Design method of composite impeller of liquid ring pump

Abstract

The application relates to a design method of a composite impeller of a liquid ring pump. The method is based on the internal flow characteristics of the liquid ring pump, and through theoretical derivation, a formula about the blade number of the composite impeller, a formula about the bias angle of the short blade, a formula about the inlet diameter of the short blade and geometric constraint relationships thereof are obtained. Through the design of the composite impeller, the blade number is increased, the constraint ability on the fluid is enhanced, the flow distribution in the impeller flow passage is improved, and the performance and stability of the liquid ring pump are improved.

Description

Technical Field

[0001] This application relates to the field of liquid ring pumps and composite impeller design, and in particular to a method for designing a composite impeller for a liquid ring pump. Background Technology

[0002] A liquid ring pump is a fluid machinery device that uses a liquid as the intermediate medium for energy transfer to pump gases. Liquid ring pumps have advantages such as large flow rate, compact structure, and isothermal compression, and are widely used in chemical, power, and mining industries. However, due to the asymmetry and complexity of the flow within the liquid ring pump, it generates significant vibration and noise during operation, which not only reduces the pump's performance but also seriously affects its reliability and stability.

[0003] The design of compound impellers with long and short blades has been widely used in centrifugal pump design to suppress the jet wake effect at the impeller blade outlet in low specific speed centrifugal pumps and improve the velocity distribution within the impeller flow channel, thereby enhancing centrifugal pump performance. Although compound impellers are now also used in liquid ring pumps, their design methods still largely follow those of centrifugal pumps, typically relying on experience to determine key geometric parameters. This inevitably leads to compound impeller designs for liquid ring pumps failing to adequately meet the operational requirements of liquid ring pumps, failing to fully utilize the advantages of long and short blades, and ultimately failing to significantly improve the performance and stability of liquid ring pumps. Summary of the Invention

[0004] Based on the background technology, it is necessary to propose a composite impeller design method for liquid ring pumps that can be adapted to the flow characteristics of the liquid ring pump itself, so as to improve the velocity and pressure distribution in the impeller flow channel and improve the performance and stability of the pump.

[0005] A design method for a composite impeller of a liquid ring pump, using formula (1) to design the number of impeller blades z:

[0006] (1)

[0007] Wherein:

[0008] q is the design gas flow rate of the impeller; P1 is the inlet pressure; P2 is the exhaust pressure; n is the impeller rotational speed. This is an empirical coefficient; is the exit radius of the long blade; s is the static moment of the water-passing section forming line about the axis of rotation (i.e., the axis of rotation of the blade) in the axial projection of the blade. This is a reduction factor given because the short blades have a different ability to confine the fluid flow than the long blades, and its value ranges from 0.85 to 0.9. The flow rate of the working fluid flowing through the impeller; The circumferential velocity is located at the inner surface of the liquid ring in the exhaust zone.

[0009] The above-mentioned composite impeller design method for liquid ring pumps is mainly based on the constraint relationship between the number of blades, impeller design performance, and main geometric parameters. The number of blades can be designed by trial calculations to obtain as many blades as possible. Since the number of blades can be increased as much as possible, it can play a better role in confining the fluid, thereby improving the velocity and pressure distribution in the impeller flow channel and improving the performance and stability of the pump.

[0010] In some embodiments, formula (1) is determined as follows:

[0011] Determine the theoretical head The relationship between the actual head H and the actual head H;

[0012] (1.1)

[0013] Determine the effective power w when the liquid ring does work on the gas and brings the pressure to P2:

[0014] (1.2)

[0015] Establish the relationship between the effective power w when the liquid ring does work on the gas and the pressure reaches P2, and the work done by the impeller on the liquid ring per unit time:

[0016] (1.3)

[0017] Determine the theoretical head Design formula:

[0018] (1.4)

[0019] Combining equations (1.1) and (1.3), we obtain the following relationship:

[0020] (1.5)

[0021] By combining equations (1.4) and (1.5), we obtain equation (1).

