A method for engine structure design based on moore loudness
By using Moore loudness assessment and a multi-body dynamics model, the vibration and sound pressure level of the engine casing were calculated. The casing structure was modified to avoid the excitation source frequency, which solved the problem that the traditional evaluation system could not reflect human subjective feelings. This effectively reduced the engine's radiated noise and improved its sound quality.
Patent Information
- Authority / Receiving Office
- CN · China
- Patent Type
- Patents(China)
- Current Assignee / Owner
- SOUTH CHINA UNIV OF TECH
- Filing Date
- 2023-03-01
- Publication Date
- 2026-07-31
AI Technical Summary
When evaluating engine NVH performance, the traditional A-weighted sound pressure level evaluation system cannot accurately reflect human subjective perception of noise, and it is difficult to effectively reduce engine radiated noise from a structural perspective, especially when the boundary conditions of the whole assembly model are inaccurate and the excitation data is uncertain.
The Moore loudness evaluation index, combined with a multi-flexible body dynamics model and fundamental wave technology, is used to calculate the vibration acceleration and sound pressure level of the engine casing surface. By modifying the natural frequency and local stiffness of the casing structure and avoiding the excitation source frequency, the engine structure is designed to reduce radiated noise.
It improves the accuracy of engine structural radiated noise assessment and ride comfort, reduces the overall characteristic loudness of the engine by more than 10%, meets lightweight requirements, and improves sound quality characteristics.
Smart Images
Figure CN116451519B_ABST
Abstract
Description
Technical Field
[0001] This invention relates to the field of analysis and control of vibration and noise in automobile engines, and more particularly to an engine structure design method based on Moore loudness. Background Technology
[0002] Engine NVH (noise, vibration, harshness) performance is a key factor for consumers when purchasing a car, and it has a significant impact on passenger comfort. Engine structural vibration mainly refers to the vibration acceleration of the engine housing surface under various excitations. Engine structural radiated noise refers to the noise radiated from the structural surfaces of engine components such as the cylinder block, cylinder head, and oil pan under engine excitation.
[0003] The evaluation of engine NVH performance is generally conducted from two perspectives: structural vibration and radiated noise. Engine structural vibration refers to surface vibration velocity and acceleration response. Radiated noise is generally evaluated using A-weighted sound pressure level and sound power level. To consider human subjective perception of noise, objective sound quality models can also be used to evaluate engine noise. Commonly used sound quality parameters include loudness and sharpness. Sound quality evaluation models can establish the relationship between objective sound quality parameters and subjective sound quality evaluation. Therefore, with the help of evaluation models, subjective sound quality evaluation can be simulated without human intervention. Traditional A-weighted engine radiated noise evaluation systems, on the one hand, do not adequately describe the relative magnitude of sound intensity on the frequency scale and cannot accurately reflect the differences in sound intensity of various frequency components; on the other hand, they do not consider human psychological characteristics and cannot accurately assess human subjective perception of engine noise.
[0004] There are two common methods for vibration and noise reduction: one is to reduce the excitation sources, addressing the noise problem at its source; the other is to modify the structural form, addressing the noise problem along the transmission path of vibration and noise. Due to the complexity of the engine system structure and the presence of multiple excitation sources, it is difficult to solve the problem from the perspective of the excitation sources. The second method is simple to operate; it involves modifying the engine structure to increase the local stiffness, altering the modal parameters of the structure, and avoiding the frequencies of the excitation sources. Ultimately, this achieves the goal of reducing the engine's radiated noise level and improving its acoustic characteristics.
