Parameter setting method for fuel cell air compressor based on load excitation amplitude analysis

Abstract

The application discloses a fuel cell air compressor parameter setting method based on load excitation amplitude analysis, which comprises the following steps: S1, constructing a super-speed electric air compressor numerical equation and a load excitation model; S2, using the load excitation model in step S1 to simulate under different primary guide vane angles to obtain a first simulation result of the load excitation amplitude; S3, using the load excitation model in step S1 to simulate under different secondary guide vane angles to obtain a second simulation result of the load excitation amplitude; S4, selecting a primary guide vane angle corresponding to the minimum load excitation amplitude in the first simulation result as a first parameter according to the first simulation result in step S2; and S5, selecting a secondary guide vane angle corresponding to the minimum load excitation amplitude in the second simulation result as a second parameter according to the second simulation result in step S3. The application can reduce the fluctuation of the super-speed electric air compressor in the speed regulation process.

Description

Technical Field

[0001] This invention relates to the field of ultra-high-speed electric air compressor technology, specifically to a method for setting parameters of a fuel cell air compressor based on load excitation amplitude analysis. Background Technology

[0002] Fuel cell vehicles (FCVs) offer advantages such as high energy efficiency, zero emissions, and zero pollution, making them a crucial direction for the development of new energy vehicles. The electric air compressor in a fuel cell vehicle compresses air and delivers it to the fuel cell stack. To meet the high power requirements of fuel cell stacks, it is necessary to further increase the air flow rate and pressure ratio of the electric air compressor. Two-stage centrifugal ultra-high-speed electric air compressors (hereinafter referred to as "ultra-high-speed electric air compressors") have attracted significant research attention due to their technological potential to provide greater flow rate and pressure ratio. To better track the output power requirements of the fuel cell stack, the ultra-high-speed electric air compressor needs to adjust its speed to tens of thousands or even hundreds of thousands of revolutions per second. However, the wide speed range adjustment of ultra-high-speed electric air compressors can cause drastic fluctuations, increasing the adjustment time, limiting the power output of the fuel cell stack, and severely affecting the power performance of fuel cell vehicles.

[0003] How to reduce the fluctuations generated when adjusting the speed of ultra-high speed electric air compressors over a wide speed range is one of the important problems that urgently need to be solved in this field. Summary of the Invention

[0004] The purpose of this invention is to provide a method for setting parameters of a fuel cell air compressor based on load excitation amplitude analysis, so as to overcome the shortcomings of the prior art. It can reduce the fluctuations generated by the ultra-high speed electric air compressor during speed regulation.

[0005] This invention provides a method for setting parameters of a fuel cell air compressor based on load excitation amplitude analysis, for use in an ultra-high speed electric air compressor in a fuel cell vehicle. The ultra-high speed electric air compressor is a two-stage centrifugal electric air compressor, which has a first-stage impeller and a second-stage impeller.

[0006] This includes the following steps:

[0007] S1. Construct the numerical equations and load excitation model of the ultra-high speed electric air compressor;

[0008] S2. Using the load excitation model in step S1, simulations are performed under different first-stage guide vane angles to obtain the first simulation results of the load excitation amplitude.

[0009] S3. Using the load excitation model in step S1, simulations are performed under different secondary guide vane angles to obtain the second simulation results of the load excitation amplitude.

[0010] S4. Based on the first simulation result in step S2, select the first-stage guide vane angle corresponding to the minimum load excitation amplitude in the first simulation result as the first parameter.

[0011] S5. Based on the second simulation results in step S3, select the second guide vane angle corresponding to the minimum load excitation amplitude in the second simulation results as the second parameter.

[0012] The fuel cell air compressor parameter setting method based on load excitation amplitude analysis described above may optionally include the following steps:

[0013] S6. Using the load excitation model in step S1, simulations are performed under different first-stage impeller installation angles to obtain the third simulation results of the load excitation amplitude.

[0014] S7. Using the load excitation model in step S1, simulations are performed under different second-stage impeller installation angles to obtain the fourth simulation result of the load excitation amplitude.

[0015] S8. Based on the third simulation result in step S6, select the first-stage impeller installation angle corresponding to the minimum load excitation amplitude in the third simulation result as the third parameter.

[0016] S9. Based on the fourth simulation result in step S7, select the second-stage impeller installation angle corresponding to the minimum load excitation amplitude in the fourth simulation result as the fourth parameter.

[0017] S10 uses the first, second, third, and fourth parameters as the design parameters for the ultra-high-speed electric air compressor.

[0018] The fuel cell air compressor parameter setting method based on load excitation amplitude analysis described above may optionally include the following steps in step S1:

[0019] S11, constructing the numerical equations for an ultra-high-speed electric air compressor;

[0020] S12, construct the numerical equations for the permanent magnet synchronous motor;

[0021] S13, Construct a load excitation model for an ultra-high-speed electric air compressor;

[0022] The formula corresponding to the load incentive model is:

[0023]

[0024] Where, ΔT L For the load excitation of the ultra-high speed electric air compressor; Q is the mass flow rate, n is the impeller speed of the ultra-high speed electric air compressor, and P is the power of the ultra-high speed electric air compressor;

[0025]

[0026]

[0027]

[0028]

[0029]

[0030] Where α1 is the inlet angle of the first-stage impeller; β1 is the installation angle of the first-stage impeller; b is the number of impeller blades; D1 is the nominal diameter of the first-stage impeller of the ultra-high-speed electric air compressor; α2 is the inlet angle of the second-stage impeller; β2 is the installation angle of the second-stage impeller; D2 is the nominal diameter of the second-stage impeller of the ultra-high-speed electric air compressor; t represents time; λ c ν1 is the loss coefficient, representing the ratio of leakage flow rate to total flow rate; ν2 is the outlet gas velocity; g is the gravitational acceleration; Z1 is the gas inlet height; Z2 is the gas outlet height; λ is the frictional resistance coefficient, which is related to the flow Reynolds number and the relative roughness of the volute wall; l is the flow path length or the length of the average streamline; c is the average airflow velocity in the volute channel; d... h S1 is the average equivalent diameter of the volute flow channel; S2 is the outlet cross-sectional area; p1 is the inlet gas pressure; p2 is the inlet and outlet gas pressure.

[0031] The parameter setting method for fuel cell air compressor based on load excitation amplitude analysis, as described above, optionally assumes that the fluid inside the ultra-high-speed electric air compressor is continuous, and that the impeller and flow channel walls are rigid walls whose physical shapes do not change with the force applied.

[0032] In the above-described method for setting parameters of a fuel cell air compressor based on load excitation amplitude analysis, the load excitation model can optionally be a model that considers impeller inlet impact loss, gas leakage loss, wheel resistance loss, and friction loss.

