A method for controlling acceleration roaring sound

Abstract

The technical solution employed by the present invention is a method for controlling acceleration roar noise, comprising the following steps: conducting an acceleration noise test on a prototype vehicle and extracting an audio spectrum of the roar noise generated by the main engine order; calculating the contribution of the noise from the left, right, and rear suspension paths to the interior noise based on the audio spectrum of the main engine order, and selecting the path with the largest contribution to the interior noise as the main noise transmission path; optimizing the vehicle body structure under the main transmission path; obtaining a vehicle body interior noise transfer function curve and identifying problematic points in the vehicle body interior noise transfer function; and, based on the problematic points in the vehicle body interior noise transfer function, selecting methods to increase the engine dynamic stiffness or optimize the beam system, panels, or component modes of the vehicle body structure, thereby optimizing the vehicle body noise, vibration, and harshness performance. This invention can address this type of noise phenomenon in the early stages of vehicle model project development, reducing development costs and improving the overall vehicle NVH performance.

Description

Technical Field

[0001] The present invention belongs to the technical field of complete vehicle production, and particularly relates to a method for controlling acceleration roaring sound. Background Art

[0002] Acceleration roar is a noise phenomenon that occurs inside a vehicle when accelerating. This sound can cause the driver and passengers to feel an uncomfortable sensation of strong pressure on the eardrums. Long-term driving can also cause symptoms such as headaches and nausea, seriously reducing vehicle comfort. The roar audio frequency band generally ranges from 20Hz to 200Hz. There are many mechanisms for its generation. The method described in this patent is mainly to address the acceleration roar inside the vehicle caused by substandard vehicle NVH performance.

[0003] For gasoline-powered vehicles, the mechanism of acceleration whine caused by substandard NVH performance is as follows: the engine's order excitation force is transmitted to the front subframe, which is connected to the vehicle body and transmits the vibration to the body beam system. If the body component modes along this transmission path are not frequency-avoided or other NVH performance characteristics of the vehicle body are substandard, the excitation force is amplified, exacerbating the vibration of the body panel structure. The panel vibration, through fluid-structure coupling with the vehicle's acoustic cavity, causes periodic changes in the volume of the enclosed air inside the vehicle, generating a high-energy whine that presses against the eardrum. Summary of the Invention

[0004] The purpose of the present invention is to address the shortcomings of the above-mentioned background technology and provide a method for controlling acceleration roar noise. This method can solve this type of noise phenomenon in the early stage of vehicle project development (mule car stage), reduce development costs, and improve the NVH performance of the entire vehicle. The technical solution adopted by the present invention is: a method for controlling acceleration roar noise, comprising the following steps:

[0005] Conduct acceleration noise tests on prototype vehicles and extract the roaring audio spectrogram generated by the main orders of the engine;

[0006] Based on the roar audio spectrogram generated by the main engine order, the contribution of the noise from the left, right, and rear suspension paths to the interior noise of the vehicle is calculated. The path with the largest contribution to the interior noise is selected as the main noise transmission path. The vehicle body structure under the main transmission path is optimized.

[0007] Obtain the body and interior noise transfer function curve and identify the problem points of the body and interior noise transfer function;

[0008] Based on the problem points of the vehicle body and interior noise transfer function, we choose to increase the engine dynamic stiffness or optimize the beam system, panels or component modes of the vehicle body structure to optimize the vehicle body noise, vibration and harshness performance.

[0009] In the above technical solution, the process of conducting an acceleration noise test on a prototype vehicle and extracting the roar audio spectrogram generated by the main order of the engine includes:

[0010] A microphone is arranged at the driver's inner ear to collect the noise value at the ear point; acceleration sensors are arranged at the active and passive ends of the engine suspension to collect the acceleration of the suspension vibration during acceleration; based on the collected noise value at the ear point and the acceleration of the suspension vibration, a roar audio spectrum generated by the main order of the engine is constructed; the vertical axis of the roar audio spectrum is the noise amplitude, and the horizontal axis is the engine speed; a roar sound target value is set, and the engine speed when the main order roar sound exceeds the roar sound target value is determined through the roar audio spectrum.

