Large-flow end cover type double-suction hydraulic turbine device for chemical industry

Abstract

The invention relates to the technical field of fluid machinery, and particularly discloses a chemical large-flow end cover type double-suction hydraulic turbine device which comprises a shell, a turbine main shaft, an impeller and a guide vane. A fluid channel is formed in the shell, two ends of the fluid channel are provided with a fluid inlet and a fluid outlet, and two sides are connected with a left end cover and a right end cover. The two ends of a turbine spindle are supported in the centers of the left and right end covers through bearings, and an impeller is fixedly connected to the spindle and located in the shell. The guide vane is arranged in the shell, located between the fluid inlet and the impeller and used for guiding fluid. Wherein a spiral suction chamber, a guide vane and a semi-spiral delivery chamber are sequentially arranged in the shell in the fluid flowing direction, and the spiral suction chamber and the semi-spiral delivery chamber jointly form a volute structure. The device is mainly used for recycling pressure energy of high-pressure fluid at the tail end of the chemical process and converting the pressure energy into mechanical energy, and the structural design of the device is beneficial to achieving efficient energy conversion and stable operation under the large-flow working condition.

Description

Technical Field

[0001] This invention relates to the field of fluid machinery technology, and in particular to a high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications. Background Technology

[0002] In chemical processes such as coal chemical engineering and petroleum hydrocracking, a large amount of high-pressure fluid containing pressure energy often exists at the end of the system. Traditionally, these high-pressure fluids are usually directly depressurized through pressure reducing valves, resulting in the waste of their contained pressure energy.

[0003] Hydraulic turbines, as energy recovery devices, can convert the pressure energy of fluids into mechanical energy to drive generators, pumps, or other equipment, making them an effective technical approach for recovering and utilizing residual pressure energy. However, in practical applications in the chemical industry, existing hydraulic turbine equipment still faces some challenges: On the one hand, there is a limited selection of turbine equipment suitable for high-flow, low-to-medium head conditions, capable of withstanding long-term corrosion from chemical media and operating in high-temperature and high-pressure environments. Some high-performance products have long relied on imports, resulting in high purchase and maintenance costs and long service cycles. On the other hand, chemical processes place extremely high demands on the operational stability and reliability of equipment, requiring energy recovery devices to operate stably and efficiently under complex media (potentially containing trace particles, corrosive, etc.) and variable operating conditions, while also facilitating routine maintenance and replacement of critical wear parts (such as mechanical seals).

[0004] Therefore, there is an urgent need for a hydraulic turbine device that is optimized for the characteristics of high-flow-rate chemical processes, and that combines high efficiency, high operational stability, corrosion resistance, and ease of installation and maintenance. Summary of the Invention

[0005] The purpose of this invention is to provide a high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications, in order to solve the above-mentioned technical problems existing in the prior art.

[0006] To achieve the above objectives, the present invention provides the following solution: a high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications, comprising: a housing having a fluid channel formed inside, with a fluid inlet and a fluid outlet at both ends of the housing; a left end cap and a right end cap connected to both sides of the housing; a turbine main shaft supported at both ends by bearings at the centers of the left and right end caps; an impeller disposed within the housing and fixedly connected to the turbine main shaft; and a guide vane disposed within the housing and located between the fluid inlet and the impeller for guiding the fluid; wherein, a spiral suction chamber, the guide vane, and a semi-spiral discharge chamber are sequentially arranged inside the housing along the fluid flow direction, and the spiral suction chamber and the semi-spiral discharge chamber together constitute a volute structure.

[0007] Optionally, the guide vanes are arranged according to a specific timing angle α, which is the angle between the line connecting the trailing edge of the guide vane and the axis and the line connecting the volute tongue and the axis. The value of α ranges from 0° to 360° / Z, where Z is the hydraulically calculated value of the number of main blades of the impeller.

[0008] Optionally, the number of blades Z0 of the guide vane satisfies the relationship: Z0 = Z + 3.

[0009] Optionally, the timing angle α is 10°.

[0010] Optionally, the impeller is a double-suction impeller, with the left and right impellers arranged in an alternating manner.

[0011] Optionally, the stagger angle between the left and right impellers is 36°.

[0012] Optionally, the impeller has an intermediate partition between its two halves.

[0013] Optionally, the impeller blades are in the form of composite blades, including the same number of main blades, intermediate blades, and short blades.

[0014] Optionally, the inlet angle of the middle blade and the short blade is consistent with that of the main blade, and their outlet angle is the same as that of the main blade at the corresponding position.