[0033] In the fuel cell air compressor parameter setting method based on load excitation amplitude analysis as described above, step S2 optionally includes the following steps:

[0034] S21, set the first-stage guide vane angle to different angles, and simulate the actual operation of the over-adjustment electric air compressor according to the set speed, inlet pressure and inlet humidity;

[0035] S22, the time step is selected as the time required for the impeller to rotate 1°;

[0036] S23, first perform steady-state calculations until the calculations converge;

[0037] S24, perform transient calculations for three impeller rotations to obtain the numerically calculated flow field and first load excitation data.

[0038] In the fuel cell air compressor parameter setting method based on load excitation amplitude analysis as described above, step S4 optionally includes the following steps:

[0039] S41, obtain the streamline diagram of the first-stage guide vane and the pressure cloud diagram of the first-stage impeller surface after numerical calculation of guide vanes with different inclination angles;

[0040] S42, the optimal first-stage guide vane angle is selected as the first parameter by using the pressure cloud map of the first-stage impeller surface.

[0041] Compared with existing technologies, this invention constructs numerical equations and a load excitation model for an ultra-high-speed electric air compressor, using the load excitation amplitude as the target value, and selects parameters for the first-stage guide vane angle, second-stage guide vane angle, first-stage impeller mounting angle, and second-stage impeller mounting angle. The parameters selected using this method can reduce the load excitation amplitude, which is beneficial for reducing vibration of the ultra-high-speed air compressor during operation. Attached Figure Description

[0042] Figure 1 This is a flowchart of the steps of the present invention;

[0043] Figure 2 This is a flowchart illustrating the specific steps of step S1 of the present invention;

[0044] Figure 3 This is a flowchart illustrating the specific steps of step S2 in this invention;

[0045] Figure 4 This is a flowchart illustrating the specific steps of step S4 in this invention;

[0046] Figure 5 This is a schematic diagram of the triangular cell proposed in Embodiment 2 of the present invention;

[0047] Figure 6 This is a schematic diagram of the velocity triangle at the impeller inlet proposed in Embodiment 2 of the present invention;

[0048] Figure 7 This is a schematic diagram illustrating the gas leakage loss principle of Embodiment 2 of the present invention;

[0049] Figure 8 This is a schematic diagram of the gas wheel resistance loss principle in Embodiment 2 of the present invention;

[0050] Figure 9 This is a schematic diagram illustrating the principle of frictional loss during gas flow in a pipeline.

[0051] Figure 10This is a schematic diagram of the impeller streamline and surface pressure at the first-stage guide vane angle obtained in Embodiment 2 of the present invention;

[0052] Figure 11 This is a graph showing the variation of load excitation amplitude under the first-stage guide vane angle obtained in Embodiment 2 of the present invention;

[0053] Figure 12 This is a schematic diagram of the impeller streamline and surface pressure at the second-stage guide vane angle obtained in Embodiment 2 of the present invention;

[0054] Figure 13 This is a graph showing the variation of load excitation amplitude under the tilt angle of the second-stage guide vane obtained in Embodiment 2 of the present invention;

[0055] Figure 14 This is a schematic diagram of the impeller streamline and surface pressure at the installation angle of the first-stage impeller obtained in Embodiment 2 of the present invention;

[0056] Figure 15 This is a graph showing the variation of load excitation amplitude under the installation angle of the first-stage impeller obtained in Embodiment 2 of the present invention;

[0057] Figure 16 This is a schematic diagram of the impeller streamline and surface pressure at the installation angle of the secondary impeller obtained in Embodiment 2 of the present invention;

[0058] Figure 17 This is a graph showing the change in load excitation amplitude under the installation angle of the secondary impeller obtained in Embodiment 2 of the present invention. Detailed Implementation

[0059] The embodiments described below with reference to the accompanying drawings are exemplary and are only used to explain the present invention, and should not be construed as limiting the present invention.

[0060] Ultra-high-speed electric air compressors operate at speeds of tens of thousands or even hundreds of thousands of revolutions per minute, supplying air to fuel cell stacks through this high-speed rotation. During this high-speed rotation, especially when adjusting speed over a wide range, significant fluctuations occur. Research has revealed that these fluctuations are primarily influenced by load excitation and electromagnetic excitation. Therefore, this application aims to reduce these fluctuations during speed adjustment by optimizing the design parameters of the ultra-high-speed electric air compressor itself. Please refer to Examples 1 and 2 below for details.

[0061] Example 1

[0062] Please refer to Figure 1 This embodiment proposes a parameter setting method for fuel cell air compressors based on load excitation amplitude analysis, which is used for ultra-high speed electric air compressors in fuel cell vehicles. The ultra-high speed electric air compressor is a two-stage centrifugal electric air compressor, which has a first-stage impeller and a second-stage impeller.

[0063] Specifically, it includes the following steps:

[0064] S1. Construct the numerical equations and load excitation model of the ultra-high speed electric air compressor. The numerical equations of the ultra-high speed electric air compressor are used for numerical calculation, and the load excitation model is used to simulate the changes in load, so as to select the optimal design parameters according to the load excitation amplitude.

[0065] In implementation, step S1 includes the following steps:

[0066] S11, constructing the numerical equations for an ultra-high-speed electric air compressor;

[0067] S12, construct the numerical equations for the permanent magnet synchronous motor;

[0068] S13, Construct a load excitation model for an ultra-high-speed electric air compressor;

[0069] The formula corresponding to the load incentive model is:

[0070]

[0071] Where, ΔT L For the load excitation of the ultra-high speed electric air compressor; Q is the mass flow rate, n is the impeller speed of the ultra-high speed electric air compressor, and P is the power of the ultra-high speed electric air compressor;

[0072]

[0073]

[0074]

[0075]

[0076]

[0077] Where α1 is the inlet angle of the first-stage impeller; β1 is the installation angle of the first-stage impeller; b is the number of impeller blades; D1 is the nominal diameter of the first-stage impeller of the ultra-high-speed electric air compressor; α2 is the inlet angle of the second-stage impeller; β2 is the installation angle of the second-stage impeller; D2 is the nominal diameter of the second-stage impeller of the ultra-high-speed electric air compressor; t represents time; λ c ν1 is the loss coefficient, representing the ratio of leakage flow rate to total flow rate; ν2 is the outlet gas velocity; g is the gravitational acceleration; Z1 is the gas inlet height; Z2 is the gas outlet height; λ is the frictional resistance coefficient, which is related to the flow Reynolds number and the relative roughness of the volute wall; l is the flow path length or the length of the average streamline; c is the average airflow velocity in the volute channel; d... hS1 is the average equivalent diameter of the volute flow channel; S2 is the outlet cross-sectional area; p1 is the inlet gas pressure; p2 is the inlet and outlet gas pressure.

[0078] Furthermore, the load excitation model is a model that considers impeller inlet impact loss, gas leakage loss, impeller resistance loss, and friction loss.