[0005] Chinese invention patent CN106250602A provides a "structural optimization method for NVH of engine oil pan", which can optimize the structure of NVH of engine oil pan. However, this solution has the following defects: (1) It only optimizes the oil pan separately and does not assemble the oil pan and the engine into a whole. The boundary conditions of the isolated oil pan model may deviate from the actual assembly, resulting in inaccurate calculation results; (2) The oil pan excitation obtained by vibration calculation has not been verified by experiments, and the accuracy of the excitation data cannot be guaranteed; (3) The optimization target is vibration velocity, which is unrelated to sound and cannot directly reflect the noise level generated by the oil pan. Using sound-related parameters to evaluate the noise of the oil pan can more directly and accurately reflect the optimization effect of the oil pan. Summary of the Invention
[0006] To address the subjective perception of noise, this invention provides an engine structure design method based on Moore's characteristic sound quality. Using an established finite element model of the engine, the modal parameters of the engine casing are calculated. Based on a multi-flexible body dynamics model and fundamental frequency techniques, the surface vibration acceleration of the engine casing and the sound pressure level at the engine's free-field point are calculated. Using the Moore's loudness calculation method, the engine's ERB level, the output excitation of the auditory filter, and finally, the engine's Moore characteristic loudness are calculated. Combining the casing modal parameters, surface vibration acceleration, and Moore characteristic loudness, the engine casing structure is redesigned to alter its natural frequency, reduce structural radiated noise, and improve the engine's sound quality.
[0007] To achieve the objective of this invention, the present invention provides an engine structure design method based on Moore sound quality, comprising the following steps:
[0008] (1) Calculate the modal parameters of the engine: Establish finite element models of the engine housing assembly, crankshaft assembly and balance shaft, and calculate the natural frequency and mode shape of the housing assembly within 1600Hz.
[0009] (2) Calculate the vibration response of the housing surface: Reduce the finite element model of the engine housing assembly, crankshaft assembly and balance shaft to establish the multi-flexible body dynamic model of the engine. Use the iterative method to calculate the solution of the reduced model, that is, the vibration response of the engine housing surface nodes after reduction. Obtain the vibration velocity of the engine housing surface through the inverse reduction transformation.
[0010] (3) Calculation of the sound pressure level (SPL) at the front end, intake and exhaust sides, and top of the engine, and the average sound pressure level at four points: First, a free-field model of the engine was established using the fundamental wave technique. The surface vibration velocity of the casing, calculated using multi-body dynamics, was used as the boundary condition to calculate the sound pressure levels at the front end, intake side, exhaust side, and top of the engine. Second, the average sound pressure level at these four points was calculated using the following formula:
[0011]
[0012] In the formula, P i The results show the sound pressure level at four different locations.
[0013] (4) The outer ear and middle ear amplify and attenuate sound, respectively. First, the sound pressure level of the sound at the field point is corrected using the outer ear transfer function and the middle ear transfer function. Second, after the sound reaches the cochlea of the inner ear, the energy of all engine noise signals within each equivalent rectangular bandwidth (ERB) is calculated, i.e., the ERB level, denoted as L. ERB The calculation formula is:
[0014]
[0015] In the formula, f i E is the center frequency (effective frequency). n P0 is the equivalent rectangular bandwidth corresponding to the center frequency, and P0 is the reference sound pressure level (2×10). -5 Pa); P j W is the corrected sound pressure level at the j-th effective frequency; a (j) is the weighting function for the j-th effective frequency, representing the contribution of each effective frequency within the frequency band to the ERB level; the calculation formulas for each parameter are as follows:
[0016] W a (j)=(1+p j g j )exp(-p j g j (3)
[0017] g j =|f j -f i | / f i (4)
[0018] p j =4f j / E n (5)
[0019] E n=24.673(0.004368f i +1) (6)
[0020] (5) Calculate the output excitation of the auditory filter. The calculation formula is as follows:
[0021]
[0022] In the formula, E0 is the reference excitation, representing the output excitation value of a pure tone with a frequency of 1kHz and a sound pressure level of 0dB in a free field at a center frequency of 1kHz via an auditory filter. W b (j) is the shape function of the filter, and the calculation formulas for each parameter are as follows:
[0023] W b (j)=(1+p ij g ij )exp(-p ij g ij (8)
[0024]
[0025]
[0026] (6) Calculate the Moore loudness of the engine within the frequency range of interest: where the Moore characteristic loudness is:
[0027]
[0028] In the formula, N i ' represents the characteristic loudness of the i-th auditory filter, E THRQ G represents the energy level of the monoauricular audible valve, G represents the low-level gain of the cochlear amplifier at a specific frequency, C = 0.046871, A is a frequency-dependent constant, and α is the exponential relationship between excitation and loudness, which is frequency-dependent. The values of each parameter can be obtained according to the ANSI loudness calculation standard.