[0011] In the above technical solution, the calculation process of the contribution of the noise inside the vehicle on a certain path includes:

[0012] Calculating the engine frequency at that moment based on the engine speed when the main order roar sound exceeds the target roar sound value;

[0013] Calculate the forces in the x, y, and z directions on the passive end of the engine mount for the path based on the engine frequency at that moment;

[0014] Based on the forces in the x, y, and z directions at the passive end of the engine mount of the path, the noise values ​​generated by the path in the x, y, and z directions are calculated respectively;

[0015] The sum of the noise values ​​generated by the path in the x, y, and z directions is calculated as the total sound pressure value along the path;

[0016] The proportion of the total sound pressure value under this path to the sum of the total sound pressure values ​​generated by all paths is calculated as the contribution of the in-vehicle noise of this path.

[0017] In the above technical solution, the process of identifying the problem points of the interior noise transfer function of the vehicle body includes:

[0018] If the noise transfer function curve in the vehicle has a peak, obtain the engine's dynamic stiffness curve; calculate whether the dynamic stiffness in the set frequency band meets the standard;

[0019] If the dynamic stiffness does not meet the standard, it is determined that the problem is caused by insufficient stiffness at the installation point of the excitation source, and the structure at the installation point of the excitation source is modified;

[0020] If the dynamic stiffness meets the requirements, the problem is determined to be caused by vibration of the body panels or beams, and then the problem is located in the specific body structure based on the body noise transmission diagram;

[0021] If peaks appear simultaneously in the dynamic stiffness curve and the in-vehicle noise transfer function curve at a certain frequency band, it is determined that the problem is caused by the mode of the auxiliary components of the excitation source, and modal testing is then performed or a corresponding dynamic vibration absorber is developed to lock in the specific frequency point.

[0022] In the above technical solution, the process of optimizing the beam system of the vehicle body structure includes: first, optimizing the layout of the beam system: the overall layout of each beam system forms a complete closed structure; second, optimizing the beam cross-section: designing the cross-section of the beam into a closed shape to avoid the appearance of an open cross-section; and finally, optimizing the stiffness of the beam system connector: the connector adopts an integrated design; adding materials with greater stiffness inside the connector; and designing a reinforced structure inside the connector.

[0023] In the above technical solution, the process of optimizing the panels of the vehicle body structure includes: first, optimizing the structure and arrangement of the beam system corresponding to the panels, improving the stiffness of the panels and optimizing the modes of the panels, and avoiding the appearance of large-area flat panels on the vehicle body; second, adding reinforcing ribs to the panels to increase the surface stiffness of the panels, and at the same time arranging the reinforcing ribs of the panels to change the phase of the panels when they vibrate, so that the vibrations within the panels cancel each other out; stamping the center of the panels into an arc; finally, adding damping plates or reinforcing plates to the panels to reduce the speed of the panels when they vibrate.

[0024] In the above technical solution, the process of optimizing the component modes of the vehicle body structure includes: using simulation methods to calculate the modal strain energy of the structure to determine the area with larger vibration deformation under the mode, and then performing targeted optimization of the corresponding structure.

[0025] In the above technical solution, the process of optimizing the engine dynamic stiffness of the vehicle body structure includes:

[0026] The dynamic stiffness in the 0-50Hz frequency band does not meet the target requirements, and the overall structure of the vehicle is modified;

[0027] The dynamic stiffness in the 50-100 Hz frequency band does not meet the target requirements, and the semi-global structure of the vehicle is modified;

[0028] The dynamic stiffness in the 100-200 Hz frequency band does not meet the target requirements, and the structure of some components needs to be modified.

[0029] In the above technical solution, the calculation formula for the force in any of the x, y, and z directions at the passive end of the engine mount in a certain path is as follows:

[0030]

[0031] Where a represents the vibration acceleration of the active end of the engine mount, f i is the engine frequency, F is the force when the passive end of the engine mount vibrates, and K is the static stiffness of the mount rubber in each direction of the path;

[0032] The x, y, z forces at the passive end of the engine mount are calculated using the x, y, z vibration accelerations at the active end of the engine mount and the x, y, z static stiffnesses of the mount rubber.

[0033] In the above technical solution, the following formula is used to calculate the noise value P generated by the path in each direction:

[0034]

[0035] Among them, NTF is calculated by the engine frequency f i The amplitude of the vehicle body noise transfer function at the location is obtained through experiments; the forces in the x, y, and z directions at the passive end of the engine mount passing through the path are used to calculate the noise values ​​P generated in the x, y, and z directions of the path.