[0079] S2, using the load excitation model from step S1, simulations are performed under different first-stage guide vane angles to obtain the first simulation results of the load excitation amplitude. The first simulation results include the load excitation amplitudes corresponding to multiple different first-stage guide vane angles. In this step, all parameters and conditions are the same except for the first-stage guide vane angle during multiple simulations.

[0080] Step S2 includes the following steps:

[0081] S21, set the first-stage guide vane angle to different angles, and simulate the actual operation of the over-adjustment electric air compressor according to the set speed, inlet pressure and inlet humidity;

[0082] S22, the time step is selected as the time required for the impeller to rotate 1°;

[0083] S23, first perform steady-state calculations until the calculations converge;

[0084] S24, perform transient calculations for three impeller rotations to obtain the numerically calculated flow field and first load excitation data.

[0085] S3. Using the load excitation model in step S1, simulations are performed under different secondary guide vane angles to obtain the second simulation results of the load excitation amplitude. The second simulation results include the load excitation amplitudes corresponding to multiple different secondary guide vane angles.

[0086] S4. Based on the first simulation result in step S2, select the first-stage guide vane angle corresponding to the minimum load excitation amplitude in the first simulation result as the first parameter; that is, the first parameter is the design parameter of the first-stage guide vane angle.

[0087] Step S4 includes the following steps:

[0088] S41, obtain the streamline diagram of the first-stage guide vane and the pressure cloud diagram of the first-stage impeller surface after numerical calculation of guide vanes with different inclination angles; S42, select the optimal first-stage guide vane angle as the first parameter based on the pressure cloud diagram of the first-stage impeller surface.

[0089] S5. Based on the second simulation results in step S3, select the second-stage guide vane angle corresponding to the minimum load excitation amplitude in the second simulation results as the second parameter. The second parameter is the design parameter for the second-stage guide vane angle.

[0090] In practical research, it was found that the amplitude variation of the load excitation is affected not only by the angles of the first-stage and second-stage guide vanes, but also by the installation angles of the first-stage and second-stage impellers. Therefore, this application makes further improvements:

[0091] A preferred implementation also includes the following steps:

[0092] S6. Using the load excitation model from step S1, simulations are performed under different first-stage impeller installation angles to obtain the third simulation results of the load excitation amplitude. The third simulation results include the load excitation amplitudes corresponding to multiple different first-stage impeller installation angles.

[0093] S7. Using the load excitation model in step S1, simulations are performed under different second-stage impeller installation angles to obtain the fourth simulation result of the load excitation amplitude. The fourth simulation result includes multiple load excitation amplitudes corresponding to different second-stage impeller installation angles.

[0094] S8. Based on the third simulation result in step S6, select the first-stage impeller mounting angle corresponding to the minimum load excitation amplitude in the third simulation result as the third parameter; that is, the third parameter is the design parameter of the first-stage impeller mounting angle.

[0095] S9. Based on the fourth simulation result in step S7, select the second-stage impeller mounting angle corresponding to the minimum load excitation amplitude in the fourth simulation result as the fourth parameter; that is, the fourth parameter is the design parameter of the second-stage impeller mounting angle.

[0096] S10 uses the first, second, third, and fourth parameters as the design parameters for the ultra-high-speed electric air compressor.

[0097] By designing according to the first, second, third, and fourth parameters determined in the above manner, the amplitude of drastic fluctuations in the speed range of the ultra-high-speed electric air compressor can be reduced.

[0098] In specific implementation, the assumptions in the construction of the load excitation model are that the fluid inside the ultra-high-speed electric air compressor is continuous, and the impeller and flow channel walls are rigid walls whose physical shapes do not change with the force applied.

[0099] Example 2

[0100] Ultra-high-speed electric air compressors do not have an internal compression process; instead, they increase gas pressure by altering the gas flow state through impeller rotation. Simply applying the first law of thermodynamics cannot represent the complex airflow characteristics between the blades of an ultra-high-speed electric air compressor. Since the object of study is compressible air with heat transfer, the main considerations are the conversion of kinetic and internal energy of the air. Therefore, the solution is simplified as follows: A steady-state analysis of the ultra-high-speed electric air compressor is performed, neglecting gas pulsation and changes in gas gravity. The governing equations of the ultra-high-speed electric air compressor are shown below:

[0101] (1) Continuity Equation The essence of the continuity equation is the law of conservation of mass. For the internal flow passage components in an ultra-high-speed electric air compressor, under the assumption of one-dimensional steady flow, the continuity equation is expressed as follows: Under good sealing conditions of the flow passage components and no external mass exchange, the mass flow rate along each cross-section of the flow channel remains conserved; that is, the increase in mass of a certain infinitesimal volume per unit time is equal to the net mass of fluid entering that infinitesimal volume per unit time. The expression form of the continuity equation is:

[0102]

[0103] Expanding the above equation in cylindrical coordinates, we get:

[0104]

[0105] In the formula, t is time (s); ρ is density (kg / m³). 3 u is the component of the velocity vector in the x-direction, m / s; v is the component of the velocity vector in the y-direction, m / s; w is the component of the velocity vector in the z-direction, m / s.

[0106] (2) Momentum Equation The momentum equation is essentially an application of Newton's second law in fluid mechanics. Specifically, it states that the rate of change of the momentum of a fluid in a infinitesimal volume with respect to time is equal to the sum of all forces exerted on the fluid in that volume. The momentum equation in cylindrical coordinates is expressed as:

[0107]

[0108] In the formula, p is the pressure acting on the infinitesimal volume, Pa; τ xy ,τ yz ,τ zx "Equal" refers to the components of viscous stress generated on a micro-element volume due to the viscosity of molecules, expressed in Pa and F. x ,F y ,F z Let N be the body force on the infinitesimal volume.

[0109] (3) Energy Equation The energy equation explains the change in the energy level of the gas after the impeller performs work on it, from the perspective of energy transformation. Similarly, under the basic assumption of one-dimensional steady flow, the flow process of gas in the flow section of the centrifugal booster belongs to a steady flow open system in thermodynamics, and the mass continuity equation also applies. The energy equation shows that the work done by the electromagnetic torque impeller on a unit mass of gas is equal to the sum of the increase in kinetic energy and enthalpy of the gas. The energy equation is mainly used to calculate and analyze the overall energy and temperature changes of different flow sections in ultra-high-speed electric air compressors. The energy equation is in the following form:

[0110]

[0111] In the formula, T is temperature, K; c p It is the specific heat capacity under specified pressure, J / (kg·K); S T This refers to the dissipation term of viscosity.

[0112] (4) Angular momentum conservation equation

[0113] The prerequisite for the conservation of angular momentum is that the stress tensor is symmetric. The equation for the conservation of angular momentum is expressed as follows:

[0114] σ=σ T (5)

[0115] In the formula, σ is the stress tensor, Pa.