[0029] The total loudness is obtained by integrating the characteristic loudness in the frequency domain.
[0030] (7) Based on the calculated Moore characteristic loudness, observe the frequency band range corresponding to the peak value of the characteristic loudness curve, and combine the shell modal parameters from step (1) and the shell surface vibration acceleration from step (2). Determine that the peak value of the curve is generated by shell resonance.
[0031] (8) Engine Structure Redesign: By modifying the engine's structural form, the local stiffness of the structure is improved, the modal parameters of the structure are changed, and the frequency of the excitation source is avoided. The ultimate goal is to reduce the engine's radiated noise level and improve the engine's acoustic characteristics. The improved structure must meet two requirements: i) When designing the structure, it must be ensured that the stiffeners in the design area do not interfere with the internal components of the engine; ii) In order to achieve the goals of lightweighting and economy, the overall weight must be controlled, and the increase in the overall weight of the improved engine shall not exceed 0.5 kg.
[0032] (9) Vibration and noise quality calculation: Recalculate according to steps (2) to (6) to obtain the Moore characteristic loudness of the engine. If the overall characteristic loudness of the engine meets the predetermined target of a reduction of more than 10% at the peak, then the scheme is determined as the final design scheme.
[0033] Compared with the prior art, the present invention has at least the following advantages:
[0034] 1) In the multi-body dynamics model, the inertia and torsional stiffness of the flywheel-end drive shaft are considered. The drive shaft is equivalent to three parts: an elastic coupling, an intermediate shaft, and a motor shaft. A parametric modeling method is used to model the drive shaft. A modeling method for the scissor teeth structure is proposed, using a nonlinear spring-damped model to represent the connection relationship between the main and auxiliary scissor teeth. This modeling method effectively improves the model's accuracy and provides modeling reference support for engine systems with balance shafts.
[0035] 2) The field point sound pressure level of the free sound field is calculated based on the fundamental wave technology theory, which can improve research efficiency and reduce costs compared with the test method.
[0036] 3) The Moore loudness evaluation index fully considers the sensitivity of the human ear to different frequencies of sound, converts the sound pressure signal into characteristic loudness to evaluate the sound quality of noise, and can reflect the relative magnitude of sound energy of different frequency components. Compared with the traditional A-weighted sound pressure level evaluation index, the Moore loudness evaluation index is consistent with the subjective evaluation results and is more accurate in evaluating engine radiated noise.
[0037] 4) An engine structure design method based on Moore loudness changes the natural frequency of the engine casing by altering its structural form, thus preventing resonance. After the engine structure is improved, the overall weight increase does not exceed a preset value, and the overall characteristic loudness of the engine is reduced by more than 10%. This method is simple, meets the requirements for engine lightweighting, and effectively reduces the structural radiated noise level of the engine.
[0038] 5) The present invention employs a structural design method to redesign the oil pan and timing cover. Attached Figure Description
[0039] Figure 1 This is a flowchart illustrating an engine structure design method based on Moore loudness provided in an embodiment of the present invention.
[0040] Figure 2 This is an equivalent diagram of the connection between the crankshaft and the transmission shaft in the multi-flexible body dynamics model of this invention.
[0041] Figure 3(a) is a diagram of the scissor tooth structure in the multi-flexible body dynamics model of the present invention.
[0042] Figure 3(b) is the equivalent diagram of the scissor teeth in the multi-flexible body dynamics model of the present invention.
[0043] Figure 4(a) shows the results of the normal vibration acceleration of the nodes on the engine oil pan.
[0044] Figure 4(b) shows the results of the normal vibration acceleration of the engine timing cover node.
[0045] Figure 5 This is a graph showing the calculated average sound pressure level of the engine at four points.