[0036] The beneficial effects of the present invention are: the present invention can quickly solve the phenomenon of acceleration roar caused by the NVH problem of the vehicle body, reduce the driver's roar under certain working conditions of the vehicle, reduce the ear pressure of the driver and passengers, and improve comfort. The present invention can completely solve the problem of acceleration roar before the mold of the vehicle body parts is formed, and avoid the situation where the vehicle body mold needs to be modified in the later stage, which significantly saves material costs. Compared with the means of reducing the roar in the car by active control at the end of the vehicle model development: such as ANC, ENC, RNC, etc., the present invention can solve this type of noise problem at the mule car stage, greatly reducing the time cost. The present invention can not only be used to solve the roar problem, but also as a means of quickly verifying the roar frequency: such as calculating the noise proportion of each path, determining the corresponding main transmission path, and then adding dynamic vibration absorbers or mass blocks to the vehicle body parts under the path. BRIEF DESCRIPTION OF THE DRAWINGS

[0037] Figure 1 It is a schematic diagram of the process of the present invention.

[0038] Figure 2 This is the vehicle body structure noise transmission diagram of the present invention.

[0039] Figure 3 This is a diagram of the second-order roaring sound in the car during acceleration according to the present invention.

[0040] Figure 4 2 is a diagram showing the proportion of noise inside the vehicle according to this specific embodiment.

[0041] Figure 5 This is the excitation force transmission path diagram of this specific embodiment.

[0042] Figure 6.1 is the NTF curve under torsional mount excitation.

[0043] Figure 6.2is the dynamic stiffness curve under torsional mount excitation.

[0044] Figure 7.1 is the modal strain energy cloud diagram.

[0045] Figure 7.2 This is the IPI curve before and after optimization.

[0046] Figure 7.3 NTF curves before and after optimization.

[0047] Figure 8 It optimizes the second-order roar sound in the front row. DETAILED DESCRIPTION

[0048] The present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments to facilitate a clear understanding of the present invention, but they do not constitute a limitation to the present invention.

[0049] The present invention provides a method for controlling acceleration roar noise, which improves acceleration roar noise by optimizing the NVH performance of the vehicle body. The specific contents include judging the NVH performance of the whole vehicle, calculating the noise transmission contribution, and identifying the NVH problem points of the vehicle body. The flowchart of the acceleration roar noise control is shown in Figure 1 The specific contents are as follows:

[0050] Step 1: Determine the NVH performance of the vehicle:

[0051] This step requires an acceleration noise test on the prototype vehicle. During the test, a microphone is placed near the driver's inner ear to capture noise at that point. Accelerometers are placed at the active and passive ends of the engine mount to capture acceleration vibrations during acceleration. After the test, the data is post-processed to extract a spectrogram of the boom generated by the main engine orders. The vertical axis of the spectrogram represents the noise amplitude in dB (unweighted), and the horizontal axis represents the speed in rpm. An appropriate target value for the boom is set. See Table 1 for a target value setting table. Based on the target value, the speed at which the boom exceeds the standard is obtained from the spectrogram.

[0052] Table 1 Relationship between vehicle type grade and target value of roar noise

[0053]

[0054] The relationship between engine speed and frequency is shown in the following formula:

[0055]

[0056] In formula (1), f i is the frequency, n is the engine main order, selected based on the engine type. For a four-cylinder engine, use 2, and for a six-cylinder engine, use 3. rpm is the engine speed.

[0057] Step 2: Calculate the noise transfer contribution:

[0058] When the car accelerates, the main-order vibration of the engine is transmitted to the car body through three paths: the left mount, the right mount, and the rear mount. The noise inside the car is the sum of the noise generated by these three transmission paths. Since each path transmits different vibrations to the car body, the amount of noise generated is also different. By calculating the noise proportion of each path, the main transmission path can be determined. The body structure under this path is optimized with a focus on optimization. Each transmission path is further divided into three directions: x, y, and z. The noise generated by vibrations in these three directions is the noise under this path. Taking the left suspension vibration transmission path as an example, the following steps are used to calculate the proportion of noise generated during its vibration:

[0059] Step 1: Calculate the x-direction force at the passive end of the engine mount.