[0116] (5) Equations of State: In most cases, the partial differential equations of a mathematical model are not closed and contain additional unknowns. To close the equation set, additional equations need to be added to the mathematical model. The aforementioned mass continuity equation, momentum equation, and energy equation are the necessary basic equations for analyzing the internal flow state of an ultra-high-speed electric air compressor. An ideal gas equation of state from engineering thermodynamics is also added to further calculate the changes in gas state parameters and close the aforementioned equation set. The equations of state are expressed as follows:

[0117]

[0118] In the formula, R is the molar gas constant, J / (mol·k).

[0119] A turbulence model is introduced to close the Naiver-Stokes equations. When the fluid velocity is low, the flow is stratified, with momentum exchange between layers occurring only due to molecular thermal motion; irregular pulsations in the flow path are absent. However, as the fluid velocity increases and reaches a certain level, the flow path becomes chaotic and irregular, i.e., turbulent. The internal structure of ultra-high-speed electric air compressors, especially the impeller blades, is extremely complex, and airflow within them is mostly turbulent. Therefore, selecting a suitable turbulence model is crucial for obtaining accurate calculation results. Numerical calculations handle turbulence problems based on scale differences; current mainstream methods include Reynolds-averaged flow (RANS), Large Eddy Simulation (LES), and Direct Numerical Simulation (DNS).

[0120] This embodiment uses the Reynolds-averaged Naiver-Stokes (RANS) turbulence model to numerically solve the internal flow of the ultra-high-speed electric air compressor. Since the influence of turbulent fluctuations—i.e., Reynolds stress—in the Reynolds-averaged Naiver-Stokes equations is unknown, this application compensates for this deficiency by introducing a turbulence model for solution.

[0121] The k-omega model does not involve complex nonlinear decay functions, resulting in more accurate results and better convergence, especially in numerical calculations near the wall at low Reynolds numbers, where it has unique advantages. The k-omega model makes assumptions about the relationship between turbulent viscosity, kinetic energy, and turbulent frequency; the specific relationships are as follows:

[0122]

[0123] In the formula μ t ω is the turbulent viscosity, Pa·s; k is the turbulent kinetic energy, J / kg; t , where is the turbulence frequency, in Hz.

[0124] The k-omega model cannot accurately predict the onset of turbulence, nor can it predict the total amount of air detached from the wall. However, the k-omega model based on the SST model can easily solve these problems. Extremely high adverse pressure gradients often exist inside the ultra-high-speed electric air compressor of a fuel cell, especially when far from the design operating conditions, making flow separation a common phenomenon. Therefore, the accuracy of the solution for the boundary layer flow becomes crucial. The k-omega model based on the SST model combines the advantages of the k-epsilon and k-omega models, providing excellent simulation results for wall flow separation. Therefore, this embodiment uses the k-omega model based on the SST model to characterize the turbulent flow inside the ultra-high-speed electric air compressor. The specific equations of the k-omega model based on the SST model are as follows:

[0125]

[0126] In the formula P k The term represents the kinetic energy generation of turbulence; D w This represents cross-diffusion. The specific representations of these two terms are as follows:

[0127]

[0128] Another SST turbulence model takes the following form:

[0129] Φ3=F1Φ1+(1-F)Φ2 (10)

[0130] In the formula, Φ1 and Φ2 represent the transformed k-epsilon and k-omega equations, respectively. The expression for the mixing function F1 is also given below:

[0131]

[0132] The specific form of 1 in the above formula is as follows:

[0133]

[0134] in,

[0135] The eddy viscosity coefficient is defined as:

[0136]

[0137] In the formula, F2 is a mixing function, and its specific form is as follows:

[0138]

[0139] in,

[0140] In the formula, y is the distance to the wall, in meters; and σ is known. ω =1.168, σ k2 =1,σ ω1 =2, a1=0.31, β * =0.09.

[0141] When setting up the physical model, the dynamic viscosity of air is set to 1.853×10-5 Pa·s, the specific heat of air is set to 1005 J / kg·K, and the thermal conductivity of air can be kept at the default value of 0.0260305 W / m·K.

[0142] In practice, permanent magnet synchronous motors (PMSMs) are typically used for driving. Therefore, it is necessary to establish numerical equations for PMSMs. Specifically, for the induced magnetic field generated by the stator input current of the PMSM and the air gap magnetic field synthesized by the magnetic field of the permanent magnets, electromagnetic field parameters are analyzed based on electromagnetic field equations and boundary condition equations. The main control equations of the PMSM are as follows:

[0143] (1) Maxwell's equations Maxwell's equations include Ampere's circuital law of electromagnetic torque, Faraday's law of electromagnetic induction, Gauss's law of magnetic flux, and Gauss's law of electric flux. They analyze electromagnetic phenomena from a macroscopic perspective and reveal the basic laws of electromagnetic torque and electromagnetic phenomena. They are the basis for the numerical analysis of permanent magnet synchronous motors. The specific expression of Maxwell's equations is as follows:

[0144] Maxwell's-Ampere law is a curl equation that describes how the line integral of the magnetic flux density vector along any closed loop equals the sum of the currents passing through the surface of that closed loop. This equation is unaffected by the distribution of the medium or the magnitude of the magnetic field. Its expression is:

[0145]

[0146] In the formula, H is the magnetic field strength (A / m); J is the current density (A / m). 2 D is the electric displacement, C / m 2 .

[0147] Faraday's law, the curl equation, connects the electric and magnetic fields of an organic system to realize electromagnetic torque, thus revealing the law of electromagnetic torque and electromagnetic induction. Its expression is:

[0148]

[0149] In the formula, E is the electric field strength, V / m; and B is the magnetic flux density, T.

[0150] Gauss's law of magnetic flux is a divergence equation that describes the fact that the directed integral of the magnetic flux density vector on a closed surface is always zero. This equation is unaffected by the distribution of the magnetic medium or the distribution of the magnetic flux density vector, and its expression is:

[0151]

[0152] Gauss's law of electric flux is a divergence equation that describes how the directed integral of the electric displacement vector on a closed surface is equal to the algebraic sum of free charges within that closed surface. This equation is unaffected by the distribution of the dielectric material or the distribution of the electric flux density vector. Its expression is:

[0153]

[0154] In the formula Charge density, C / m 2 .

[0155] The continuity equation is a divergence equation, describing the differential expression of the current conservation law, and its expression is:

[0156]

[0157] The medium composition equation describes the relationship between electromagnetic torque and the field quantities related to the electromagnetic properties of the medium through constitutive relations. It explains the characteristics of electromagnetic torque in three types of electromagnetic media: dielectric, magnetic, and conductive media. It also explains in detail the mechanism of molecular magnetization, polarization, and electron conduction in electromagnetic torque media under the action of electromagnetic fields. Its expression is as follows:

[0158]

[0159] In the formula, ε is the permittivity, F / m.

