多激励系统隔振降噪功率流及振声能量耦合理论与应用
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摘要
中央空调、风力机等多激励系统通常安装在同一空间内，由于系统内各种设备振动能量很大，同时随着结构的日趋轻薄，振动和噪声问题日益突出。本文以多项隔振降噪工程为背景，以薄板结构上大型热泵机组系统等隔振降噪模型为研究对象，综合运用导纳法、四端参数法分析了双层隔振系统、双激励双输出二级隔振系统、具有路径质量的单层隔振系统的振动传递路径的功率流，研究了薄板振动产生的振声能量耦合声场特性以及多激励系统中不同设备激励源对形成的振声声场的贡献，并应用振级落差隔振评价指标和功率流传递率隔振评价指标对隔振降噪效果进行了评价。
首先基于振动系统功率流隔振理论以及导纳理论、四端参数法和电-力类比等方法，建立了中央空调机房内振动设备的隔振系统模型。根据热泵机组的特点建立了符合实际条件的主机双激励双输出的二级隔振系统模型、支撑管路一级隔振系统、具有中间质量块的单激励单输出二级隔振系统和水泵机组的具有质量块的一级隔振系统。以单激励单输出二级隔振系统为例，计算并得到了传递功率流和以及输入、输出速度的变化曲线，明确了具有中间小质量的双层隔振系统的隔振设备重量、上层隔振器刚度、下层隔振器的阻尼以及中间质量块重量等参数对隔振效果的影响。计算结果表明，隔振器的阻尼对隔振效率影响很小，通过调节隔振器刚度和中间小质量块的重量可以调节二阶共振峰值的频率，达到避开设备工作频率和增强隔振降噪效果的目的。
其次，通过求解薄板振动微分方程和三维矩形空间声波波动方程，确定了当薄板以对称模态振动时能激起矩形空间内的对称声模态的振声耦合边界条件。通过实测振动数据对楼板结构振动产生的振声声场进行模拟，得出在频率在400Hz以下的区域，由于楼板结构的自阻尼，振声强、弱耦合差别很小，而400Hz以上的区域振声具有强耦合的特性，振声强耦合声场噪声值较大且与实测声场数据吻合。
以中央空调机房内的主要设备为研究对象，根据建立的隔振系统模型，通过计算分析得到了路径参数发生变化时隔振系统的输入、传递功率流，输入、输出速度以及传递载荷的变化曲线。对于热泵机组，空间条件允许的情况下增大中间质量块的重量对降低传递功率流最有效，而只有大幅降低隔振器刚度才能有效降低传递功率流，但必须保证机组启动时的平稳运行。水泵质量块可以起到调节一阶共振峰值位置并降低共振峰后频率的传递功率流，而且能降低设备工作重心使设备安全运行的效果，而增大水泵区域基础面密度也是经济实用的降低传递功率流的方法。
再后根据实测设备振动数据，应用ANSYS有限元分析软件和Virtual LabAcoustics声学仿真软件模拟了隔振前后机房内设备振动所形成的振声耦合声场，分析了热泵机组、水泵和管路振动对振声声场的频带噪声贡献以及各组成设备在整个隔振系统中比较敏感的隔振降噪频带。热泵机组振动对噪声的贡献是全频带的，这是由于其振动能量最大且位置最接近基础结构中心有关。而水泵机组多靠近机房角落所以不易激起低频振动，且本身振动能量相比热泵机组小多个数量级，所以水泵振动对振声声场的主要贡献频带在300Hz以上。虽然管路隔振支点很多，但总的振动能量多集中在200Hz以下，这与其产生的噪声声场高噪声频带主要集中在200Hz相吻合。
最后提出了参考功率流传递率评价隔振降噪效果的方法。通过绘制隔振降噪系统的速度振级落差曲线和功率流传递率曲线，并与模拟振声声场得到的频带降噪量频谱图相比较，得出以能量作为评价依据的功率流传递率曲线与频带降噪量曲线总体走势更加吻合的结论，为准确评价降噪效果提供依据。
Multi-incentive systems, such as central air conditioning system and wind turbine
system, are usually installed in the same space. Since the equipments’vibration energy ishuge and at the same time the structures become increasingly thin and light, vibration andnoise problems have become increasingly prominent. Based on a number of vibrationisolation and noise reduction engineering cases, large-scale heat pump unit systemvibration isolation model on the thin plate structure is chosen as research object. Vibrationtransfer path power flow characteristics of double-deck isolation system, double sourcesexciting and double output double-deck isolation system and single-deck isolation systemwith inertia block are analyzed by comprehensive use of four-end parameter method andmobility method. Vibration-acoustic energy coupling characteristics of sound fieldgenerated by civil thin plate vibration are studied and then analyze multi-incentivesystem’s different machines excitation sources’contribution to the sound field. Applyvibration level difference and power flow transmission ratio to evaluate vibration isolationand noise reduction effect.