[0060]

[0061] In formula (1), a Lx X-axis vibration acceleration of the active end of the engine mount, f i is the engine frequency, F Lx K is the force in the x direction when the passive end of the engine mount vibrates. x is the static stiffness of the left mount rubber in the x direction. The force F in the y direction of the passive end of the engine mount Ly , z-direction force F at the passive end of the engine mount Lz The corresponding parameters are calculated using formula (1). The x-, y-, and z-axis vibration accelerations of the engine mount active end are obtained by installing a PCB vibration sensor on the engine mount bracket. This sensor is a three-phase sensor that can simultaneously collect data in three directions. The x-, y-, and z-axis static stiffness parameters of the mount rubber are provided by the rubber manufacturer and can be referred to in the rubber's factory report.

[0062] Step 2: Calculate the noise value P generated by the x-path Lx Noise is the convolution of force and the vehicle body noise transfer function. In the logarithmic domain, the convolution calculation becomes a linear calculation, and the calculation formula is as follows:

[0063]

[0064] In formula (2), NTF Lx Take the calculated engine frequency f i The amplitude of the vehicle body noise transfer function at the location is obtained through experiments. The noise value P generated by the y-direction path Ly , the noise value P generated by the z-direction path Lz It is calculated by using the corresponding parameters using formula (2).

[0065] Step 3: Calculate the total sound pressure value along the left suspension transmission path

[0066] 2OL Left =20LgP Lx +20LgP Ly +20LgP Lz (3)

[0067] In formula (3), P Left P is the interior noise value caused by the left suspension vibration. Lx 、P Ly 、P Lz These are the interior noises caused by the vibrations of the left suspension in the x, y, and z directions.

[0068] Step 4: Calculate the contribution of the noise generated along this path to the noise inside the vehicle

[0069]

[0070] In formula (4), P Left 、P Right 、P Rear They are the interior noise generated by the vibration of the left suspension, right suspension and rear suspension respectively. Right 、P Rear The calculation is performed from the first to the third step with the corresponding parameters.

[0071] By repeating the above steps, the contribution of the noise generated in the three paths of left suspension, right suspension and rear suspension to the interior noise is obtained; the contribution of the three suspension vibrations to the interior noise is compared, and the path with the largest contribution is determined as the main transmission path.

[0072] The main transfer path is combined with the vehicle body noise schematic (such as Figure 2 ), select the vehicle body structure under the optimized main transfer path.

[0073] Step 3: Identify NVH problem points of the vehicle body structure along the main transmission path:

[0074] The core indicator of vehicle body NVH performance is the in-vehicle noise transfer function (NTF). Through experiments, the vehicle body NTF curve is obtained and then NVH problem areas are identified. Optimizing the vehicle body NTF improves vehicle body NVH performance. The vehicle body NTF is strongly correlated with the engine's dynamic stiffness. This paper first deduces the relationship between dynamic stiffness K and vehicle body NTF, and then provides qualitative conclusions. The specific details are as follows:

[0075]

[0076] In formula (5), α, ρ, and c are the calculation coefficient, air density, and sound speed, respectively.

[0077] Conclusion: The vehicle body NTF is inversely proportional to the engine dynamic stiffness. Improving the engine dynamic stiffness is beneficial to improving the vehicle body NVH performance.

[0078] The vehicle body NTF problem point identification criteria formulated by the present invention are as follows:

[0079] Criterion 1: If a peak appears in the vehicle body NTF curve, first calculate whether the dynamic stiffness value meets the standard. Obtain the dynamic stiffness curve of the engine through the test, and use formula (6) to calculate the dynamic stiffness K at the set frequency band. The value on the engine stiffness curve is the value of the denominator IPI in formula (6), and each frequency corresponds to a value.

[0080]

[0081] Where f is the calculation frequency, which starts at 50 and ends at a frequency that depends on the vehicle model, as described in Table 2.