[0015] Optionally, the circumferential offset angle θ of the short blade is the included angle between two adjacent main blades; the inlet diameter D of the short blade S Satisfying the relation: D S =0.5(D1-D2)+D2, where D1 is the impeller outlet diameter and D2 is the impeller inlet diameter; the outlet deflection angle of the short blade is -5°.

[0016] Compared with the prior art, the present invention discloses at least the following beneficial effects:

[0017] This invention utilizes a volute structure comprised of a spiral suction chamber and a semi-spiral discharge chamber to achieve efficient flow guidance and energy conversion of high-pressure fluids in chemical processes. This allows the fluid to enter subsequent working units smoothly and uniformly, effectively reducing hydraulic losses and inlet impact. Guide vanes, positioned between the fluid inlet and the impeller, pre-guide and rectify the fluid, optimizing the flow field entering the impeller and improving energy recovery efficiency. The rigid connection between the double-suction impeller and the turbine shaft, combined with bearings supported by left and right end covers, significantly increases the device's flow handling capacity, adapting to high-flow-rate conditions, and ensures the rotor's smooth and reliable operation. The overall detachable end-cover structure greatly facilitates the inspection and replacement of vulnerable internal components (such as mechanical seals), while the modular design of the casing and end covers makes installation and maintenance more convenient. This integrated structural design ultimately enables the device to achieve efficient and stable pressure energy recovery under complex chemical media and harsh operating conditions, while also possessing good maintainability and long-term operational stability. Attached Figure Description

[0018] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the drawings used in the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0019] Figure 1 This is a front view of the high-flow-rate end-cap type double-suction turbine device for chemical applications of the present invention.

[0020] Figure 2 This is a left view of the high-flow-rate end-cap type double-suction turbine device for chemical applications of the present invention.

[0021] Figure 3 This is an isometric view of the high-flow-rate end-cap type double-suction turbine device for chemical applications of the present invention.

[0022] Figure 4 This is an exploded view of the high-flow-rate end-cap type double-suction turbine device for chemical applications of the present invention.

[0023] Figure 5 This is an exploded view of the interior of the housing of the device of the present invention;

[0024] Figure 6 This is a first-view cross-sectional view of the device of the present invention;

[0025] Figure 7 This is a second-view cross-sectional view of the device of the present invention;

[0026] Figure 8 This is a schematic diagram of the timing arrangement of the guide vanes in the device of the present invention;

[0027] Figure 9 This is a schematic diagram of the impeller blades in the device of the present invention;

[0028] Figure 10 This is a schematic diagram showing the translation of the impeller cover plate in the device of the present invention;

[0029] Figure 11 This is a characteristic curve diagram of the device of the present invention;

[0030] Figure 12 This is a time-domain diagram of pressure pulsation at the diaphragm monitoring point of the device of the present invention at different guide vane timings;

[0031] Figure 13 The present invention presents time-domain diagrams of the resultant radial forces of the staggered impellers under different flow rates in the device of the present invention.

[0032] Figure 14 The load distribution curves of the long, medium, and short impeller blades of the device of the present invention under low flow conditions are shown. Figure 1 ;

[0033] Figure 15 The load distribution curves of the long, medium, and short impeller blades of the device of the present invention under low flow conditions are shown. Figure 2 .

[0034] In the diagram: 1. Base; 2. Turbine main shaft; 3. Left end cover; 4. Housing; 5. Flange; 6. Right end cover; 7. Guide vane; 8. Impeller. Detailed Implementation

[0035] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0036] To make the above-mentioned objects, features and advantages of the present invention more apparent and understandable, the present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0037] Reference Figures 1 to 15As shown, this embodiment provides a high-flow-rate, end-cap type, double-suction hydraulic turbine device for chemical applications. The device includes a base 1, a turbine main shaft 2, a left end cap 3, a housing 4, a flange 5, a right end cap 6, guide vanes 7, and an impeller 8. The housing 4 is fixedly mounted on the base 1, and its interior forms a flow channel. Fluid inlets and outlets are located at both ends of the housing 4, and it is connected to external high-pressure and low-pressure fluid pipelines via the flange 5. The fluid inlet is connected to the external high-pressure fluid pipeline, and the fluid outlet is connected to the low-pressure fluid pipeline. Inside the housing 4, a spiral suction chamber, guide vanes 7, and a semi-spiral discharge chamber are sequentially arranged along the fluid flow direction. The spiral suction chamber and the semi-spiral discharge chamber together constitute the turbine's volute structure. This structure adopts a radially split, two-end-supported form, resulting in a compact overall structure that facilitates installation and maintenance.