[0022] In some embodiments, the number of blades z of the impeller is designed using formula (1), including the following steps:

[0023] Step 1: Based on the given impeller design performance parameters: gas flow rate q; inlet pressure P1; exhaust pressure P2; rotational speed n, calculate the outlet radius of the long blades. The design formula is:

[0024] (2)

[0025] In the formula: It is the rotational angular velocity;

[0026] To achieve the maximum compression ratio, the following conditions must be met:

[0027] (3)

[0028] Step 2: Calculate the static moment s of the line forming the cross-section about the axis of rotation in the axial projection of the blade. The design formula is:

[0029] (4)

[0030] In the formula: The hub radius of the impeller; ; This is an empirical coefficient, ranging from 0.4 to 0.6;

[0031] Step 3: Given the cross-sectional area S1 of the long blade in the vertical plane and the cross-sectional area S2 of the short blade in the vertical plane, initially determine the blade displacement coefficient of the liquid ring pump. To calculate the flow rate of the working fluid flowing through the impeller. The design formula is:

[0032] (5)

[0033] In the formula: For the hub ratio, ;

[0034] Step 4: Based on the impeller speed n and the distance from the inner surface of the liquid ring to the impeller center during the exhaust zone... Calculate the circumferential velocity at the inner surface of the liquid ring in the exhaust zone. ;

[0035] Step 5: Calculate the number of blades z using formula (1).

[0036] In some embodiments, it also includes:

[0037] Step 6: Reverse verification of the liquid ring pump blade displacement coefficient To determine whether the design requirements are met, and to verify the rationality of the number of blades z, the formula is as follows:

[0038] (6)

[0039] Step 7: Adjust the displacement coefficient of the liquid ring pump blades Repeat steps 3 and 5 to recalculate the number of blades z;

[0040] Step 8; Repeat steps 6 and 7 until the displacement coefficient of the liquid ring pump blades is obtained according to formula (6). It meets the design requirements.

[0041] In some embodiments, the maximum value of the number of blades z is constrained using the following relationship:

[0042] (7).

[0043] In some embodiments, the inlet angle of the long blade is determined. and the exit angle of long blades The values ​​are all between 45° and 60°.

[0044] In some embodiments, the method further includes: determining the empirical coefficient. The design formula is as follows:

[0045] (8)

[0046] In some embodiments, the method further includes: determining the inlet radius of the short blade. The design formula is as follows:

[0047] (9)

[0048] In the formula:

[0049] This is the distance from the inner surface of the liquid ring to the center of the impeller during the suction zone.

[0050] In some embodiments, the The design formula is as follows:

[0051] (10)

[0052] (11)

[0053] The design formula is as follows:

[0054] (12)

[0055] (13)

[0056] (14)

[0057] In the formula: For the central angle, R is the shell radius; The ratio of the casing width to the impeller width; denoted as , where is the compression ratio at different locations within the exhaust zone; 'a' represents the blade submersion depth, which is the minimum immersion depth of the blade in the working medium. H1 is the blade shape correction factor; H1 is the height of the working fluid column.

[0058] In some embodiments, the method further includes determining the offset angle of the short blade, using the following formula:

[0059] (15)

[0060] In the formula: l is the length of the cross-sectional area formed by the water flow; This refers to the efficiency of the liquid ring pump.

[0061] In some embodiments, the formula (15) is determined as follows:

[0062] When determining the intake zone, the flow rate through the flow channel in front of the short blades is:

[0063] (15.1)

[0064] The flow rate through the rear channel of the short blades when determining the intake zone is:

[0065] (15.2)

[0066] By ensuring that the flow rates through the channels on both sides of the short blade are consistent, the following relationship is obtained:

[0067] (15.3)

[0068] Sure and The average speed is

[0069] (15.4)

[0070] The offset angle is obtained by combining equations (15.2), (15.3), and (15.4). The design formula is the aforementioned formula (15).

[0071] In some embodiments, the method further includes: determining the exit radius of the short blade. , = , where k1 takes values ​​from 0.9 to 0.95. Attached Figure Description

[0072] The accompanying drawings, which form part of this application, are used to provide a further understanding of this application. The illustrative embodiments of this application and their descriptions are used to explain this application and do not constitute an undue limitation of this application.

[0073] To more clearly illustrate the technical solutions in the embodiments of this application, the accompanying drawings used in the description of the embodiments will be briefly introduced below. Obviously, the accompanying drawings described below are only some embodiments of this application. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0074] Figure 1 This is a schematic diagram of the vertical plane of the shaft of a liquid ring pump.

[0075] Figure 2 This is a schematic diagram showing the distance from the inner surface of the liquid ring in the suction zone of this liquid ring pump to the center of the impeller.

[0076] Figure 3 This is a schematic diagram of the flow composition within the liquid ring pump channel.

[0077] Figure 4 The graph shows the mesh independence verification of the liquid ring pump designed based on the design method of this application.

[0078] Figure 5 This application presents a schematic diagram of the structure of the composite impeller of the liquid ring pump designed using the design method.

[0079] Figure 6 A comparative schematic diagram of turbulent vortex dissipation clouds at rated flow for a liquid ring pump with a liquid ring pump composite impeller designed using the design method of this application, which is a prototype pump.

