unstable initial flow velocity, and ensure that the fluid enters the blade working area in an orderly state; as the fluid flows outward along the radial direction, the gradually expanding spacing can provide more sufficient flow space for the fluid, reduce congestion and collision in the flow passage, and reduce local flow resistance; and at the trailing edge outlet area, the larger spacing design helps the fluid to smoothly diffuse when leaving the blade, avoids pressure fluctuation caused by limited space, further optimizes the transition process of the fluid entering the subsequent volute flow passage, and thus improves the operation stability and energy conversion efficiency of the water pump.

[0042] Further, the front and rear pressure protrusions have the same width of 2.5 mm. In this way, the protrusions can form a uniform pressure disturbance zone on the blade surface, avoid secondary flow caused by sudden change of local pressure gradient due to width difference, and facilitate unified control of parameters during mold processing, thereby reducing manufacturing difficulty and cost, and ensuring that the vortex suppression effects generated on the pressure surface and the suction surface can be matched, forming a stable pressure field throughout the flow passage, further enhancing the control ability of the fluid boundary layer, and reducing the occurrence of flow separation phenomenon. BRIEF DESCRIPTION OF DRAWINGS

[0043] Figure 1 FIG. 1 is a schematic diagram of the overall structure of a multi-impeller centrifugal water pump according to an embodiment of the present application;

[0044] Figure 2 FIG. 2 is a front view of the overall structure of a centrifugal impeller according to an embodiment of the present application;

[0045] Figure 3 FIG. 3 is a rear view of the overall structure of a centrifugal impeller according to an embodiment of the present application;

[0046] Figure 4 FIG. 4 is a schematic diagram of the left view structure of a multi-impeller centrifugal water pump according to an embodiment of the present application;

[0047] In the figure: 1, pump shell; 2, water inlet; 3, water outlet; 4, main shaft; 5, impeller hub; 6, blade; 7, first curved segment; 8, second curved segment; 9, third curved segment; 10, fourth curved segment; 13, leading edge; 14, trailing edge; 15, driving motor; 16, starting point A; 17, intermediate point B; 18, transition point C; 19, end point D. DETAILED DESCRIPTION

[0048] In order to make the objects, technical schemes and advantages of the present application clearer, the following further describes the present application with reference to the embodiments; it should be understood that the specific embodiments described herein are only used to explain the present application and not to limit the present application.

[0049] The preferred embodiments of the present application are described below with reference to the drawings. Those skilled in the art should understand that these embodiments are only used to explain the technical principles of the present application and not to limit the protection scope of the present application.

[0050] Please refer to Figures 1-4 As shown in the drawings, they are respectively a schematic diagram of the overall structure of a multi-impeller centrifugal water pump according to an embodiment of the present application; a front view of the overall structure of a centrifugal impeller according to an embodiment of the present application; a rear view of the overall structure of a centrifugal impeller according to an embodiment of the present application; and a schematic diagram of the left view structure of a multi-impeller centrifugal water pump according to an embodiment of the present application.

[0051] The multi-impeller centrifugal water pump according to an embodiment of the present application comprises:

[0052] A pump shell 1, which is provided with a water inlet 2 and a water outlet 3, and a fluid chamber is arranged between the water inlet 2 and the water outlet 3;

[0053] A main shaft 4, which is provided with a centrifugal impeller along the axial direction, the centrifugal impeller comprising an impeller hub 5 and a plurality of blades 6, and the centrifugal impeller is rotatably arranged in the fluid chamber and fixedly connected with the main shaft 4 through a flat key;

[0054] The blade 6 is a rear-bent circular arc, each blade 6 comprising a pressure surface, a suction surface, a leading edge 13 and a trailing edge 14; the leading edge 13 of the blade 6 is connected with the outer circle of the hub, and the trailing edge 14 is bent towards the opposite direction of the rotation direction of the impeller;

[0055] The pressure surface is provided with three front pressurizing protrusions, and the suction surface is provided with one rear pressurizing protrusion;

[0056] Among them, the center lines of the front pressurizing protrusions are respectively a first curve segment 7, a second curve segment 8 and a third curve segment 9;

[0057] Among them, the first curve segment 7 extends from the starting point A16 of the leading edge 13 to the intermediate point B17, and the profile fitting formula is:

[0058]

[0059] The second curve segment extends from the intermediate point B17 to the transition point C18, and the profile fitting formula is:

[0060]

[0061] The third curve segment extends from the transition point C18 to the terminal point D19 of the trailing edge, and the profile fitting formula is:

[0062]

[0063] wherein, is a first curve segment, is a second curve segment, is a third curve segment, is a hub diameter, is an impeller outer diameter, and θ is a polar angle, is a first curvature coefficient, is a second curvature coefficient, is a third curvature coefficient respectively, is a first radial correction amount, is a second radial correction amount, is a third radial correction amount;

[0064] wherein the first curve segment 7, the second curve segment 8 and the third curve segment 9 are established in a polar coordinate system with the center of impeller rotation as the origin;

[0065] the center line of the rear pressure convex is a fourth curve segment 10, and the profile fitting formula of the fourth curve segment 10 is:

[0066]

[0067] wherein is a fourth curve segment, is a fourth curvature coefficient, is a fourth radial correction amount.