[0160] When the material used is an isotropic medium, the parameters ε, μ, and σ in the above formula are all scalars; when the material used is an anisotropic medium, the parameters ε, μ, and σ in the above formula are all tensors.

[0161] (2) Based on the fundamental equations of the electromagnetic field, the partial differential equations under the separation of magnetic and electric fields can be derived using the potential function equations. The expression is as follows:

[0162]

[0163] In the formula, A is the vector magnetomotive force, AT; Φ is the scalar electric potential, V.

[0164] The relationship between the vector magnetic potential A and the scalar potential Φ, after applying the Lorentz condition, is expressed as follows:

[0165]

[0166] When performing finite element calculations on electromagnetic fields using the potential function equation, the first step is to solve for the vector magnetopotential A or the scalar electric potential Φ. Then, the magnitude of the magnetic field strength or electric field strength can be obtained according to the definition. Under constant magnetic field and passive electrostatic field conditions, the potential function equation can be simplified to the Laplace equation; under active electrostatic field conditions, it can be simplified to the Poisson equation; and under time-harmonic field conditions, it can be simplified to the Helmholtz equation.

[0167] (3) Two-dimensional electromagnetic field boundary condition equations: The electromagnetic field of a permanent magnet synchronous motor is actually a three-dimensional field, but the axial structure remains basically unchanged during the analysis. Therefore, the analysis is based on the two-dimensional electromagnetic field generated by the permanent magnet synchronous motor under the condition that the axial structure remains unchanged. In the analysis, the electromagnetic field problem of the ultra-high-speed permanent magnet motor is regarded as a quasi-steady magnetic field, and phenomena such as hysteresis and eddy current effects in the electromagnetic field are not considered. The equations satisfied by the two-dimensional electromagnetic field boundary condition problem are:

[0168]

[0169] In the formula, γ is the magnetic reluctance; J c H is the equivalent surface current of the permanent magnet, A; t The tangential component of the magnetic field strength is given in A / m and J. s Source current density, A / m 2 .

[0170] In reality, electromagnetic fields do not have boundary conditions. However, for the sake of simplicity, electromagnetic torque boundary conditions are introduced. Commonly used boundary conditions include the first type of Dirichlet boundary condition and the second type of homogeneous boundary condition. In this paper, the first type of Dirichlet boundary condition for electromagnetic torque is used, which is expressed as:

[0171] A| Γ1 =A0 (24)

[0172] When the tangential component of the magnetic field strength is zero, the boundary value problem of a two-dimensional electromagnetic field solved using the variational method satisfies the following equation:

[0173]

[0174] in,

[0175] Using discretization methods, the two-dimensional electromagnetic field is divided into a finite number of cells to solve the conditional variational problem. Triangular elements are typically used as the finite number of cells, such as... Figure 5 As shown.

[0176] Numbering the nodes of the triangular finite element cell counterclockwise as a, b, c, the shape function of the triangular cell can be expressed as:

[0177]

[0178] Let n = a, b, c, then in equation (26) Represented as:

[0179]

[0180] in,

[0181] For simplified analysis, the above equations need to be transformed into a system of linear difference algebraic equations. The magnetic potential at any point in the electromagnetic field can be expressed as:

[0182]

[0183] In the formula A z A represents the magnetic potential.

[0184] Substituting the original function of the interpolation function and the first-order partial derivatives of the interpolation function with respect to variables x and y into equation (24), and differentiating the energy functional within the triangular finite element cell, we can obtain:

[0185]

[0186] in,

[0187] By summing the derivatives of the element energy functions over the entire computational domain with respect to the magnetic potential at the same node, the linear equation system obtained when the overall system energy value is minimized is as follows:

[0188] [k1]{A}={p1} (30)

[0189] Finally, the magnetic field strength at each point in the electromagnetic field is obtained by solving equation (30).

[0190] For the load excitation model, due to the effect of the internal flow field of the ultra-high-speed electric air compressor, there is a difference between the actual load torque and the ideal load torque. This difference is the load excitation of the ultra-high-speed electric air compressor. In actual operation, the presence of load excitation can cause vibration and noise in the ultra-high-speed electric air compressor, and in severe cases, even impeller damage, resulting in significant losses. To study and analyze the influence of key parameters of the ultra-high-speed electric air compressor on the load excitation, a mathematical model of the load excitation is presented. To simplify the analysis, the following assumptions are made within an acceptable range for the load excitation equation: ① The fluid inside the ultra-high-speed electric air compressor is continuous; ② The impeller and flow channel walls are rigid walls, and their physical shapes do not change with the force applied. Therefore, the load excitation equation of the ultra-high-speed electric air compressor is:

[0191]

[0192] In the formula ΔT L For the load excitation of the ultra-high speed electric air compressor, N·m; The actual load torque of the ultra-high speed electric air compressor is expressed in N·m; T L The ideal load torque for an ultra-high-speed electric air compressor is N·m.

[0193] The energy output by the actual load torque of an ultra-high-speed electric air compressor is equal to the sum of the energy gained by the gas flowing through the compressor per unit time and the energy lost per unit time. The energy gained by the gas flowing through the compressor includes the energy gained from gas compression, the energy gained from work done by gravity, and the energy gained from the increase in gas velocity. This equation can be expressed as:

[0194]

[0195] In the formula, n is the impeller speed of the ultra-high-speed electric air compressor, in rpm; P p W; P is the energy gained by gas compression per unit time. g W; P is the energy gained by the gravitational potential energy of a gas per unit time. e P is the energy gained by the increase in gas velocity per unit time. hyd W represents the energy lost per unit time as gas flows through the ultra-high-speed electric air compressor.

[0196] Substituting the energy gained by the gas per unit time into equation (32), we get:

[0197]

[0198] In the formula, p1 and p2 are the inlet and outlet gas pressures, Pa; S1 and S2 are the inlet and outlet cross-sectional areas, m2; ν1 and ν2 are the inlet and outlet gas velocities, m / s; and g is the gravitational acceleration, m / s². 2 Q is the mass flow rate, kg / s; Z1 and Z2 are the gas inlet and outlet heights, m.

[0199] The energy loss per unit time of gas flowing through an ultra-high-speed electric air compressor can be further divided into impeller inlet impact loss, gas leakage loss, impeller resistance loss, and friction loss. Impeller inlet impact loss is the energy loss caused by the collision between the airflow and the impeller at the impeller inlet. The magnitude of this loss is closely related to the airflow direction, airflow velocity, and the impeller rotation speed of the ultra-high-speed electric air compressor. The velocity triangle at the impeller inlet is shown in the figure below. Figure 6 As shown.