[0082] According to the specific vehicle model, a specific dynamic stiffness target is set. The principle for setting the dynamic stiffness target value is: 5000β in the x and y directions, and 10000γ in the z direction; β and γ are calculation coefficients, which are related to the vehicle model grade. The present invention summarizes the relationship between the vehicle model grade and the calculation coefficient, as shown in Table 2. If the dynamic stiffness does not meet the standard, it is determined that the problem is caused by insufficient stiffness at the installation point of the excitation source, and then the structure at the installation point of the excitation source is modified. Specifically, if the calculated dynamic stiffness value is less than the target value, it is determined that the stiffness at the installation point of the excitation source is insufficient, and the local structure of the installation point of the excitation source can be modified. If it is greater than the target value, it means that the stiffness of the installation point of the excitation source meets the requirements, and the body panel or beam system needs to be modified.

[0083] Table 2 Relationship between vehicle type grade and dynamic stiffness calculation coefficient

[0084]

[0085] Criterion 2: If the dynamic stiffness meets the requirements, the peak in the vehicle's NTF curve is caused by vibration of the body panels or beams. Operational Deformation Modal Analysis (ODS) and Modal Strain Energy Analysis (ESE) are performed at the NTF peak frequency. Visual dynamic graphs are used to identify the beam or panel with the greatest deformation at that frequency. A finite element model with an acoustic cavity is then used to perform a node contribution analysis to determine whether the structure with the greatest deformation is caused by a particular beam or panel.

[0086] Criterion 3: If peaks appear simultaneously in the dynamic stiffness curve and the in-vehicle NTF curve at a certain frequency band, it can be determined that the cause of the vehicle body NVH exceeding the standard is caused by the modal of the excitation source's accessory components. Modal tests should be conducted or corresponding dynamic vibration absorbers should be developed to lock in the specific frequency point.

[0087] The above three criteria are arranged in a certain order, with Criterion One having the highest level. Similarly, in the early stages of vehicle development, there will basically be problems that exceed the standards, which are the identification contents of Criterion One.

[0088] Step 4: Optimize vehicle body NVH performance:

[0089] Improving the engine's dynamic stiffness and optimizing the beams and panels of the vehicle body structure can improve the vehicle's NFT, thereby optimizing the vehicle's NVH performance. The optimization of the vehicle body structure is achieved through simulation. This paper develops corresponding optimization criteria for engine dynamic stiffness and different vehicle body structures. The specific contents are as follows:

[0090] Optimization criterion 1: Optimize the engine dynamic stiffness. The calculation formula of the engine dynamic stiffness is as follows:

[0091]

[0092] In formula (7), K r 、M r 、C r are the stiffness, mass, and damping of the r-th order mode, Φ Lr is the r-th order vibration mode vector at the measuring point L.

[0093] The above formula shows that the engine dynamic stiffness includes the response of all body structures, including the engine, partial body, semi-partial body, and the entire body. The optimization approach for engine dynamic stiffness varies in different frequency bands, which can be summarized as follows:

[0094] 1) The dynamic stiffness within 0-50Hz does not meet the target requirements, and the overall structure of the vehicle needs to be modified; optimization criteria 2, 3, and 4 need to be executed in sequence;

[0095] 2) The dynamic stiffness within 50-100Hz does not meet the target requirements, and the semi-global structure of the vehicle needs to be modified; optimization criteria 2, 3, and 4 need to be executed in sequence;

[0096] 3) The dynamic stiffness within 100-200 Hz does not meet the target requirements, and the local component structure needs to be modified; only optimization criterion four needs to be executed in sequence.

[0097] Optimization criterion two: Optimizing the beam system. First, optimize the beam system layout. The optimization method is to ensure that the overall arrangement of each beam system forms a complete, closed structure. Next, optimize the beam cross-section. The optimization method is to design the beam cross-section to be as closed as possible, avoiding open cross-sections. Finally, optimize the stiffness of the beam system connectors. The optimization method is to adopt an integrated connector design, add high-stiffness materials inside the connector, and design reinforcement structures inside the connector.