[0046] Figure 6 This is the calculated Moore characteristic loudness curve of the engine.
[0047] Figure 7(a) is a schematic diagram of the original engine timing cover of the present invention.
[0048] Figure 7(b) is a structural design diagram of the engine timing cover of the present invention.
[0049] Figure 7(c) is a schematic diagram of the original design of the engine oil pan according to the present invention.
[0050] Figure 7(d) is a structural design diagram of the engine oil pan of the present invention.
[0051] Figure 8(a) is a comparison of the Moore characteristic loudness of the engine front end before and after the structural design.
[0052] Figure 8(b) is a comparison of the Moore characteristic loudness on the engine intake side before and after the structural design.
[0053] Figure 8(c) is a comparison of the Moore characteristic loudness on the exhaust side of the engine before and after the structural design.
[0054] Figure 8(d) is a comparison of the average Moore characteristic loudness at four points of the engine before and after the structural design. Detailed Implementation
[0055] To make the objectives, technical solutions, and advantages of this invention clearer and more explicit, the invention will be further described in detail below with reference to the accompanying drawings and embodiments.
[0056] like Figure 1 As shown, the present invention provides an engine structure design method based on Moore loudness, comprising the following steps:
[0057] (1) Calculate the modal parameters of the engine: Establish finite element models of the engine housing assembly, crankshaft assembly and balance shaft, and calculate the natural frequency and mode shape of the housing assembly within 1600Hz (this frequency is determined by the frequency range of interest. In this embodiment, the frequency of interest is within 1600Hz. It can be understood that in other embodiments, it can be changed to other values), as shown in Table 1.
[0058] Table 1 Modal results of engine casing
[0059]
[0060]
[0061] In some embodiments of the present invention, the finite element models of the housing assembly and the crankshaft assembly are both assemblies of multiple parts.
[0062] (2) Calculation of vibration response on the housing surface: The finite element models of the engine housing assembly, crankshaft assembly, and balance shaft are reduced and transformed to establish a multi-body dynamics model of the engine. In the multi-body dynamics model, the drive shaft is equivalent to three parts: an elastic coupling, an intermediate shaft, and a motor shaft. Parametric modeling is used to model the drive shaft, such as... Figure 2 As shown in Figure 3, a modeling method for scissor teeth is proposed, which uses a nonlinear spring-damped model to represent the connection relationship between the main and auxiliary scissor teeth. The solution of the reduced model is calculated using an iterative method (the iterative method is an existing solution method and will not be elaborated here), which is the vibration response of the engine housing surface nodes after reduction. The vibration velocity of the engine housing surface is obtained through the inverse transformation of the reduction.
[0063] The nonlinear spring-damped model, as shown in Figure 3, is established based on the assembly relationship of the scissor teeth. The modeling process is completed in the software. The selection of stiffness and damping is based on the bearing connection pair in the analysis software.
[0064] In some embodiments of the present invention, the reduction method for performing the reduction transformation is modal synthesis. The reduction transformation process is an existing method and will not be described in detail here.
[0065] In some embodiments of the present invention, the results of the normal vibration acceleration of the engine oil pan and timing cover nodes are shown in Figure 4.
[0066] (3) Calculate the sound pressure level (SPL) at the front end, intake side, exhaust side, and top of the engine, and the average sound pressure level at these four points: First, establish a free-field model of the engine using the fundamental wave technique, and use the surface vibration velocity of the casing calculated by multi-flexible body dynamics as the boundary condition to calculate the sound pressure level at the front end, intake side, exhaust side, and top of the engine; Second, calculate the average sound pressure level P at these four points of the engine. a The calculation formula is:
[0067]
[0068] In the formula, P i ( i =1,2,3,4) are the sound pressure level results at the four field points respectively.
[0069] In some embodiments of the present invention, the calculated four-point average sound pressure level of the engine is as follows: Figure 5 As shown.