[0080] Figure 7 A comparative schematic diagram of pressure cloud diagrams of a liquid ring pump with a liquid ring pump composite impeller designed using the design method of this application at rated flow rate. Detailed Implementation

[0081] To make the above-mentioned objectives, features, and advantages of this application more apparent and understandable, the specific embodiments of this application are described in detail below with reference to the accompanying drawings. Many specific details are set forth in the following description to provide a thorough understanding of this application. However, this application can be implemented in many other ways different from those described herein, and those skilled in the art can make similar modifications without departing from the spirit of this application. Therefore, this application is not limited to the specific embodiments disclosed below.

[0082] This application proposes a composite impeller design method for a liquid ring pump. The idea behind this design method is to improve the velocity and pressure distribution within the impeller flow channel and enhance the pump's performance and stability by increasing the number of blades, while minimizing the impact on the pump's inlet and outlet discharge.

[0083] Specifically, the number of blades z of the impeller is determined by formula (1):

[0084] (1)

[0085] Wherein, q is the design gas flow rate of the impeller; P1 is the inlet pressure; and P2 is the exhaust pressure. This is an empirical coefficient; is the exit radius of the long blade; s is the static moment of the water-passing section forming line about the axis of rotation (i.e., the axis of rotation of the blade) in the axial projection of the blade. This is a reduction factor given because the short blades have a different ability to confine the fluid flow than the long blades, and its value ranges from 0.85 to 0.9. The flow rate of the working fluid flowing through the impeller; The circumferential velocity is located at the inner surface of the liquid ring in the exhaust zone.

[0086] Formula (1) is the constraint relationship between the number of blades and the impeller design performance and main geometric parameters. Based on this, the number of blades can be designed. Therefore, as many blades as possible can be obtained through trial calculation. Since the number of blades can be increased as much as possible, it can play a better role in constraining the fluid, thereby improving the velocity and pressure distribution in the impeller flow channel and improving the performance and stability of the pump.

[0087] The theoretical derivation of formula (1) is given below.

[0088] The main geometric parameters of the liquid ring pump are shown in Table 1.

[0089] Table 1 Geometric parameters of liquid ring pump

[0090]

[0091] Please refer to the following: Figure 1 Understanding the above geometric parameters, Figure 1 This is a schematic diagram of the vertical plane (also known as the plane) of the liquid ring pump shaft.

[0092] When designing the impeller of a liquid ring pump, its original performance parameters are given, including at least: (1) gas flow rate q; (2) suction pressure P1; (3) exhaust pressure P2; (4) rotational speed n.

[0093] As pump theory dictates, a finite number of blades inevitably leads to relative velocity slip at the impeller outlet, resulting in a jet wake effect and a reduction in theoretical head. Pfleurde's formula shows that the theoretical head is reduced due to velocity slip. The following deviations exist between the actual head H and the actual head:

[0094] (1.1)

[0095] In the formula, s is the static moment of the line forming the cross-section of the water passage in the axial projection of the blade about the axis of rotation. This is an empirical coefficient. Based on the working principle of the liquid ring pump, the liquid ring converts the work done by the impeller on it into the pressure energy of the gas. From the theoretical head calculation formula of the pump, the work done by the impeller on the liquid ring per unit time (i.e., the power gained by the liquid ring at the impeller) is... Furthermore, according to thermodynamic formulas, the effective power w when the liquid ring does work on the gas and causes its pressure to reach P2 is:

[0096] (1.2)

[0097] In the absence of velocity slip and without considering other losses, we should have:

[0098] (1.3)

[0099] Theoretical lift It consists of two parts: one part is the power of the impeller doing work on the liquid ring when it is in the suction zone; the other part is the power of the impeller doing work on the liquid ring when it is in the discharge zone. In the suction zone, the theoretical head of the pump can be obtained from the formula:

[0100] (1.41)

[0101] The impeller outlet circumferential velocity is given in m / s. The impeller inlet velocity is denoted as , in m / s. Since the liquid ring begins its movement from the hub in the suction zone, therefore, in the formula... It is 0.