[0038] In this embodiment, the impeller 8 is a double-suction impeller, which is fixedly connected to the turbine main shaft 2 by a key, forming the rotor component of the turbine. The impeller 8 is located downstream of the guide vane 7 and is installed in the cavity formed by the spiral suction chamber and the semi-spiral discharge chamber. The two ends of the turbine main shaft 2 are respectively supported on bearing seats inside the left end cover 3 and the right end cover 6. The left end cover 3 and the right end cover 6 are detachably fixed to the two end faces of the housing 4 by bolts. Mechanical seal elements are provided between the end covers and the main shaft to prevent leakage of internal high-pressure and corrosive media. Furthermore, the end cover structure design greatly facilitates the quick replacement and daily maintenance of vulnerable parts such as mechanical seals.

[0039] To reduce the impact and alternating load on the blades caused by the high-pressure fluid flowing out of the suction chamber and to improve the operational stability of the turbine, a special type of guide vane 7 is used to guide the fluid. This allows the guide vane 7 to better constrain the fluid, creating more circulation before it enters the impeller 8, thereby reducing impact and enhancing flow stability. The core measure lies in the timing arrangement of the guide vane 7.

[0040] Specifically, the timing angle α of guide vane 7 is defined as the angle between the line connecting the trailing edge of guide vane 7 and the axis and the line connecting the volute tongue and the axis, with a value range of [0°, 360° / Z], where Z is the hydraulically calculated value of the number of main blades of impeller 8. The number of blades of guide vane 7 is Z0 = Z + 3, therefore the actual value range of α is [0°, 40°], as shown in the schematic diagram below. Figure 8 As shown.

[0041] The study selected α=0° as the initial scheme (denoted as CL0), and then rotated the guide vane 7 clockwise by 10° each time, resulting in four schemes with α=0°, 10°, 20°, and 30°, named CL0, CL1, CL2, and CL3, respectively. Numerical calculation results show that under optimal operating conditions, the hydraulic efficiency of the time-series scheme CL1 is only 0.43% lower than that of the highest CL0. Regarding pressure pulsation, CL1 has the smallest pulsation amplitude and dominant frequency amplitude at the volute tongue, decreasing by 23.62% and 6.59% respectively compared to the maximum amplitude; simultaneously, CL1 also exhibits smaller pressure pulsation within the guide vane 7 and impeller 8 flow channels. Furthermore, when α=10°, the radial force on impeller 8 is minimized, and the fluid-structure interaction is significantly reduced.

[0042] Considering factors such as hydraulic efficiency, turbulent kinetic energy, and pressure pulsation distribution, CL1 (timing angle α = 10°) is the optimal timing position among the four timing schemes. Numerical simulations and experimental verification show that when the timing angle α = 10°, the turbine exhibits high hydraulic efficiency under optimal operating conditions. Furthermore, the amplitude of pressure pulsation and the dominant frequency are significantly reduced at the volute tongue, the guide vane 7 flow channel, and the impeller 8 flow channel. The radial force on the impeller 8 is minimized, resulting in optimal overall operational stability. This special guide vane 7 structure effectively guides the fluid, creating more circulation before it enters the impeller 8, reducing the impact of the fluid on the blades, and thus optimizing the flow state.

[0043] In one specific embodiment, the double-suction impeller 8 adopts a back-to-back symmetrical structure, which can offset axial force and only bear radial force. However, the angle setting of the left and right impellers 8 has an important impact on the hydraulic performance of the turbine. To optimize performance, based on the original impeller 8 design, while keeping other parameters unchanged, two new models are constructed by offsetting the two impellers 8 by a certain angle: staggered 0° impeller 8 (i.e., adding a baffle in the middle of the original symmetrical impeller 8) and staggered 36° impeller 8 (arranging the two impellers 8 staggered by 36° based on the 0° base).

[0044] Numerical calculation results show that the pressure pulsation amplitude of the staggered impeller 8 is lower than that of the symmetrically arranged impeller 8, which can effectively improve the operational stability of the hydraulic turbine device.

[0045] Regarding the equivalent stress distribution: the stress in the staggered 36° impeller 8 is mainly concentrated at the connection between the blade inlet and the shroud, as well as on the intermediate partition plate. The stress extends along the connection area between the blade and the partition plate towards the outlet, resulting in a reduction of low-stress areas on the front and rear shroud plates. Analysis shows that the addition of the intermediate partition plate helps to disperse stress and reduce the overall equivalent stress of the impeller 8, thereby stabilizing the blades; while the staggered 36° arrangement causes the intermediate partition plate to bear more stress, extending it towards the blade outlet.