[0200] Figure 6 In this context, ν1 represents the absolute velocity of the incoming airflow, u1 represents the tangential velocity of the impeller of the ultra-high-speed electric air compressor, w1 represents the relative velocity of the gas, and w c This represents the velocity component of the impeller inlet impact loss in an ultra-high-speed electric air compressor. From the geometric relationships between the velocity triangles in the first-stage impeller, the expression for the velocity component of the impeller inlet impact loss in an ultra-high-speed electric air compressor is:

[0201]

[0202] In the formula, α1 is the inflow angle of the first-stage impeller, rad; β1 is the installation angle of the first-stage impeller, rad; and b is the number of impeller blades.

[0203] The absolute velocity of the airflow can be expressed as:

[0204]

[0205] In the formula, D1 is the nominal diameter of the first-stage impeller of the ultra-high-speed electric air compressor, in meters; h is the diameter of the impeller shaft, in meters.

[0206] The formula for the tangential velocity of the impeller of an ultra-high-speed electric air compressor is:

[0207]

[0208] Solving equations (3.3) to (3.5) simultaneously, we obtain the first-stage impeller inlet impact loss work H1:

[0209]

[0210] First-stage impeller inlet impact loss power P 11 It can be represented as:

[0211]

[0212] Similarly, the impact loss power P at the inlet of the second-stage impeller can be obtained. 12 :

[0213]

[0214] In the formula, α2 is the inflow angle of the second-stage impeller, rad; β2 is the installation angle of the second-stage impeller, rad; and D2 is the nominal diameter of the second-stage impeller of the ultra-high-speed electric air compressor, m.

[0215] Gas leakage, a form of energy loss in gas systems, is primarily caused by pressure differences, leading to gas flow from high-pressure areas to low-pressure areas and resulting in gas leakage. Figure 7 .

[0216] The leakage of gas is directly proportional to the flow rate of the ultra-high-speed electric air compressor, and the power loss can be expressed as:

[0217]

[0218] In the formula λ c The loss factor represents the ratio of leakage flow rate to total flow rate.

[0219] Wheel drag loss, a component of energy loss in gas, occurs when the impeller rotates. This friction between the outer surfaces of the impeller disk and cover and the gas in the surrounding gaps causes the impeller to perform additional work, including overcoming frictional resistance torque, in addition to the work it does on the gas. This extra work is called wheel drag loss. Figure 8 .

[0220] Wheel resistance loss power can be expressed as:

[0221]

[0222] In the formula β df This is the wheel resistance loss coefficient.

[0223] Frictional energy loss in gas flow occurs due to friction between the gas and the walls of the volute channel, as well as friction between gas layers with different flow velocities during the flow process. Figure 9 According to fluid mechanics, the flow within a channel can be divided into two parts: the mainstream region and the boundary layer near the wall. Within the boundary layer, the velocity gradient along its thickness is significant, and internal friction or viscosity exists between the fluids. To maintain fluid flow, the gas in the mainstream region must continuously inject energy into the gas within the boundary layer to overcome these internal friction or viscosity forces.

[0224] The friction loss in the flow channel of an ultra-high-speed electric air compressor can be calculated using the following formula:

[0225]

[0226] In the formula, λ is the frictional resistance coefficient, which is related to the flow Reynolds number and the relative roughness of the volute wall; l is the flow path length or the length of the average streamline, in meters; d h denoted as , where is the average equivalent diameter of the volute channel, in meters (m); and c is the average airflow velocity within the volute channel, in m / s.

[0227] Frictional power loss can be expressed as:

[0228]

[0229] Substituting the energy change expressions of each part of the ultra-high-speed electric air compressor into equation (33), and after simplification, we can obtain the expression for the actual load torque of the ultra-high-speed electric air compressor:

[0230]

[0231] The expression for coefficient a1 is:

[0232]

[0233] The expression for coefficient a2 is:

[0234]

[0235] The expression for coefficient a3 is:

[0236]

[0237] The expression for coefficient a4 is:

[0238]

[0239] The expression for coefficient a5 is:

[0240]

[0241] The ideal load torque of an ultra-high-speed electric air compressor is affected by the power and speed of the ultra-high-speed electric air compressor. This can be expressed by the following equation:

[0242]

[0243] In the formula, P represents the power of the ultra-high-speed electric air compressor, in kW.

[0244] Finally, the load excitation formula for the ultra-high speed electric air compressor is obtained:

[0245]

[0246] Based on the foregoing, this embodiment constructs the numerical equations and load excitation model of the ultra-high-speed electric air compressor. In order to realize the design parameters for the first-stage guide vane angle, the second-stage guide vane angle, the first-stage impeller mounting angle, and the second-stage impeller mounting angle, it is necessary to calculate them according to the following method.

[0247] In the analysis of the influence of changes in rotational speed and load torque on the load excitation amplitude, it can be found that the collision between the airflow and the impeller leading edge is the main cause of load excitation. Therefore, the load excitation amplitude is reduced by adjusting the airflow injection angle by changing the guide vane tilt angle. Five new guide vane angles were designed, with tilt angles of 70°, 80°, 90°, 100°, and 110°. First, the original first-stage guide vanes were replaced with the new-angle guide vanes, and then the new model was put into simulation software for numerical calculation. The boundary conditions were set to simulate the actual operation of a high-speed electric air compressor under common operating conditions, with a rotational speed n = 80000 rpm; inlet pressure of 101.325 kPa; inlet temperature of 290.15 K; and outlet mass flow rate and temperature were set to the parameters from the experimental results under common operating conditions. The time step was selected as the time required for the impeller to rotate 1°, i.e., 2.083 × 10⁻⁶. -6 The calculation method is the same as before: first, a steady-state calculation is performed, and after the calculation converges, the physical model is adjusted to perform a transient calculation with three impeller rotations to obtain the flow field and load excitation data obtained from the numerical calculation. This embodiment is based on the above five tilt angles for calculation. In the implementation process, other angles can also be selected for calculation according to the actual situation.

[0248] The streamline diagrams of the first-stage guide vanes and the pressure contour diagrams of the first-stage impeller surface were obtained after numerical calculations of guide vanes with different inclination angles, as follows: Figure 10Before the impeller leading edge, the airflow is not affected by the impeller's rotation, resulting in less disturbance and a more uniform and fixed direction, as indicated by the red circle at point A in the streamline diagram. However, between the blades, the blades exert a strong rotational effect on the airflow, causing a significant change in its direction, as indicated by the red circle at point B in the streamline diagram. The streamline diagram shows that after adding guide vanes, the airflow direction after the guide vanes is affected by the vane angle, resulting in a significant change in airflow direction and a corresponding change in the collision angle between the airflow and the impeller leading edge. This demonstrates that adding guide vanes with different angles can effectively streamline the inlet airflow and alter the collision angle between the incident airflow and the impeller leading edge. In the area indicated by the red circle in the surface impeller pressure cloud diagram, the leading edge pressure does not increase with the guide vane angle as expected; instead, the leading edge surface pressure is lowest at a 90° angled guide vane.