[0070] (4) The outer ear and middle ear respectively enhance and attenuate sound. First, the sound pressure level of the sound at the field point is corrected using the outer ear transfer function and the middle ear transfer function (the correction method adopts existing methods and will not be elaborated here). Second, after the sound reaches the cochlea of the inner ear, the energy of all engine noise signals within each equivalent rectangular bandwidth (ERB) is calculated, i.e., the ERB level, denoted as L. ERB The calculation formula is:
[0071]
[0072] In the formula, f i E is the center frequency (effective frequency). n P0 is the equivalent rectangular bandwidth corresponding to the center frequency, and P0 is the reference sound pressure level (2×10). -5 Pa); P j W is the corrected sound pressure level at the j-th effective frequency; a (j) is the weighting function for the j-th effective frequency, representing the contribution of each effective frequency within the frequency band to the ERB level; the calculation formulas for each parameter are as follows:
[0073] W a (j)=(1+p j g j )exp(-p j g j (3)
[0074] g j =|f j -f i | / fi (4)
[0075] p j =4f j / E n (5)
[0076] E n =24.673(0.004368f i +1) (6)
[0077] g j These are intermediate variables in the calculation process.
[0078] (5) Based on energy and L ERB Calculate the output excitation E of the auditory filter i The calculation formula is:
[0079]
[0080] In the formula, E0 is the reference excitation, representing the output excitation value of a pure tone with a frequency of 1kHz and a sound pressure level of 0dB in a free field at a center frequency of 1kHz via an auditory filter. W b (j) is the shape function of the filter, and the calculation formulas for each parameter are as follows:
[0081] W b (j)=(1+p ij g ij )exp(-p ij g ij (8)
[0082]
[0083]
[0084] (6) Calculate the Moore loudness of the engine within the frequency range of interest: where the Moore characteristic loudness (both Moore characteristic loudness and Moore total loudness are Moore loudness. Moore characteristic loudness refers to the distribution characteristics in the frequency domain, and the Moore total loudness is obtained by integrating (summing) the Moore characteristic loudness in the frequency domain) is:
[0085]
[0086] In the formula, N i ' represents the characteristic loudness of the i-th auditory filter, E THRQG is the energy level of the monoauricular audible valve, G is the low-level gain of the cochlear amplifier at a specific frequency, C is a coefficient (in some embodiments of the present invention, the value is 0.046871), A is a frequency-related constant, and α is the exponential relationship between excitation and loudness, which is frequency-related. The values of each parameter can be obtained according to the ANSI loudness calculation standard.
[0087] The total loudness is obtained by integrating the Moore characteristic loudness in the frequency domain.
[0088] Step (6) calculates the frequency distribution characteristics of the characteristic loudness, which can reflect the relationship between the characteristic loudness and frequency; the input of the loudness calculation model is the four-point average sound pressure level, which takes into account the radiation noise of the entire engine (the four-point average sound pressure level is the average of four locations: the front end of the engine, the exhaust side, the intake side, and the top, so the radiation noise of the entire engine is taken into account).
[0089] In some embodiments of the present invention, the calculated engine Moore characteristic loudness curve is as follows: Figure 6 As shown.
[0090] (7) Based on the calculated Moore characteristic loudness, observe the frequency band range corresponding to the peak value of the characteristic loudness curve, and combine the shell modal parameters in step (1) and the shell surface vibration acceleration in step (2) for analysis to determine that the peak value of the curve is generated by shell resonance.
[0091] In some embodiments of the present invention, according to Figure 6 As shown, the characteristic loudness of both the intake and exhaust sides has a significant peak at 13.9 ERB, indicating a serious noise problem on the intake and exhaust sides in the 740-850Hz range. The characteristic loudness of the front end peaks at 18.5 ERB, corresponding to the noise problem at the front end in the 1362-1543Hz frequency band. The four-point average characteristic loudness corresponds to the peak frequency bands of the characteristic loudness on the intake and exhaust sides and the front end, respectively, and can represent the overall noise loudness of the engine. Combining the noise test, modal, and vibration results, it can be seen that because the vibration excitation frequency is close to the local mode of the oil pan (829Hz) and the local modes of the timing cover (1445Hz and 1468Hz), the oil pan and timing cover resonate, resulting in a large vibration response on the surfaces of the oil pan and timing cover, which in turn leads to a large characteristic loudness of the engine noise at the corresponding frequencies. The peak value of the curve is determined to be generated by the housing resonance.