[0046] Under optimal flow conditions, the maximum radial force of the staggered 36° impeller 8 is only 0.63 times that of the staggered 0° scheme and 0.6 times that of the symmetrical arrangement scheme, indicating a significant reduction in radial force. This demonstrates that the staggered arrangement can improve the fluid pressure distribution around the impeller 8 and enhance the unit's operational stability.

[0047] Furthermore, the pressure pulsation amplitude of the staggered 36° impeller 8 at each monitoring point was lower than that of the other two impeller models 8. Based on the comprehensive pressure pulsation analysis results, the staggered 36° impeller 8 performed best and was therefore determined as the preferred impeller model 8 used in this embodiment.

[0048] In one specific embodiment, the impeller 8 employs several optimized designs to improve performance and lifespan. Firstly, the impeller 8 adopts a composite blade design combining long, medium, and short blades, such as... Figure 9 As shown. Specifically, based on the number Z of main blades determined by hydraulic calculations, medium blades (number Z) and short blades (number 2Z) are added in each flow channel. The inlet angle of the medium and short blades is consistent with that of the main blades, and the outlet angle is the same as that of the main blades at the corresponding positions.

[0049] Key design parameters for the splitter blades (short blades) include inlet diameter, circumferential offset angle, outlet deflection angle, and number of blades. In this design, the circumferential offset angle of the short blades is set to θ (θ is the angle between the two long blades, which is 30° here), and the inlet diameter D... S The calculation formula is D S =0.5(D1-D2)+D2, and the outlet deflection angle is -5°. Based on the above parameters, the design of the long-medium-short composite impeller 8 was completed.

[0050] Numerical calculations were performed on the split-blade scheme (Z=6+6+12, i.e., 6 long blades, 6 medium blades, and 12 short blades) and the original scheme containing only main blades (Z=6), under the condition of matching the same volute and guide vanes 7. The results show that the composite impeller 8 can increase the head under low flow conditions, make the outlet pressure distribution of impeller 8 more uniform, and reduce overall pressure pulsation; among them, the Z=6+6+12 model shows a particularly significant reduction in pressure pulsation at the volute tongue. The introduction of medium and short blades effectively optimizes the blade load distribution, increases the blade frequency by increasing the number of blades, and disperses the flow channel excitation energy, thereby suppressing pressure pulsation under the design conditions. The composite impeller 8 also significantly reduces radial and axial forces, enhances operational stability, improves the vortex structure and outlet slip phenomenon in the flow channel, and reduces pressure pulsation caused by interference between impeller 8 and the tongue. Under optimal conditions, the axial force of the Z=6+6+12 impeller 8 is reduced by 0.071% compared to the original impeller 8. In summary, Z=6+6+12 was determined to be the optimal 8-blade arrangement for the impeller.

[0051] In one specific embodiment, the front cover plate of impeller 8 underwent an optimized translational design. To investigate the impact of changes in the inlet width of impeller 8 on performance, based on the original hydraulic calculation parameters of impeller 8 (inlet width b1), two new impeller 8 models were constructed by translating the front cover plate of impeller 8 to the left and right by 5% respectively: after translating to the left, the inlet width decreased to b2 = 0.95b1, and after translating to the right, the inlet width increased to b3 = 1.05b1, as detailed below. Figure 10 As shown.

[0052] Three impeller 8 models (original width b1, reduced width b2, and increased width b3) were matched with the same set of guide vanes 7 and volute, and numerical calculations were performed. The results show that the impeller 8 model shifted 5% to the right and increased the inlet width to b3 has the highest hydraulic efficiency and the best operational stability.

[0053] In one specific embodiment, the turbine spindle 2 is preferably made of SUS630 stainless steel. SUS630 stainless steel has a density of 7800 kg / m³. 3 With its elastic modulus of 210 GPa and Poisson's ratio of 0.3, this material possesses excellent corrosion resistance and high temperature and pressure resistance, making it highly suitable for conveying corrosive, high temperature and high pressure media commonly found in the chemical industry. This ensures the long-term reliable operation of the core transmission components under harsh working conditions.