Firstly, based on the vibration power flow isolation theory, mobility method, four-endparameter method and electricity-force analogy method, build vibration equipments’vibration isolation system models. By the heat pump unit’s features and the actual situation,build double sources exciting and double output double-deck vibration isolation system,supporting pipe single-deck vibration isolation system，single source exciting and singleoutput double-deck vibration isolation system with middle mass and water pump’ssingle-deck vibration isolation system with middle mass. For the single source exciting andsingle output double-deck isolation system with middle mass, calculate and obtain curvesof the system’s imported and transmitted power flow, machine and foundation speedamplitude, and identify the impact on vibration isolation effect when the vibration isolationsystem with little intermediate mass’path parameters change, such as machine mass, upper isolator stiffness, lower isolator damping and intermediate mass weight. Results show thatthe isolator damping has little effect on the vibration isolation efficiency. Second-orderresonance peak frequency can be changed by adjusting the isolator stiffness and theintermediate mass weight, and then avoid machine working frequency and improve thevibration isolation effect.
Secondly, by solving the thin plate vibration differential equation and threedimensional rectangular space acoustic wave equation, knows that symmetric acousticmodes in the rectangular space are promoted when the thin plat vibrates with symmetricalmode. Vibration and acoustic are strong coupling. Simulate the vibration-acoustic couplingsound field generated by the thin plat vibration using the actual vibration data. Resultsshow that vibration-acoustic strong or weak coupling are no difference below 400Hzbecause of the damping of the floor structure. Above 400Hz vibration and acoustic arestrong coupling and simulating results are consistent with the actual measured sound fielddata.
Thirdly, main equipments in the central air conditioning room as research projects,based on the vibration isolation models, calculate and obtain curves of the system’simported and transmitted power flow, machine and foundation speed amplitude andtransmitted forces to the foundation with the change of the path parameters. For the heatpump unit, increase the weight of the floating raft mass is the most effective method toreduce the transmitted power flow only if the vibration isolation space is enough. Thoughsmall stiffness of the isolator can reach the same result with the increasing weight of themass, it must assure machine’s stable running when it starts. Pump mass can adjust theresonance peak position and reduce the transmitted power flow above the resonancefrequency, and lower the machine’s gravitational center to ensure safe operation. Increasesurface density is an economic and practical method to reduce transmitted power flow.
Fourthly, based on the actual measured vibration data, apply ANSYS finite elementanalysis software and Virtual Lab Acoustics acoustic simulation software to simulate thevibration-acoustic coupling sound field generated by the machines’vibration before andafter vibration isolation. Analyze the contribution of the heat pump vibration, the pumpvibration and the pipe vibration to the vibration-acoustic coupling sound field. Point outthe machines most important vibration isolation band. For heat pump’s vibration energy isthe largest and its location is the nearest to the center of the foundation, so its contributionto the sound field is the whole band. Since the pump is installed near to the corner of the foundation and its vibration energy is much smaller than heat pump’s, so it is difficult tostimulate low-frequency vibration. Contribution to the noise sound field is mainlyconcentrated above 300Hz. Though having many pipes’isolation fulcrums, the totalvibrational energy is concentrated below 200Hz and it is consistent with the noisespectrum characteristic.
Finally, propose a reference of power flow transmission ratio to evaluate vibrationisolation and noise reduction effect. Draw vibration isolation and noise reduction system’scurves of speed vibration level difference and power flow transmission ratio, and comparewith the band noise reduction spectrogram, results show that power flow transmission ratiocurve is more consistent with the band noise reduction spectrogram, it can be used as anevaluation criteria for the judgment of the noise reduction effect.