[0068] Specifically, the "polar angle" θ is an angle that gradually increases from the blade leading edge A point to the blade trailing edge D point along the center line of the pressure convex. It is measured in a two-dimensional polar coordinate plane with the center of the impeller shaft as the origin, and is used to define and draw the three specially designed convex shapes on the pressure surface. The θ range of the first curve segment 7 is set to [0-π / 4), wherein θ=0 is the starting point A; the θ range of the second curve segment 8 is set to [π / 4-π / 2), wherein θ=π / 4 is the intermediate point B; the θ range of the third curve segment 9 is set to [π / 2-3π / 4], wherein θ=π / 2 is the transition point C and θ=3π / 4 is the end point D; the θ range of the fourth curve segment 10 is set to [3π / 4-π], wherein θ=3π / 4 is the starting point E and θ=π is the end point F. is 0.1m, is 0.5m, is 0.01, is 0.05, is 0.01, is 0.25, is 0.33, The value is 0.12. This setting, by clearly defining the polar angle range and key parameter values ​​of each curve segment, enables precise quantitative design of the blade profile. When θ is in the range [0-π / 4), the first curve segment controls the change in bulge height according to a quadratic function law, in conjunction with... A hub diameter of 0.1m ensures a smooth transition of the leading edge bulge from the starting point A16 to the midpoint B17, preventing severe impact when fluid enters; after θ enters [π / 4-π / 2), the exponential function of the second curve segment... The radial dimension is rapidly increased under the influence of =0.05, combined with A correction of 0.33 creates a highly efficient energy conversion zone in the middle of the impeller; while when θ is in the range of [π / 2-3π / 4], the logarithmic decay characteristic of the third curve segment is... =0.01 and With a setting of 0.12, ensure that the height of the bulge gradually converges to the outer diameter of the impeller. At a distance of 0.5m, a stable transition at the fluid outlet is achieved. The leading edge region is where the fluid impact is most direct. A smaller value is typically chosen, allowing the bulge height to increase slowly. Its main function is to smoothly guide the fluid and prevent flow separation and cavitation, rather than to significantly increase pressure. The middle section is the primary region where the fluid performs work and converts energy. A larger value is typically chosen, combined with the rapid growth of the exponential function, to significantly increase the bulge height, thereby applying a significant pressure boost to the fluid. This is the key region for increasing head. The rear section, near the outlet, requires a stable flow field. A smaller value is usually selected, and the saturation characteristics of the logarithmic function are used to make the height of the bulge converge smoothly. Its main function is to stabilize the flow field, prepare for the fluid to enter the volute smoothly, and avoid chaotic flow at the outlet. This protrusion, located on the suction side, primarily functions to control the boundary layer and suppress flow separation commonly seen on suction surfaces. Its value needs to be coordinated with the protrusion on the pressure side, and its improvement effect on the wake region and eddies is typically observed through simulation. The radial correction amount is dynamically adjusted based on the overall hydraulic performance requirements of the impeller. As a correction parameter for the leading edge region, it is mainly used to compensate for the influence of fluid inlet pre-swirl on pressure distribution. When positive pre-swirl exists in the system, it can be appropriately increased. To enhance the guiding effect; For the intermediate energy conversion region, and The synergistic effect ensures that the pressure energy improves efficiency while avoiding flow blockage caused by excessive local pressure. Generally, it increases linearly with the increase of the design flow rate. Focusing on the export transition phase, its value needs to match the volute inlet size; when the volute inlet cross-sectional area is large, it can be reduced. This extends the diffusion process in the flow channel and ensures the compatibility between the fluid and the volute flow channel. As a correction term of the suction surface post-pressurization protrusion, it needs to be complementary to each correction amount of the pressure surface, to balance the pressure difference on both sides of the blade by fine-tuning the protrusion height, and further optimize the boundary layer flow state. The coupling regulation and control among each correction amount is realized through a multi-objective optimization algorithm based on CFD simulation, to achieve the best balance state of the hydraulic efficiency, cavitation margin and operation stability on the premise of meeting the blade structure strength constraint. Among them, and In the formula, only the numerical values are taken for fitting the corresponding line segment. The coordinated cooperation of each parameter makes the front-pressurization protrusion of the blade pressure surface not only be able to optimize the flow characteristics in different regions, but also be able to ensure the continuity and smoothness of the overall profile, thereby improving the hydraulic efficiency of the water pump. It should be pointed out that the data in the embodiment are the results obtained through the method of the present application through pre-experiment verification before the operation of the method, and each preset value can be adjusted according to the specific use, as long as the method of the present application can clearly define different specific conditions in the single item judgment process through the obtained numerical values.