[0102] The theoretical head in the exhaust zone is:

[0103] (1.42)

[0104] In the formula Let be the circumferential velocity at the inner surface of the liquid ring in the exhaust zone, in m / s. Since the liquid ring in the intake zone undergoes circular motion after exiting the impeller and the exit angles of the blades are all the same, therefore, in equations 1.4 and 1.5... They are equal. Therefore, the theoretical head is:

[0105] (1.4)

[0106] In the exhaust zone, the liquid ring converts the energy received from the impeller into the pressure energy of the gas. Furthermore, due to the influence of the jet wake at the blade exit caused by the velocity slippage of the finite number of blades, the actual head H is less than the theoretical head. Therefore, combining equations 1.1 and 1.3, we have:

[0107] (1.5)

[0108] In the formula This is a reduction factor given because the constraint ability of short blades on the liquid flow is different from that of long blades. The value is determined by simulation to be approximately 0.85~0.9. Combining equations 1.4 and 1.5, the design formula for the number of blades z required to compress the gas and raise its pressure to P2 is the aforementioned formula (1).

[0109] Regarding the conversion factor This can be obtained through simulation (e.g., using CFX fluid dynamics software). For example, design two liquid ring pumps with identical geometries. Suppose these pumps have only long blades, eight in total, and the blade thickness is made as thin as possible. Then, cut off 50% of the long blades from the inlet of one of the pumps, leaving the rest unchanged, and perform simulation calculations. Due to the thin blades and small number of blades, the inlet displacement coefficients of the two pumps can be considered approximately equal. In the simulation results of the two pumps, the outlet pressures are different. This is mainly because the work capacity of the cut-off blades is less than that of the long blades. By analyzing the functional relationship between the outlet pressures of the two pumps and the 50% cut-off, this coefficient can be quantitatively given through simulation. .

[0110] In some embodiments, the number of blades z of the impeller is designed using formula (1), including the following steps:

[0111] Step 1: Based on the given impeller design performance parameters: gas flow rate q; inlet pressure P1; exhaust pressure P2; rotational speed n, calculate the outlet radius of the long blades. The design formula is:

[0112] (2)

[0113] In the formula: To achieve the maximum compression ratio, the following conditions must be met:

[0114] (3)

[0115] For the exit radius of long blades, the impeller exit circumferential velocity can be calculated first according to formula (2). Then, the exit radius r2 of the long blade can be calculated. The inlet radius of the long blade, i.e., the hub radius, can be taken by an empirical formula:

[0116]

[0117] In the formula This is an empirical coefficient, typically ranging from 0.4 to 0.6.

[0118] Step 2: Calculate the static moment s of the line forming the cross-section about the axis of rotation in the axial projection of the blade. The design formula is:

[0119] (4)

[0120] Based on step 1, we can see that Since r2 and r2 are both known, the static moment s can be calculated using formula (4) in this step.

[0121] Step 3: Given the cross-sectional area S1 of the long blade in the vertical plane and the cross-sectional area S2 of the short blade in the vertical plane, initially determine the blade displacement coefficient of the liquid ring pump. To calculate the flow rate of the working fluid flowing through the impeller. The design formula is:

[0122] (5)

[0123] When there is sufficient working fluid in the liquid ring pump, the flow rate of the working fluid flowing through the impeller should be equal to the flow rate of the fluid flowing through the impeller. Figure 1 Flow rate at section AB Consistency, that is, having Where B is at the hub and A is on the inner surface of the casing. The initial determination of the impeller displacement coefficient of the liquid ring pump is then performed. That is, a crowding-out coefficient is given based on experience. This allows for the trial calculation of the number of blades z.

[0124] Step 4: Based on the impeller speed n and the distance from the inner surface of the liquid ring to the impeller center during the exhaust zone... Calculate the circumferential velocity at the inner surface of the liquid ring in the exhaust zone. .

[0125] In this step, It can be determined based on empirical data; or it can be calculated based on design formulas, which will be discussed further later.

[0126] It should be noted that the calculation order of steps 2 to 4 can be flexibly adjusted and is not limited to the example above. For example, the static moment s can be calculated first, and then the flow rate can be calculated. Circumferential velocity at the inner surface of the liquid ring in the exhaust zone .

[0127] Step 5: Calculate the number of blades z using formula (1).

[0128] In this application, the displacement coefficient of the liquid ring pump blades is initially determined. Then, the number of blades z can be calculated using formula (1). After multiple trial calculations, the maximum number of blades can be obtained. Since the number of blades can be increased as much as possible, it can play a better role in confining the fluid, thereby improving the velocity and pressure distribution in the impeller channel and improving the performance and stability of the pump.

[0129] The number of blades z calculated above refers to the total number of long and short blades, which is an even number. Since long and short blades are arranged alternately in the circumferential direction, their numbers are the same, both being z / 2.