[0092] (8) Engine structure redesign: By modifying the engine's structural form, the local stiffness of the structure is improved, the modal parameters of the structure are changed, and the frequency of the excitation source is avoided. Ultimately, the goal is to reduce the radiated noise level of the engine and improve its acoustic characteristics. The improved structure must meet two requirements: (a) When designing the structure, it must be ensured that the stiffeners in the design area do not interfere with the internal components of the engine; (b) In order to achieve the goals of lightweighting and economy, the overall weight must be controlled, and the increase in the overall weight after improvement must not exceed a preset value (in some embodiments of the present invention, this preset value is 0.5 kg).
[0093] In some embodiments of the present invention, modifications to the engine structure include deleting or adding reinforcing ribs, etc. Within the frequency range of interest, the aim is to reduce Moore characteristic loudness. Figure 7 shows a comparison of the original and structural designs for the engine oil pan and timing cover.
[0094] (9) Vibration and noise quality calculation: Recalculate according to steps (2) to (6) to obtain the Moore characteristic loudness of the engine. If the overall characteristic loudness of the engine meets the predetermined target of a reduction of more than 10% at the peak, then the scheme is determined as the final design scheme.
[0095] In some embodiments of the present invention, Figure 8(a) shows the characteristic loudness curves of the main field points. After the improvement, the characteristic loudness of the intake and exhaust sides is significantly reduced in the 13.9 ERB frequency band, and the characteristic loudness of the front end is significantly reduced in the 18.5 ERB frequency band. Figure 8(b) is a comparison of the average Moore characteristic loudness of the engine at four points before and after the structural design. The reduction of the average Moore characteristic loudness at the peak of the four points is greater than the predetermined target of 10%, which demonstrates the effectiveness of the method of the present invention.
[0096] The multi-flexible body dynamics modeling method of this invention effectively improves the accuracy of the model and provides modeling reference support for engine systems with balance shafts; the field point sound pressure level of the free sound field is calculated based on the fundamental wave technology theory, which can improve research efficiency and reduce costs compared with the test method; compared with the traditional A-weighted sound pressure level evaluation index, the Moore loudness evaluation index is consistent with the subjective evaluation results and provides a more accurate evaluation of engine radiated noise; an engine structure design method based on Moore loudness is simple in process, meets the requirements of engine lightweighting, and effectively reduces the structural radiated noise level of the engine.
[0097] The above description of the disclosed embodiments enables those skilled in the art to make or use the invention. Various modifications to these embodiments will be readily apparent to those skilled in the art, and the general principles defined herein may be implemented in other embodiments without departing from the spirit or scope of the invention. Therefore, the invention is not to be limited to the embodiments shown herein, but is to be accorded the widest scope consistent with the principles and novel features disclosed herein.