[0069] Specifically, the main shaft 4 is driven by a driving motor 15, the output shaft of the driving motor 15 is fixedly connected with one end of the main shaft 4 through a shaft coupling, and the rotating speed of the motor can be adjusted according to the actual operation requirement, so as to realize the dynamic control of the flow and head of the water pump. The pump shell 1 is casted by cast iron, and the inner wall thereof is precisely machined with a surface roughness Ra≤1.6 μm, so as to reduce the flow resistance of the fluid in the cavity. The water inlet 2 and the water outlet 3 are both provided with a standard flange interface, the flange sealing surface adopts a tenon and slot structure, and is matched with a nitrile rubber gasket to realize reliable sealing and prevent medium leakage. The internal flow channel of the fluid cavity is designed as a smooth transition curved surface structure, the flow channel cross-sectional area gradually expands from the water inlet 2 to the impeller inlet, forming a diffusion section, which can reduce the flow velocity of the fluid entering the impeller and reduce the inlet impact.

[0070] Specifically, the pump shell 1 is provided with an independent volute flow channel corresponding to the centrifugal impeller, the volute flow channel has a start end communicated with the water inlet 2 and an end communicated with the water outlet 3. In this way, the independent volute flow channel can convert the kinetic energy of the fluid into static pressure energy through the gradually expanding flow channel cross section according to the fluid velocity and pressure distribution characteristics of the centrifugal impeller, so as to reduce the collision and vortex loss of the fluid in the pump shell 1. At the same time, the geometric shape design of the volute flow channel can make the fluid maintain a stable pressure gradient during the flow process, avoid local pressure fluctuation caused by sudden change of the flow channel, further improve the hydraulic efficiency and operation stability of the water pump, and ensure that the high-pressure fluid flowing out of the impeller can be smoothly discharged through the water outlet 3, so as to meet the pressure and flow requirements in actual application.

[0071] Specifically, the projection of the blade 6 in the axial direction of the impeller along the main shaft 4 accounts for 55% to 65% of the outer diameter of the impeller hub 5. In this way, the blade 6 can have sufficient working area to drive the fluid while avoiding the increase in the overall weight of the impeller and the waste of materials due to the excessively large projection ratio, and can also prevent the flow channel from being too narrow to cause flow congestion. When the projection ratio is in this range, the contact area of the blade 6 with the fluid is moderate, which can effectively capture the fluid and transfer energy, and also reserve reasonable space for the flow of the fluid in the flow channel, thereby reducing the flow resistance. Through experiments, it has been verified that the projection ratio of 55% to 65% can make the blade 6 form a good interaction with the fluid during rotation, thereby improving the head and flow of the water pump while reducing the additional energy consumption caused by the improper size of the blade 6, and ensuring the balance between efficient work and lightweight design of the impeller.

[0072] Specifically, the thickness of the blade 6 is 3 mm to 4 mm. In this way, the structural strength and hydraulic performance of the blade 6 can be considered. The thickness range of 3 mm to 4 mm can ensure that the blade 6 has sufficient rigidity to withstand the pressure of the fluid and the centrifugal force, while avoiding the problem of the flow channel being too narrow and the flow resistance increasing due to the excessively thick blade 6. Thinner blade 6 can reduce the disturbance to the fluid, making the fluid flow more smoothly in the flow channel, and reducing the boundary layer separation and vortex loss caused by the excessively large thickness of the blade 6; and the appropriate thickness ensures that the blade 6 is not easily deformed or damaged under long-term high-speed rotation conditions, thereby prolonging the service life of the impeller. In addition, this thickness range also facilitates the welding or integral molding processing of the blade 6 and the impeller hub 5, ensuring the structural stability of the connection part, so that the blade 6 can reliably transmit torque and work efficiently.