[0130] Furthermore, the design method of this application also includes:

[0131] Step 6: Reverse verification of the liquid ring pump blade displacement coefficient To determine whether the design requirements are met, and to verify the rationality of the number of blades z, the formula is as follows:

[0132] (6)

[0133] Step 7: Adjust the displacement coefficient of the liquid ring pump blades Repeat steps 3 and 5 to recalculate the number of blades z;

[0134] Step 8; Repeat steps 6 and 7 until the displacement coefficient of the liquid ring pump blades is obtained according to formula (6). It meets the design requirements.

[0135] In steps 6-8, the number of blades z calculated in step 5 is substituted into the liquid ring pump blade displacement coefficient formula (6), and then the result is compared with the displacement coefficient. If the number of blades z calculated in step 5 is within the design range, then it is reasonable. This ensures that the pump's inlet and outlet discharge is minimized while still increasing the number of blades, thus providing better fluid confinement.

[0136] In some embodiments, the design method of this application further includes:

[0137] The maximum value of the number of blades z is constrained by the following relationship:

[0138] (7)

[0139] While having as many blades as possible can improve the efficiency of a liquid ring pump, the more blades there are, the greater the impeller inlet displacement becomes. Therefore, the displacement coefficient formula for liquid ring pumps is used to constrain the number of blades z.

[0140] The design method in this application also includes: determining the inlet angle of the long blade. and the exit angle of long blades The values ​​are all between 45° and 60°. Due to the increased number of blades, the inlet and outlet angles of the long blades can be appropriately larger: 45° to 60°, while the blade thickness t is typically 0.04 to 0.05 times r². After calculating the number of blades z, the inlet angle of the long blades can be determined according to specific needs within the above range. and the exit angle of long blades The value of .

[0141] Furthermore, the design method in this application also includes: determining the empirical coefficients. The design formula is as follows:

[0142] (8)

[0143] Understandably, empirical coefficient Alternatively, it can be given directly without having to use formula (8) for calculation.

[0144] In some embodiments, the design method of this application further includes: determining the inlet radius of the short blade. The design formula is as follows:

[0145] (9)

[0146] The theoretical derivation process of formula (9) is explained below with reference to the appendix.

[0147] Based on the motion analysis of the liquid ring within the impeller, it can be concluded that in the suction zone, as... Figure 2 As shown, the distance from the inner surface of the liquid ring to the center of the impeller for:

[0148] (10)

[0149] (11)

[0150]

[0151] In the formula, R is the shell radius, in mm; The ratio of the casing width to the impeller width. The central angle is [value missing]. Similarly, in the exhaust zone, the distance from the inner surface of the liquid ring to the center of the impeller is [value missing]. for:

[0152] (12)

[0153] In the formula The compression ratio is the ratio at different locations within the exhaust zone, and it is the central angle. The function of can be calculated by equations (13) and (14):

[0154] (13)

[0155] (14)

[0156] In the formula, 'a' represents the blade immersion depth in mm, which is the minimum immersion depth of the blade in the working medium. H1 is the blade shape correction factor, typically 0.9~1.1; H1 is the height of the working fluid column, in mm, representing the pressure. It can be seen from equations (11) and (12) that... and With the central angle The inlet of the short blades should be positioned above the gas-liquid interface of the liquid ring pump to ensure better confinement of the liquid ring. However, the inlet diameter should not be too long, lest it increase inlet displacement; therefore, the inlet radius of the short blades should be... The value of should satisfy equation (9).

[0157] In the exhaust zone, the liquid ring moves in the opposite direction to the intake zone. As the liquid ring moves from the outside to the inside of the impeller in the exhaust zone, to reduce blade displacement at the impeller outlet and minimize energy loss when the liquid ring enters the impeller from the outside, a shorter blade outlet radius is used. Should be taken k1 is an empirical coefficient, and simulation results show that its value is 0.9~0.95.

[0158] In this application, in order to reduce the impact loss at the inlet of the short blade, the inlet angle of the short blade is... and the exit angle of this short blade Should be with long leaves and The inlet angle or outlet angle at the truncated circle is the same.

[0159] In some embodiments, the design method of this application further includes: determining the offset angle of the short blade, using the following formula:

[0160] (15)

[0161] The theoretical derivation of formula (15) will be explained below with reference to the attached diagram.

[0162] Due to the inertial force of the fluid and the limited ability of the blades to constrain the fluid, the velocity and pressure distribution in the impeller flow channel of the liquid ring pump is uneven. Therefore, the offset of the blades must be considered when designing short blades.