[0098] Optimization criterion three: Panel optimization. First, optimize the structure and layout of the corresponding beam system to improve panel stiffness and optimize panel modalities, avoiding the appearance of large flat panels on the vehicle body. Second, add a certain number of ribs to the panel to increase its surface stiffness. Appropriately arranging the ribs shifts the phase of panel vibration, allowing vibrations within the panel to cancel each other out. This reduces overall panel vibration and prevents vibration from compressing the acoustic cavity. Panel ribs can be primarily categorized as unidirectional, bidirectional, and random. Unidirectional ribs are formed by stamping only along the transverse or longitudinal direction of the panel. Bidirectional ribs are formed along both the transverse and longitudinal directions. Irregular and bidirectional ribbed panels exhibit higher stiffness and modalities than unidirectional ribbed panels. Stamping the center of the panel into an arc can also significantly improve panel stiffness and modalities. Arc-shaped panels significantly enhance the high-frequency modes of thin panels. Finally, damping sheets or reinforcement plates are added to the panels to reduce the speed at which the panels vibrate.

[0099] Optimization criterion four: Optimization of component modes. Using simulation methods, the modal strain energy of the structure is calculated to determine the area with larger vibration deformation under this mode, and then targeted optimization of the corresponding structure is carried out.

[0100] The present invention also provides a computer-readable storage medium, on which a method program for controlling an acceleration roaring sound is stored. When the method program for controlling an acceleration roaring sound is executed by a processor, the steps of the method described in the above technical solution are implemented.

[0101] The present invention provides a specific embodiment: when a certain SUV is accelerating, the second-order roaring sound exceeds the standard at 3750r / min. The noise frequency corresponding to 3750r / min is calculated to be 125Hz using the formula (1) in this patent. The spectrum diagram of the second-order roaring sound during acceleration is shown in Figure 3 Using the noise contribution calculation method provided by the present invention, it is calculated that the noise generated by the rear suspension vibration accounts for 50%. The calculation results are shown in Figure 4 , so the main transmission path of the noise is the engine rear mount. Combined with the vehicle body noise principle diagram, the excitation force transmission path diagram is developed, such as Figure 5 As shown. Through the vehicle body dynamic stiffness and vehicle body NTF test, as shown Figure 6.1 and Figure 6.2 , combined with the judgment criteria provided by the present invention, it was finally confirmed that the excessive roaring sound was caused by the coupling of the second-order excitation frequency of the engine and the natural frequency of the front subframe of 125Hz. Using simulation methods, the modal strain energy of the front subframe was calculated, and the modal strain energy cloud diagram ( Figure 7.1 ), it was found that the place with the largest deformation at this frequency was the clevis at the connection between the subframe and the body. After optimizing the clevis of the front subframe, the dynamic stiffness and the simulation results of the interior noise transfer function before and after optimization were compared. Figure 7.2 、 Figure 7.3 .

[0102] The results show that under the excitation of the torsional mount at 125Hz, the amplitude of the vehicle body noise transfer function decreased by 13dB compared with the pre-optimization level, and the dynamic stiffness amplitude in the 20Hz-300Hz frequency band decreased by 4dB compared with the pre-optimization level. A new subframe was trial-produced, and the prototype vehicle with the new subframe was subjected to acceleration noise testing. The test results showed that the second-order acceleration roar at 125Hz decreased by 11dB, verifying the effectiveness of the optimization scheme. The optimization effect is shown in Figure 8

[0103] The contents not described in detail in this specification belong to the prior art known to those skilled in the art.

Claims

1. A method for controlling acceleration roaring sound, characterized by: The following steps are involved: Acceleration noise tests were conducted on the prototype vehicle to extract the roaring audio spectrogram generated by the main order of the engine. A microphone was placed near the driver's inner ear to collect the noise value at the ear point. Acceleration sensors were placed at the active and passive vibration ends of the engine mount to collect the acceleration of the mount vibration during acceleration. Based on the collected ear-point noise values ​​and suspension vibration acceleration, a roar audio spectrogram generated by the main order of the engine is constructed; the vertical axis of the roar audio spectrogram is the noise amplitude, and the horizontal axis is the engine speed; a roar sound target value is set, and the engine speed when the main order roar sound exceeds the roar sound target value is determined by the roar audio spectrogram; Based on the roar audio spectrogram generated by the main engine order, the contribution of the noise from the left, right, and rear suspension paths to the interior noise is calculated. The path with the largest contribution to the interior noise is selected as the main noise transmission path. Select the body structure under the optimized main transfer path; Obtain the body and interior noise transfer function curve and identify the problem points of the body and interior noise transfer function; If the interior noise transfer function curve has a peak, the engine dynamic stiffness curve is obtained; Calculate whether the dynamic stiffness of the set frequency band meets the standard; If the dynamic stiffness does not meet the requirements, the problem is determined to be caused by insufficient stiffness at the excitation source installation point, and the structure at the excitation source installation point is modified. If the dynamic stiffness meets the requirements, the problem is determined to be caused by vibration of the body panel or beam system, and the problem is then located in the specific body structure by combining the body noise transfer function diagram. If the dynamic stiffness curve and the interior noise transfer function curve show peaks at the same time in a certain frequency band, the problem is determined to be caused by the mode of the excitation source accessory components, and modal testing is then carried out or a corresponding dynamic vibration absorber is developed to target the specific frequency point. Based on the problem points of the vehicle interior noise transfer function, the engine dynamic stiffness can be increased or the beams, panels, or component modes of the vehicle body structure can be optimized to optimize the vehicle body noise, vibration, and harshness performance. The process of optimizing the engine dynamic stiffness of the body structure includes: The dynamic stiffness in the 0-50Hz frequency band does not meet the target requirements, and the overall structure of the vehicle is modified; The dynamic stiffness in the 50-100 Hz frequency band does not meet the target requirements, and the semi-global structure of the vehicle is modified; The dynamic stiffness in the 100~200Hz frequency band does not meet the target requirements, and the structure of local components needs to be modified.