[0073] Specifically, the spacing between adjacent blades 6 gradually increases from the inside to the outside along the radial direction of the impeller, and the spacing at the leading edge 13 is 0.3 times the diameter of the impeller inlet, and the spacing at the trailing edge 14 is 0.5 times the diameter of the impeller outlet. In this way, the flow velocity of the fluid in the flow channel can change uniformly along the radial direction, avoiding the flow separation caused by the excessively small spacing. The smaller spacing at the leading edge 13 can ensure that the fluid entering the flow channel has sufficient constraint, thereby reducing the inlet impact loss; as the fluid flows to the outer diameter of the impeller, the gradually increasing spacing provides sufficient space for the diffusion and energy conversion of the fluid, making the conversion process of kinetic energy to pressure energy more stable and efficient, while reducing the turbulence intensity in the flow channel, thereby further improving the hydraulic performance of the water pump.

[0074] Specifically, the front and rear pressure protrusions have the same width of 2.5 mm. With such a setting, the pressure disturbances generated by the pressure protrusions on the pressure surface and the suction surface when interacting with the fluid are matched, avoiding an imbalance in the flow field distribution around the blade 6 due to the difference in the width of the protrusions on both sides. The width of 2.5 mm can form an effective local pressure gradient to guide the flow of fluid and enhance the control effect on the boundary layer, without increasing the flow resistance or inducing new vortex flow due to the excessive width of the protrusions. The same width also simplifies the machining process of the blade 6, ensures the symmetry of the pressure surface and the suction surface during the manufacturing process, reduces the imbalance of the impeller rotation caused by the structural asymmetry, improves the stability and reliability of the water pump during operation, and facilitates the unified analysis and optimization of the effect of the protrusions during the design stage.

[0075] Specifically, the number of blades 6 is 6, and the equivalent diameter of the flow passage formed between adjacent blades 6 gradually increases along the flow direction. The number of 6 blades 6 ensures an adequate number of flow passages while avoiding the problem of excessively narrow flow passages caused by too many blades 6, allowing the fluid to be evenly distributed between the blades 6 and reducing mutual interference. The gradual increase in the equivalent diameter of the flow passage along the flow direction is adapted to the pressure recovery process of the fluid in the impeller. As the fluid flows from the inlet to the outlet of the impeller, its pressure continuously increases and its flow velocity gradually decreases. The increased equivalent diameter provides a reasonable space for the deceleration and pressure recovery of the fluid, effectively reducing the flow loss in the flow passage, improving the energy conversion efficiency, avoiding vortex phenomena caused by sudden changes in the cross section of the flow passage, and ensuring the smoothness of fluid flow.

[0076] Specifically, the bending angle of the trailing edge 14 is 25°. This bending angle is determined based on fluid dynamics simulation and experimental optimization, which ensures the effective flow guiding effect of the trailing edge 14 of the blade 6 on the fluid and avoids flow separation caused by an excessively large bending angle. The bending angle of 25° allows the fluid to smoothly flow out along the tangent direction of the volute-shaped flow passage when leaving the blade 6, reducing the impact of the fluid on the inner wall of the pump housing 1 and reducing local vortex loss. At the same time, this angle is well coordinated with the pressure surface profile and the suction surface profile of the blade 6, making the pressure distribution on the surface of the blade 6 more uniform and further improving the work efficiency of the impeller. In actual applications, the bending angle of the trailing edge 14 can be adjusted according to the design flow rate, head, and other parameters of the water pump, but 25° is the optimal value range verified, which can balance the hydraulic performance and structural strength under most working conditions.

[0077] Specifically, a spiral flow channel is formed between adjacent blades 6, and the inlet width of the spiral flow channel is greater than the outlet width. This design allows the fluid to obtain a larger initial flow area when entering the flow channel, reduces the flow velocity at the inlet, and reduces the local resistance loss caused by excessive flow velocity. As the fluid flows along the spiral trajectory to the outlet, the width of the flow channel gradually decreases, and the equivalent diameter of the flow channel increases, forming a reasonable velocity gradient and promoting the smooth conversion of fluid kinetic energy into pressure energy. The spiral flow channel shape can guide the fluid to make an orderly rotational motion, avoiding the flow dead zone and backflow phenomenon that easily occurs in the traditional straight blade 6 flow channel, making the velocity distribution of the fluid more uniform in the entire flow channel, and effectively improving the flow capacity and hydraulic efficiency of the flow channel. At the same time, the structure design of wide inlet and narrow outlet cooperates with the bending angle of the blade 6 trailing edge 14, further optimizes the flow state of the fluid when leaving the impeller, and lays a good foundation for the subsequent diffusion process into the volute flow channel.