[0163] According to the velocity slip theory with a finite number of blades, the flow inside the impeller channel can be approximated as an axial vortex flow (e.g., ...). Figure 3 (as shown in region a) and the cross-flow passing through the stationary impeller (as shown in region a) Figure 3 The vector sum is shown in region b). The impeller of the liquid ring pump of this application includes, but is not limited to, forward-curved blades, radial straight blades, and backward-curved blades. In the suction zone, the liquid ring moves outward from the impeller, and the working liquid velocity at the suction surface of the blades is higher, and its internal flow is composed of... Figure 3As shown in region c; in the exhaust zone, the liquid ring moves into the impeller, and the working fluid velocity is higher at the blade pressure surface, where the internal flow is combined into... Figure 3 As shown in region d. Therefore, when designing the short blade offset, at the short blade inlet, the short blade should be offset towards the suction surface; at the short blade outlet, the short blade should be offset towards the pressure surface, with equal offset angles. Since the liquid ring flows in opposite directions in the suction and exhaust zones and requires equal offset angles, the short blade profile does not need to be twisted.

[0164] The offset angle is defined as the line connecting the midpoint of the inlet edge of the short blade and the impeller center, and the angle between two adjacent long blades. The angle between the midpoint of the truncated circle and the line connecting the center of the impeller, such as Figure 1 As shown in θ. Starting from the angle that keeps the flow rates on both sides of the short blade essentially the same, the flow rate through the left-side flow channel C (the front flow channel, located in front of the short blade in the rotational direction) in the intake region can be:

[0165] (15.1)

[0166] In the formula The average axial velocity component of the liquid flow within channel C (m·s) -1 Let the flow rate through the flow channel D (the rear flow channel, located behind the short blade in the rotational direction) on the right side of the short blade be:

[0167] (15.2)

[0168] In the formula Let be the average axial velocity component of the fluid flow within channel D, in m·s. -1 If the flow rates through the channels on both sides of the short blade are made the same, then:

[0169] (15.3)

[0170] Since the flow rate through the channel between the two long blades is constant, the average axial velocity of the liquid flow between the two long blades can be calculated. At the same time, it can be approximated as and The average speed is Right now:

[0171] (15.4)

[0172] In the formula, l is the length of the cross-section formed by the water flow, in mm; For the efficiency of the liquid ring pump, an estimated value is taken based on experience. Then, by combining equations (15.2), (15.3), and (15.4), the offset angle can be obtained. The design formula is the aforementioned formula (15).

[0173] In the exhaust zone, the working fluid velocity at the blade pressure surface is higher, and the short blades are biased towards the pressure surface, with the bias angle being the same as in the intake zone.

[0174] Experimental and Simulation Verification Description

[0175] Based on the 2BEA-203 liquid ring pump, i.e., the prototype impeller liquid ring pump (prototype pump), a composite impeller liquid ring pump designed using the above method was tested and compared. The parameters of the two pumps are shown in Table 2.

[0176] Table 2

[0177]

[0178] The liquid ring pump composite impeller designed using the above method is as follows: Figure 5 As shown, the impeller of the prototype pump has no short blades.

[0179] Performance testing was conducted using a liquid ring pump performance testing bench. The bench mainly consists of a liquid ring pump, suction line, discharge line, and replenishment line. The suction line is equipped with a flow meter, inlet regulating valve, and inlet pressure sensor to monitor and regulate the inlet flow rate. The outlet line is equipped with an outlet check valve and pressure sensor to monitor the outlet pressure.

[0180] To test grid independence, different grid numbers were used to test the inlet pressure and efficiency of the liquid ring pump at its rated flow rate. The results are as follows: Figure 5 As shown, when the number of grids is above 1.7 million, the vacuum degree and efficiency of the liquid ring pump change little with the number of grids. Taking into account both the calculation accuracy and the amount of calculation, the number of grids for the composite impeller and the prototype impeller liquid ring pump models are determined to be 2.6 million and 2.803 million, respectively.

[0181] Simulations of a composite impeller liquid ring pump and a prototype impeller liquid ring pump were performed using Fluent, employing the VOF gas-liquid two-phase flow model, which has strong free interface capture capabilities. The RNG k-ε turbulence model was used, and the PISO pressure-velocity coupled algorithm was selected for the solution. A no-slip wall was used, with a gas phase inlet and mass inlet and pressure outlet boundary conditions. Inlet flow rates were 0.02, 0.035, and 0.05 kg / s, respectively. The time step was 1×10⁻⁶. -5 After the impeller rotates 4 revolutions, the inlet vacuum of the pump tends to stabilize, and it can be considered that the numerical calculation of the internal flow of the liquid ring pump tends to converge.