[0249] To verify whether the change in load excitation amplitude is related to the pressure cloud on the impeller surface. Figure 1 The numerical calculation results are exported as a time-domain plot, and a fast Fourier transform is performed. Then, the dominant frequency amplitude is exported. Figure 11 .pass Figure 11 The load excitation frequency for guide vanes with different inclination angles was 17333 Hz. The amplitude of the load excitation was 0.07215 N·m at an inclination angle of 70°, 0.06619 N·m at 80°, 0.06017 N·m at 90°, 0.06679 N·m at 100°, and 0.07280 N·m at 110°. Comparison revealed that after modifying the guide vane inclination angle, the amplitude of the load excitation first decreased and then increased with increasing inclination angle, reaching its minimum at 90°, consistent with the results of impeller surface pressure feedback. However, this contradicts the expected assumption that the load excitation amplitude would continuously increase with increasing guide vane inclination angle. This indicates that factors other than the airflow injection angle affect the load excitation results in this numerical calculation. Analysis reveals that the main reason is the interference between the guide vanes and the impeller's moving and stationary blades, leading to an increased amplitude of the load excitation. This interference occurs due to uneven pressure distribution along the circumference of the fluid, generating localized high pressure at the tail of the guide vanes and the front of the impeller during impeller rotation. The interference intensity is minimal when the guide vane inclination angle is 90°. However, changes in the guide vane inclination angle still have some impact on the load excitation amplitude, albeit a smaller one compared to the impact of the moving and stationary blade interference. The fact that the rate of change of amplitude is not the same on both sides of the 90° angle supports this view; the increase in amplitude is less on the side with the smaller angle than on the side with the larger angle.

[0250] The second-stage guide vane angle is calculated as follows. In the variation of the first-stage guide vane angle, the amplitude first increases and then decreases with increasing angle, reaching a minimum near 90°. The second-stage guide vane angle is then adjusted to verify whether the variation pattern is consistent with that of the first-stage guide vane angle, and whether there are any differences. The original second-stage guide vanes are replaced with second-stage guide vanes with angles of 70°, 80°, 90°, 100°, and 110°, and the new model is then placed in simulation software for numerical calculation. The boundary conditions are set consistently as in the numerical calculation of the first-stage guide vane angle variation and are not changed.

[0251] The streamline diagrams of the second-stage guide vanes and the pressure contour diagrams of the second-stage impeller surface were obtained after numerical calculations of guide vanes with different inclination angles, as follows: Figure 12 Streamlines at different angles from the first-stage guide vane. Figure 1 Before the impeller leading edge, the airflow is not affected by the impeller's rotation, resulting in less disturbance and a more uniform and fixed direction, as indicated by the red circle at point A in the streamline diagram. However, between the blades, the blades exert a strong rotational effect on the airflow, causing a significant change in its direction, as indicated by the red circle at point B in the streamline diagram. The streamlines after the guide vanes are affected by the guide vane angle, resulting in a significant change in airflow direction and consequently, a change in the collision angle between the airflow and the impeller leading edge. Notably, the streamlines after the second-stage guide vanes are lighter in color than those after the first-stage guide vanes, indicating a faster airflow velocity. As shown by the red circle on the pressure cloud diagram of the second-stage impeller surface, the color contrast between the leading edge and surrounding areas is higher than that of the first-stage impeller, leading to more drastic changes in the amplitude of the load excitation.

[0252] To verify whether the change in load excitation amplitude is related to the pressure cloud on the impeller surface. Figure 1 The numerical calculation results are exported as a time-domain plot, and a fast Fourier transform is performed. Then, the dominant frequency amplitude is exported. Figure 13 .pass Figure 13The dominant frequency of the load excitation for the second-stage guide vanes at different inclination angles was 17333 Hz. The amplitude of the load excitation was 0.07413 N·m at an inclination angle of 70°, 0.06739 N·m at 80°, 0.06017 N·m at 90°, 0.06859 N·m at 100°, and 0.07477 N·m at 110°. Comparison revealed that the amplitude of the load excitation after modifying the second-stage guide vane inclination angle first decreased and then increased with increasing inclination angle, reaching its minimum at 90°, consistent with the trend of the first-stage guide vane inclination angle. However, the change in the load excitation amplitude was greater after changing the second-stage guide vane inclination angle because the airflow velocity and pressure after the second-stage guide vane were higher than those after the first-stage guide vane. As can be seen from the above steps, the optimal tilt angle for both the first-stage and second-stage guide vanes is 90°.

[0253] To obtain the optimal installation angles for the first and second stage impellers, the following methods are used: Besides adjusting the airflow angle by installing inlet guide vanes to streamline the airflow at the impeller inlet, the airflow collision angle can also be adjusted by changing the impeller installation angle to alter the collision angle between the impeller leading edge and the airflow, thereby reducing the load excitation amplitude. The installation angles of the first and second stage impellers in ultra-high-speed air compressors are not the same. Therefore, the installation angle of the first stage impeller is changed first for comparative analysis, followed by the installation angle of the second stage impeller. This allows for control of a single variable while comparing the impact of changes in the angles of the first and second stage impellers on the load excitation amplitude. Four new first-stage impeller angles (34°, 37°, 43°, and 46°) are selected for comparative analysis with the original impeller. The impellers with the new angles are added to the numerical calculation model to calculate the impeller leading edge streamlines and impeller surface pressure under the four blade installation angles. These results are then plotted together with the numerical calculation results of the original impeller. Figure 14 .

[0254] exist Figure 14 The streamlines in the diagram show the change in the impeller mounting angle. As the mounting angle increases, the velocity loss due to airflow collisions decreases, as indicated by the red circle in the diagram. The streamlines in front of the impeller have a relatively uniform direction because they are not significantly disturbed. The streamlines between the impellers change direction due to the impeller rotation. In the pressure cloud diagram of the first-stage impeller, marked by the red circle, the smaller the impeller mounting angle, the more pronounced the pressure at the impeller leading edge compared to the surrounding areas, indicating higher pressure.

[0255] The numerical calculation results are output as a load excitation time-domain plot. Then, the load excitation time-domain plot is transformed using a Fast Fourier Transform to obtain the load excitation frequency-domain plot, as follows: Figure 15At a frequency of 17333 Hz, the amplitude of the load excitation is 0.06281 N·m for an installation angle of 34°, 0.06178 N·m for 37°, 0.06017 N·m for 40°, 0.05993 N·m for 43°, and 0.05857 N·m for 46°. The load excitation amplitude decreases as the installation angle of the first-stage impeller increases, with the smallest amplitude observed at a 46° installation angle. Therefore, a design value of 46° for the first-stage impeller installation angle is optimal.