Claims
1. A method of engine structural design based on Moore loudness, characterized by, Includes the following steps: (1) Calculate the modal parameters of the engine housing assembly: Establish finite element models of the engine housing assembly, crankshaft assembly, and balance shaft, and calculate the natural frequencies and mode shapes of the housing assembly; (2) Calculate the vibration response of the housing surface: Reduce the finite element model of the engine housing assembly, crankshaft assembly and balance shaft to establish the multi-flexible body dynamics model of the engine, calculate the solution of the reduced model, that is, the vibration response of the engine housing surface nodes after reduction, and obtain the vibration velocity of the engine housing surface through the inverse reduction transformation. (3) Calculate the sound pressure level at the front end, intake side, exhaust side, and top of the engine, and the average sound pressure level at these four points: A free-field model of the engine is established using the fundamental wave technique. With the vibration velocity of the casing surface as the boundary condition, the sound pressure levels at the front end, intake side, exhaust side, and top of the engine are calculated, and the average sound pressure level P at these four points is also calculated. a ; (4) First, the sound pressure level of the sound at the field point is corrected using the outer ear transfer function and the middle ear transfer function. After the sound reaches the inner ear cochlea, the energy of all engine noise signals within each equivalent rectangular bandwidth is calculated, i.e., the ERB level, denoted as L. ERB ; (5) calculating an output excitation E of the auditory filter i ; (6) calculating the Moore characteristic loudness N of the engine in the frequency range of interest i ’; (7) Based on the calculated Moore characteristic loudness N i Based on the frequency band range corresponding to the peak value of the characteristic loudness curve, combined with the shell modal parameters in step (1) and the shell surface vibration acceleration in step (2), it is determined that the peak value of the curve is generated by shell resonance. (8) Engine structure redesign: By modifying the engine structure, the local stiffness of the structure is improved, the modal parameters of the structure are changed, and the frequency of the excitation source is avoided, so as to reduce the radiated noise level of the engine and improve the sound quality characteristics of the engine. The improved structure must meet two requirements: i. When designing the structure, it must be ensured that the stiffeners in the design area do not interfere with the internal parts of the engine. ii. In order to achieve the goals of lightweighting and economy, the overall weight must be controlled and the increase in overall weight after improvement must not exceed the preset value; (9) Vibration and noise quality calculation: Recalculate according to steps (2) to (6) to obtain the Moore characteristic loudness of the engine. If the overall characteristic loudness of the engine meets the predetermined target of a reduction of more than 10% at the peak, then the scheme is determined as the final design scheme.
2. The method of claim 1, wherein: Step (1) The finite element models of the housing assembly and the crankshaft assembly are both assemblies of multiple parts.
3. The method of claim 1, wherein: The multi-flexible body dynamics model in step (2) is a nonlinear time-domain dynamics model. The model considers the inertia and torsional stiffness of the flywheel end drive shaft, and the drive shaft is equivalent to three parts: the flexible coupling, the intermediate shaft and the motor shaft. The drive shaft is modeled using a parametric modeling method. A nonlinear spring damping model is used to represent the connection relationship between the main and auxiliary teeth of the scissor teeth, based on the structural form of the scissor teeth.
4. The method of claim 1, wherein: In step (2), the solution of the reduced model is calculated using an iterative method.
5. The method of claim 1, wherein: In step (3), the average sound pressure level P of the four field points a is calculated by the following formula: In the formula, P i The sound pressure level results of the four field points, respectively.
6. The engine structure design method based on Moore loudness according to claim 1, characterized in that, In step (4), L ERB The calculation formula is: In the formula: W a (j)=(1+p j g j )exp(-p j g j ) (3) g j =|f j -f i | / f i (4) p j = 4f j / E n (5) In the formula, f i E is the center frequency. n P is the equivalent rectangular bandwidth corresponding to the center frequency, and P0 is the reference sound pressure level; j W is the corrected sound pressure level at the j-th effective frequency; a (j) is the weighting function for the j-th effective frequency, representing the contribution of each effective frequency in the frequency band to the ERB level.
7. The Moore loudness engine design method of claim 1, wherein: In step (5), the formula for calculating the output excitation of the auditory filter is: where P0is a reference sound pressure, E0is a reference excitation, W b (j) is a shape function of the filter.
8. The method of claim 7, wherein, In step (5), the shape function expression of the filter is: In the formula: f i E is the center frequency. n L is the equivalent rectangular bandwidth corresponding to the center frequency. ERB Let be the sum of the energy of all engine noise signals within each equivalent rectangular bandwidth.
9. The engine configuration design method based on Moore loudness according to any one of claims 1 to 8, characterized in that, In step (6), the Moore characteristic loudness is: In the formula, N i Let E be the Moore characteristic loudness of the i-th auditory filter. THRQ G represents the energy level of the monoauricular audible valve, G is the low-level gain of the cochlear amplifier at a specific frequency, C is a coefficient, A is a frequency-dependent constant, and α is the exponential relationship between excitation and loudness.
10. The method of claim 1, wherein, In step (6), the total loudness is obtained by integrating the Moore characteristic loudness in the frequency domain.