[0182] (I) Comparison of external characteristics

[0183] The experimentally tested liquid ring pump achieved a speed of 0.02 kg·s. -10.035 kg·s -1 and 0.05 kg·s -1 The efficiencies under different operating conditions (caused by different inlet flow rates) were 6.3%, 11.9%, and 11.8%, respectively; the inlet pressures were 26.4 kPa, 46.2 kPa, and 60.3 kPa, respectively. Table 3 compares the performance of the prototype impeller liquid ring pump and the composite impeller liquid ring pump obtained from the simulation. Since axial and radial clearance leakage were not considered in the numerical simulation, the vacuum degree and efficiency obtained from the numerical simulation are higher than the experimental values.

[0184] Table 3 shows that the trends in inlet pressure and efficiency obtained from the experimental tests and numerical simulations are consistent. Since the numerical simulations did not consider axial clearance leakage between the impeller and the pump casing, the simulated values ​​for vacuum and efficiency of both the prototype pump and the composite impeller liquid ring pump are higher than the experimental values. The inlet pressure of both the composite impeller liquid ring pump and the prototype impeller liquid ring pump increases with increasing flow rate; the efficiency also increases with increasing flow rate. Furthermore, the efficiency and vacuum of the composite impeller liquid ring pump are significantly better than those of the prototype impeller liquid ring pump at all operating points. (0.02, 0.035, and 0.05 kg·s) -1 Under operating conditions, the efficiency of the composite impeller was increased by 2, 1.4 and 4.3 percentage points respectively compared with the prototype impeller; the vacuum level was increased by 4.12, 3.1 and 2.21 kPa respectively.

[0185] Table 3 Comparison of numerical simulation performance of liquid ring pumps

[0186]

[0187] (II) Comparison of internal flow fields

[0188] Figure 6 The diagram illustrates the distribution of turbulent vortex dissipation intensity at the axial mid-section of two liquid ring pumps under rated operating conditions after 0.16 s of rotation. The left side represents the composite impeller liquid ring pump of this application; the right side represents the prototype pump.

[0189] Depend on Figure 6 It can be seen that the turbulent vortex dissipation in the internal flow field of the two types of liquid ring pumps is mainly concentrated in the gas phase region, the pump casing, and the blade outlet. The stronger turbulent vortex dissipation at the pump casing is mainly due to the dynamic and static interference between the pump casing and the liquid ring. The stronger turbulent vortex dissipation at the blade outlet is mainly due to the stronger exit jet wake and tip vortex disturbance at the blade outlet during the intake region. In the exhaust region, the turbulent vortex dissipation at the blade outlet is mainly due to liquid flow impact losses.

[0190] Depend on Figure 6 It can be seen that the turbulent vortex dissipation intensity of the composite impeller liquid ring pump is lower than that of the prototype pump in the gas phase region, at the pump casing, and at the blade outlet. Figure 6The A portion on the left is smaller than the A portion on the right. This indicates that the design of the composite impeller effectively reduces the jet wake effect at the blade outlet and improves the flow distribution within the pump channel.

[0191] Figure 7 The diagram illustrates the pressure distribution of the axial mid-section of two liquid ring pumps under rated operating conditions after 0.16 s of rotation. The left side shows the composite impeller liquid ring pump of this application; the right side shows the prototype pump.

[0192] Depend on Figure 7 It can be seen that the pressure in both types of liquid ring pumps gradually increases radially from the hub, reaching its highest point at the exhaust casing. The pressure gradient within the impeller flow channel of the composite impeller liquid ring pump is smaller than that of the prototype impeller, as shown in the figure, where the pressure change is a gradual increase radially rather than an abrupt change (reflected in the gradual color change); this indicates that the composite impeller helps improve the pressure distribution within the impeller flow channel.

[0193] The technical features of the above embodiments can be combined in any way. For the sake of brevity, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0194] The above embodiments merely illustrate several implementation methods of this application, and while the descriptions are relatively specific and detailed, they should not be construed as limiting the scope of the patent application. It should be noted that those skilled in the art can make numerous modifications and improvements without departing from the concept of this application, and these all fall within the protection scope of this application. Therefore, the protection scope of this patent application should be determined by the appended claims.

Claims

1. A method for designing a composite impeller for a liquid ring pump, characterized in that, The number of blades z of the impeller is designed using formula (1): （1） Wherein: q is the design gas flow rate of the impeller; P1 is the inlet pressure; P2 is the exhaust pressure; This is an empirical coefficient; denoted as the exit radius of the long blade; s is the static moment of the water-passing section forming line about the axis of rotation in the axial projection of the blade. This is a reduction factor given because the short blades have a different ability to confine the fluid flow than the long blades, and its value ranges from 0.85 to 0.