2. A method according to claim 1, characterized in that: The calculation process of the noise contribution of the vehicle interior noise on a certain path includes: Calculating the engine frequency at that moment based on the engine speed when the main order roar sound exceeds the target roar sound value; Calculate the forces in the x, y, and z directions on the passive end of the engine mount for the path based on the engine frequency at that moment; Based on the forces in the x, y, and z directions at the passive end of the engine mount of the path, the noise values ​​generated by the path in the x, y, and z directions are calculated respectively; The sum of the noise values ​​generated by the path in the x, y, and z directions is calculated as the total sound pressure value along the path; The proportion of the total sound pressure value under this path to the sum of the total sound pressure values ​​generated by all paths is calculated as the contribution of the in-vehicle noise of this path.

3. A method according to claim 1, characterized in that: The process of optimizing the beam system of the vehicle body structure includes: first, optimizing the layout of the beam system: the overall layout of each beam system forms a complete closed structure; second, optimizing the beam cross-section: designing the cross-section of the beam into a closed shape to avoid the appearance of an open cross-section; and finally, optimizing the stiffness of the beam system connector: the connector adopts an integrated design; adding materials with greater stiffness inside the connector; and designing a reinforced structure inside the connector.

4. A method according to claim 1, characterized in that: The process of optimizing the panels of the vehicle body structure includes: first, optimizing the structure and layout of the beam system corresponding to the panel, improving the stiffness of the panel and optimizing the mode of the panel to avoid the appearance of large-area flat panels on the body; second, adding reinforcing ribs to the panel to increase the panel surface stiffness, and at the same time arranging the reinforcing ribs of the panel to change the phase of the panel when it vibrates, so that the vibrations within the panel cancel each other out; punching the center of the panel into an arc; finally, adding damping plates or reinforcing plates to the panel to reduce the speed of the panel when it vibrates.

5. A method according to claim 1, characterized in that: The process of optimizing the component modes of the vehicle body structure includes: using simulation methods to calculate the modal strain energy of the structure to determine the areas with larger vibration deformation under this mode, and then performing targeted optimization of the corresponding structure.

6. A method according to claim 2, characterized in that: The formula for calculating the force in any of the x, y, and z directions at the passive end of the engine mount for a certain path is as follows: ； Where a represents the vibration acceleration of the active end of the engine mount, f i is the engine frequency, F is the force when the passive end of the engine mount vibrates, and K is the static stiffness of the mount rubber in each direction of the path; The x, y, z forces at the passive end of the engine mount are calculated using the x, y, z vibration accelerations at the active end of the engine mount and the x, y, z static stiffnesses of the mount rubber.

7. A method according to claim 6, characterized in that: The noise value P generated by this path in all directions is calculated using the following formula: ； Among them, NTF is calculated engine frequency The amplitude of the vehicle body noise transfer function at the location is obtained through experiments; the forces in the x, y, and z directions at the passive end of the engine mount passing through the path are used to calculate the noise values ​​P generated in the x, y, and z directions of the path.

Images