[0078] Specifically, the outlet end of the blade 6 is provided with a fillet transition. The value of the fillet radius is set to 0.8mm-1.2mm. By using a circular arc to smoothly transition at the junction of the blade 6 trailing edge 14 and the suction surface and the pressure surface, local stress concentration caused by a right-angle edge is avoided, and at the same time, the circular arc structure can guide the fluid to form a stable boundary layer at the outlet of the blade 6, reducing the vortex loss caused by flow separation. The fillet structure can also reduce the direct scouring of high-speed fluid on the blade 6 trailing edge 14, delay the generation and development of cavitation, and especially when conveying media containing small particles, the wear rate of the blade 6 can be significantly reduced, thereby prolonging the maintenance period and service life of the water pump.

[0079] Thus, the technical solutions of the present application have been described in conjunction with the preferred embodiments shown in the drawings, but those skilled in the art will readily understand that the protection scope of the present application is obviously not limited to these specific embodiments. Those skilled in the art can make equivalent changes or replacements to the related technical features without departing from the principles of the present application, and the technical solutions after such changes or replacements will all fall within the protection scope of the present application.

[0080] The above description is only the preferred embodiments of the present application and is not intended to limit the present application; for those skilled in the art, the present application can have various changes and variations. Any modification, equivalent replacement, improvement, etc. made within the spirit and principles of the present application shall be included in the protection scope of the present application.

Claims

1. A multi-impeller centrifugal water pump characterized by, It includes: Pump shell, the pump shell is equipped with water inlet and water outlet, water inlet and water outlet are provided with fluid chamber; Main shaft, the main shaft is provided with centrifugal impeller along the axial direction, the centrifugal impeller includes impeller hub and several blades, and the centrifugal impeller is rotationally arranged in the fluid chamber and fixedly connected with the main shaft by means of a flat key; The blade is a rear bending type circular arc, each blade includes a pressure surface, a suction surface, a leading edge and a trailing edge; The leading edge of the blade is connected with the outer circumference of the hub, and the trailing edge is bent towards the opposite direction of the rotation direction of the impeller; The pressure surface is provided with three front pressure increasing protrusions, and the suction surface is provided with a rear pressure increasing protrusion; Wherein, the center line of each front pressure increasing protrusion is respectively a first curve segment, a second curve segment and a third curve segment; The first curve segment extends from the leading edge starting point A to the intermediate point B, and the profile fitting formula thereof is: ; The second curve segment extends from the intermediate point B to the transition point C, and its profile fitting formula is: ; The third curve segment extends from the transition point C to the trailing edge end point D, and its profile fitting formula is: ; wherein, is a first curve segment, is a second curve segment, is a third curve segment, is a hub diameter, is an impeller outer diameter, and θ is a polar angle, is a first curvature coefficient, is a second curvature coefficient, is a third curvature coefficient, respectively, is a first radial correction amount, is a second radial correction amount, is a third radial correction amount; Wherein, the first curve segment, the second curve segment and the third curve segment are established in the polar coordinate system with the center of the impeller rotation as the origin; The center line of the rear press protrusion is a fourth curve segment, and the profile fitting formula is: ; wherein is a fourth curve segment, is a fourth curvature coefficient, is a fourth radial correction amount.

2. The centrifugal water pump of claim 1, wherein The inside of the pump shell is provided with an independent volute type flow channel corresponding to the centrifugal impeller, the volute type flow channel is in communication with the water inlet at the beginning end, and the volute type flow channel is in communication with the water outlet at the end.

3. The centrifugal water pump of claim 2, wherein The projection of the blade in the axial direction of the impeller accounts for 55% to 65% of the outer diameter of the impeller hub.

4. The centrifugal water pump of claim 3, wherein The thickness of the blade is 3mm to 4mm.

5. The centrifugal water pump of claim 4, wherein The spacing between adjacent blades gradually increases from inside to outside along the radial direction of the impeller, the spacing at the leading edge is 0.3 times the diameter of the impeller inlet, and the spacing at the trailing edge is 0.5 times the diameter of the impeller outlet.

6. The centrifugal water pump of claim 5, wherein The width of the front pressure increasing protrusion and the rear pressure increasing protrusion is the same, which is 2.5mm.

7. The centrifugal water pump of claim 6, wherein The number of blades is 6, and the equivalent diameter of the flow channel formed between adjacent blades gradually increases along the flow direction.

8. The centrifugal water pump of claim 7, wherein The bending angle of the trailing edge is 25°.

9. The centrifugal water pump of claim 8, wherein The spiral flow channel is formed between adjacent blades, and the inlet width of the spiral flow channel is greater than the outlet width.

10. The centrifugal water pump of claim 9, wherein The blade outlet end is provided with a fillet transition.

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