9. The flow rate of the working fluid flowing through the impeller; The circumferential velocity at the inner surface of the liquid ring in the exhaust zone; The number of blades z of the impeller is designed using formula (1), including the following steps: Step 1: Based on the given impeller design performance parameters: gas flow rate q; inlet pressure P1; exhaust pressure P2; rotational speed n, calculate the outlet radius of the long blades. The design formula is: （2） In the formula: It is the rotational angular velocity; To achieve the maximum compression ratio, the following conditions must be met: （3） Step 2: Calculate the static moment s of the line forming the cross-section about the axis of rotation in the axial projection of the blade. The design formula is: （4） In the formula: The hub radius of the impeller; ; This is an empirical coefficient, ranging from 0.4 to 0.6; Step 3: Given the cross-sectional area S1 of the long blade in the vertical plane and the cross-sectional area S2 of the short blade in the vertical plane, initially determine the blade displacement coefficient of the liquid ring pump. To calculate the flow rate of the working fluid flowing through the impeller. The design formula is: （5） In the formula: For the hub ratio, b is the impeller width; Step 4: Based on the impeller speed n and the distance from the inner surface of the liquid ring to the impeller center during the exhaust zone... Calculate the circumferential velocity at the inner surface of the liquid ring in the exhaust zone. ; Step 5: Calculate the number of blades z using formula (1).

2. The method according to claim 1, characterized in that, Formula (1) is determined in the following way: Determine the theoretical head The relationship between the actual head H and the actual head H; （1.1） Determine the effective power w when the liquid ring does work on the gas and brings the pressure to P2: （1.2） Establish the relationship between the effective power w when the liquid ring does work on the gas and the pressure reaches P2, and the work done by the impeller on the liquid ring per unit time: （1.3） Determine the theoretical head Design formula: （1.4） Combining equations (1.1) and (1.3), we obtain the following relationship: （1.5） By combining formulas (1.4) and (1.5), we obtain the formula (1).

3. The method according to claim 1, characterized in that, Also includes: Step 6: Reverse verification of the liquid ring pump blade displacement coefficient To determine whether the design requirements are met, and to verify the rationality of the number of blades z, the formula is as follows: （6） Step 7: Adjust the displacement coefficient of the liquid ring pump blades Repeat steps 3 and 5 to recalculate the number of blades z; Step 8; Repeat steps 6 and 7 until the displacement coefficient of the liquid ring pump blades is obtained according to formula (6). It meets the design requirements.

4. The method according to claim 1 or 3, characterized in that, It also includes: constraining the maximum value of the number of blades z, using the following relationship: （7）。 5. The method according to claim 1, characterized in that, Determine the inlet angle of the long blade and the exit angle of long blades The values ​​are all between 45° and 60°.

6. The method according to claim 1, characterized in that, The method further includes: determining the inlet radius of the short blade. The design formula is as follows: （9） In the formula: This is the distance from the inner surface of the liquid ring to the center of the impeller during the suction zone; This is the distance from the inner surface of the liquid ring to the center of the impeller in the exhaust zone.

7. The method according to claim 6, characterized in that, The The design formula is as follows: （10） （11） The design formula is as follows: （12） （13） （14） In the formula: For the central angle, R is the shell radius; The ratio of the casing width to the impeller width; This represents the compression ratio at different locations within the exhaust zone; 'a' represents the blade submersion depth, which is the minimum immersion depth of the blade in the working medium. H1 is the blade shape correction factor; H1 is the height of the working fluid column.

8. The method according to claim 6, characterized in that, The method also includes determining the offset angle of the short blade, using the following formula: （15） In the formula: l is the length of the cross-sectional area formed by the water flow; This refers to the efficiency of the liquid ring pump.

9. The method according to claim 8, characterized in that, The formula (15) is determined in the following way: The flow rate through the front channel of the short blades when determining the intake zone The formula is: （15.1） The flow rate passing through the rear channel of the short blades when determining the intake zone The formula is: （15.2） By ensuring that the flow rates through the channels on both sides of the short blade are consistent, the following relationship is obtained: （15.3） Sure and The average speed is （15.4） The offset angle is obtained by combining equations (15.2), (15.3), and (15.4). The design formula is the aforementioned formula (15): In the formula: The average axial velocity component of the fluid flow within the flow channel on the front side of the short blade; The average axial velocity component of the liquid flow within the flow channel behind the short blade.

10. The method according to claim 6, characterized in that, The method further includes: determining the exit radius of the short blade. , = , where k1 takes values ​​from 0.9 to 0.95.

Images