[0054] The specific working process of the turbine device in this embodiment is as follows: At the end of processes such as coal chemical engineering and petroleum hydrocracking, fluid media with high pressure energy still enters the spiral suction chamber of the casing 4 through the flange 5 from the external pipeline. Guided by the suction chamber, the fluid flows through the guide vanes 7 with a specific timing angle, is rectified and guided, and then smoothly enters the inlet of the double-suction impeller 8. The high-pressure fluid impacts the blades of the impeller 8, driving the impeller 8 to rotate. The impeller 8 transmits torque to the turbine main shaft 2 through a key, thereby converting the pressure energy of the fluid into the mechanical energy of the main shaft. The fluid after work flows out from the outlet of the impeller 8 and enters the semi-spiral discharge chamber for energy collection and integration. Finally, it is discharged through the flange 5 at the other end to the low-pressure pipeline or the next process. The rotational mechanical energy of the turbine main shaft 2 can be directly used to drive a generator to generate electricity or to drive other machinery such as centrifugal pumps, thereby achieving efficient recovery and utilization of residual pressure energy in the chemical process.

[0055] As can be seen from the above technical solution, the high-flow end-cap type double-suction hydraulic turbine device provided in this embodiment achieves efficient, stable, and safe recovery of pressure energy under high-flow, medium-low head, and complex chemical media conditions through the comprehensive design and coordination of multiple features such as the spiral suction chamber and semi-spiral discharge chamber volute structure, the guide vane 7 with a specific timing angle, the impeller 8 with long, medium, and short composite blades arranged in a staggered 36° pattern and optimized front cover position, the corrosion-resistant SUS630 stainless steel main shaft, and the easy-to-maintain left and right end-cap type 6 structure. All performance indicators have reached or exceeded the level of similar imported products.

[0056] In the description of this invention, it should be understood that the terms "longitudinal", "lateral", "up", "down", "front", "rear", "left", "right", "vertical", "horizontal", "top", "bottom", "inner", "outer", etc., indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings, and are only for the convenience of describing this invention, and are not intended to indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation, and therefore should not be construed as a limitation of this invention.

[0057] The embodiments described above are merely preferred embodiments of the present invention and are not intended to limit the scope of the present invention. Various modifications and improvements made by those skilled in the art to the technical solutions of the present invention without departing from the spirit of the present invention should fall within the protection scope defined by the claims of the present invention.

Claims

1. A high-flow-rate, end-cap type, double-suction hydraulic turbine device for chemical applications, characterized in that, include: The housing (4) has a fluid channel inside it, and the two ends of the housing (4) are respectively provided with a fluid inlet and a fluid outlet; the left end cover (3) and the right end cover (6) are respectively connected to the two sides of the housing (4). The turbine spindle (2) is supported at both ends by bearings at the center of the left end cover (3) and the right end cover (6); The impeller (8) is located inside the housing (4) and is fixedly connected to the turbine main shaft (2); The guide vane (7) is disposed inside the housing (4) and located between the fluid inlet and the impeller (8) for guiding the fluid; The shell (4) contains a spiral suction chamber, a guide vane (7) and a semi-spiral discharge chamber arranged sequentially along the fluid flow direction. The spiral suction chamber and the semi-spiral discharge chamber together form a volute structure.

2. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 1, characterized in that, The guide vanes (7) are arranged according to a specific timing angle α, which is the angle between the line connecting the trailing edge of the guide vane (7) and the axis and the line connecting the volute tongue and the axis. The value of α ranges from 0° to 360° / Z, where Z is the hydraulically calculated value of the number of main blades of the impeller (8).

3. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 2, characterized in that, The number of blades Z0 of the guide vane (7) satisfies the relationship: Z0 = Z + 3.

4. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 2 or 3, characterized in that, The timing angle α is 10°.

5. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 1, characterized in that, The impeller (8) is a double-suction impeller, and the impellers (8) on the left and right sides are arranged in an alternating manner.

6. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 5, characterized in that, The stagger angle of the left and right impellers (8) is 36°.

7. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 5 or 6, characterized in that, The impeller (8) has a middle partition plate between its two halves.

8. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 1 or 5, characterized in that, The blades of the impeller (8) are in the form of composite blades, including the same number of main blades, intermediate blades and short blades.

9. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 8, characterized in that, The inlet angle of the middle blade and the short blade is consistent with that of the main blade, and their outlet angle is the same as that of the main blade at the corresponding position.

10. The high-flow-rate end-cap type double-suction hydraulic turbine device for chemical applications according to claim 9, characterized in that, The circumferential offset angle θ of the short blade is the included angle between two adjacent main blades; the inlet diameter D of the short blade S Satisfying the relation: D S =0.5(D1-D2)+D2, where D1 is the outlet diameter of the impeller (8) and D2 is the inlet diameter of the impeller (8); the outlet deflection angle of the short blade is -5°.

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