[0256] Based on the original two-stage impeller, four sets of new impeller angles (48°, 51°, 57°, and 60°) were selected for comparative analysis with the original impeller. The impellers with the new angles were incorporated into a numerical calculation model to obtain the impeller leading-edge streamlines at the four blade installation angles. These streamlines were then plotted together with the numerical calculation results of the original impeller. Figure 16 .

[0257] exist Figure 16 The diagram shows the change in impeller mounting angle. As the mounting angle increases, the velocity loss component due to airflow collision decreases, as indicated by the red circle in the figure. Furthermore, it reveals that the airflow velocity at the second-stage impeller inlet is higher than that at the first-stage impeller inlet. In the pressure cloud diagram of the second-stage impeller, the increased airflow velocity and pressure result in higher pressure not only at the impeller leading edge but also on the rotor shaft compared to the first-stage impeller. Moreover, the presence of the volute bend at the second-stage impeller inlet further worsens the airflow uniformity. Therefore, the amplitude of the load excitation not only decreases with increasing second-stage impeller mounting angle but also appears to have a greater impact on the load excitation amplitude than that of the first-stage impeller.

[0258] The numerical calculation results are output as a load excitation time-domain plot. Comparison shows that the load excitation decreases with increasing installation angle. The load excitation time-domain plot is then converted to a load excitation frequency-domain plot for detailed analysis. Figure 17 At a frequency of 17333 Hz, the amplitude of the load excitation was found to be 0.06579 N·m at an installation angle of 48°, 0.06203 N·m at 51°, 0.06017 N·m at 54°, 0.05799 N·m at 57°, and 0.05507 N·m at 60°. The load excitation amplitude was minimized when the installation angle of the second-stage impeller was 60°. Furthermore, when the installation angles of the first and second-stage impellers changed by the same angle, the change in the installation angle of the second-stage impeller had a greater impact on the load excitation amplitude. Therefore, a 60° installation angle for the second-stage impeller is optimal.

[0259] The above description, based on the embodiments shown in the figures, details the structure, features, and effects of the present invention. The above description is only a preferred embodiment of the present invention, but the present invention is not limited to the scope of implementation shown in the figures. Any changes made in accordance with the concept of the present invention, or equivalent embodiments modified to have equivalent changes, that do not exceed the spirit covered by the specification and figures, should be within the protection scope of the present invention.

Claims

1. A method for setting parameters of a fuel cell air compressor based on load excitation amplitude analysis, used for an ultra-high-speed electric air compressor in a fuel cell vehicle, wherein the ultra-high-speed electric air compressor is a two-stage centrifugal electric air compressor, and the two-stage centrifugal electric air compressor is provided with a first-stage impeller and a second-stage impeller. characterized in that Includes the following steps, S1. Construct the numerical equations and load excitation model of the ultra-high speed electric air compressor; S2. Using the load excitation model in step S1, simulations are performed under different first-stage guide vane angles to obtain the first simulation results of the load excitation amplitude. S3. Using the load excitation model in step S1, simulations are performed under different secondary guide vane angles to obtain the second simulation results of the load excitation amplitude. S4. Based on the first simulation result in step S2, select the first-stage guide vane angle corresponding to the minimum load excitation amplitude in the first simulation result as the first parameter. S5. Based on the second simulation results in step S3, select the second guide vane angle corresponding to the minimum load excitation amplitude in the second simulation results as the second parameter. Step S1 includes the following steps: S11, constructing the numerical equations for an ultra-high-speed electric air compressor; S12, construct the numerical equations for the permanent magnet synchronous motor; S13, Construct a load excitation model for an ultra-high-speed electric air compressor; The formula corresponding to the load incentive model is: wherein ΔT L is the load excitation of the super-high-speed electric air compressor; Q is the mass flow, n is the impeller speed of the super-high-speed electric air compressor, and P is the power of the super-high-speed electric air compressor. Wherein, a1 is the first-stage impeller inlet flow angle; b1 is the first-stage impeller installation angle; b is the number of impeller blades; D1 is the nominal diameter of the first-stage impeller of the super-speed electric air compressor; a2 is the second-stage impeller inlet flow angle; b2 is the second-stage impeller installation angle; D2 is the nominal diameter of the second-stage impeller of the super-speed electric air compressor; t represents time; lc is the loss coefficient, representing the ratio of leakage flow rate to total flow rate; v1 is the outlet gas velocity; v2 is the inlet gas velocity; g is the gravitational acceleration; Z1 is the gas inlet height; Z2 is the gas outlet height; l is the friction resistance coefficient, which is related to the flow Reynolds number and the relative roughness of the volute wall surface; l is the length of the flow path or the length of the average flow line; c is the average gas flow velocity in the volute flow passage; d h is the average equivalent diameter of the volute flow passage; S1 is the outlet cross-sectional area; S2 is the inlet cross-sectional area; p1 is the inlet gas pressure; and p2 is the outlet gas pressure. The application also discloses a super-speed electric air compressor.

2. The method of claim 1, wherein: It also includes the following steps, S6. Using the load excitation model in step S1, simulations are performed under different first-stage impeller installation angles to obtain the third simulation results of the load excitation amplitude. S7. Using the load excitation model in step S1, simulations are performed under different second-stage impeller installation angles to obtain the fourth simulation result of the load excitation amplitude. S8. Based on the third simulation result in step S6, select the first-stage impeller installation angle corresponding to the minimum load excitation amplitude in the third simulation result as the third parameter. S9. Based on the fourth simulation result in step S7, select the second-stage impeller installation angle corresponding to the minimum load excitation amplitude in the fourth simulation result as the fourth parameter. S10 uses the first, second, third, and fourth parameters as the design parameters for the ultra-high-speed electric air compressor.

3. The method of claim 1, wherein: The assumptions in the construction of the load excitation model are that the fluid inside the ultra-high-speed electric air compressor is continuous, and the impeller and flow channel walls are rigid walls whose physical shapes do not change with the force applied.

4. The method of claim 1, wherein: The load excitation model is a model that considers impeller inlet impact loss, gas leakage loss, wheel resistance loss and friction loss.

5. The method of claim 1, wherein: Step S2 includes the following steps: S21, set the first-stage guide vane angle to different angles, and simulate the actual operation of the over-adjustment electric air compressor according to the set speed, inlet pressure and inlet humidity; S22, the time step is selected as the time required for the impeller to rotate 1°; S23, first perform steady-state calculations until the calculations converge; S24, perform transient calculations for three impeller rotations to obtain the numerically calculated flow field and first load excitation data.

6. The method of claim 5, wherein: Step S4 includes the following steps: S41, obtain the streamline diagram of the first-stage guide vane and the pressure cloud diagram of the first-stage impeller surface after numerical calculation of guide vanes with different inclination angles; S42, selecting the best first-stage guide vane angle as the first parameter through the first-stage impeller surface pressure cloud map.

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