首先基于振动系统功率流隔振理论以及导纳理论、四端参数法和电-力类比等方法，建立了中央空调机房内振动设备的隔振系统模型。根据热泵机组的特点建立了符合实际条件的主机双激励双输出的二级隔振系统模型、支撑管路一级隔振系统、具有中间质量块的单激励单输出二级隔振系统和水泵机组的具有质量块的一级隔振系统。以单激励单输出二级隔振系统为例，计算并得到了传递功率流和以及输入、输出速度的变化曲线，明确了具有中间小质量的双层隔振系统的隔振设备重量、上层隔振器刚度、下层隔振器的阻尼以及中间质量块重量等参数对隔振效果的影响。计算结果表明，隔振器的阻尼对隔振效率影响很小，通过调节隔振器刚度和中间小质量块的重量可以调节二阶共振峰值的频率，达到避开设备工作频率和增强隔振降噪效果的目的。
其次，通过求解薄板振动微分方程和三维矩形空间声波波动方程，确定了当薄板以对称模态振动时能激起矩形空间内的对称声模态的振声耦合边界条件。通过实测振动数据对楼板结构振动产生的振声声场进行模拟，得出在频率在400Hz以下的区域，由于楼板结构的自阻尼，振声强、弱耦合差别很小，而400Hz以上的区域振声具有强耦合的特性，振声强耦合声场噪声值较大且与实测声场数据吻合。
以中央空调机房内的主要设备为研究对象，根据建立的隔振系统模型，通过计算分析得到了路径参数发生变化时隔振系统的输入、传递功率流，输入、输出速度以及传递载荷的变化曲线。对于热泵机组，空间条件允许的情况下增大中间质量块的重量对降低传递功率流最有效，而只有大幅降低隔振器刚度才能有效降低传递功率流，但必须保证机组启动时的平稳运行。水泵质量块可以起到调节一阶共振峰值位置并降低共振峰后频率的传递功率流，而且能降低设备工作重心使设备安全运行的效果，而增大水泵区域基础面密度也是经济实用的降低传递功率流的方法。
再后根据实测设备振动数据，应用ANSYS有限元分析软件和Virtual LabAcoustics声学仿真软件模拟了隔振前后机房内设备振动所形成的振声耦合声场，分析了热泵机组、水泵和管路振动对振声声场的频带噪声贡献以及各组成设备在整个隔振系统中比较敏感的隔振降噪频带。热泵机组振动对噪声的贡献是全频带的，这是由于其振动能量最大且位置最接近基础结构中心有关。而水泵机组多靠近机房角落所以不易激起低频振动，且本身振动能量相比热泵机组小多个数量级，所以水泵振动对振声声场的主要贡献频带在300Hz以上。虽然管路隔振支点很多，但总的振动能量多集中在200Hz以下，这与其产生的噪声声场高噪声频带主要集中在200Hz相吻合。
最后提出了参考功率流传递率评价隔振降噪效果的方法。通过绘制隔振降噪系统的速度振级落差曲线和功率流传递率曲线，并与模拟振声声场得到的频带降噪量频谱图相比较，得出以能量作为评价依据的功率流传递率曲线与频带降噪量曲线总体走势更加吻合的结论，为准确评价降噪效果提供依据。
Multi-incentive systems, such as central air conditioning system and wind turbine
system, are usually installed in the same space. Since the equipments’vibration energy ishuge and at the same time the structures become increasingly thin and light, vibration andnoise problems have become increasingly prominent. Based on a number of vibrationisolation and noise reduction engineering cases, large-scale heat pump unit systemvibration isolation model on the thin plate structure is chosen as research object. Vibrationtransfer path power flow characteristics of double-deck isolation system, double sourcesexciting and double output double-deck isolation system and single-deck isolation systemwith inertia block are analyzed by comprehensive use of four-end parameter method andmobility method. Vibration-acoustic energy coupling characteristics of sound fieldgenerated by civil thin plate vibration are studied and then analyze multi-incentivesystem’s different machines excitation sources’contribution to the sound field. Applyvibration level difference and power flow transmission ratio to evaluate vibration isolationand noise reduction effect.
Firstly, based on the vibration power flow isolation theory, mobility method, four-endparameter method and electricity-force analogy method, build vibration equipments’vibration isolation system models. By the heat pump unit’s features and the actual situation,build double sources exciting and double output double-deck vibration isolation system,supporting pipe single-deck vibration isolation system，single source exciting and singleoutput double-deck vibration isolation system with middle mass and water pump’ssingle-deck vibration isolation system with middle mass. For the single source exciting andsingle output double-deck isolation system with middle mass, calculate and obtain curvesof the system’s imported and transmitted power flow, machine and foundation speedamplitude, and identify the impact on vibration isolation effect when the vibration isolationsystem with little intermediate mass’path parameters change, such as machine mass, upper isolator stiffness, lower isolator damping and intermediate mass weight. Results show thatthe isolator damping has little effect on the vibration isolation efficiency. Second-orderresonance peak frequency can be changed by adjusting the isolator stiffness and theintermediate mass weight, and then avoid machine working frequency and improve thevibration isolation effect.
Secondly, by solving the thin plate vibration differential equation and threedimensional rectangular space acoustic wave equation, knows that symmetric acousticmodes in the rectangular space are promoted when the thin plat vibrates with symmetricalmode. Vibration and acoustic are strong coupling. Simulate the vibration-acoustic couplingsound field generated by the thin plat vibration using the actual vibration data. Resultsshow that vibration-acoustic strong or weak coupling are no difference below 400Hzbecause of the damping of the floor structure. Above 400Hz vibration and acoustic arestrong coupling and simulating results are consistent with the actual measured sound fielddata.
Thirdly, main equipments in the central air conditioning room as research projects,based on the vibration isolation models, calculate and obtain curves of the system’simported and transmitted power flow, machine and foundation speed amplitude andtransmitted forces to the foundation with the change of the path parameters. For the heatpump unit, increase the weight of the floating raft mass is the most effective method toreduce the transmitted power flow only if the vibration isolation space is enough. Thoughsmall stiffness of the isolator can reach the same result with the increasing weight of themass, it must assure machine’s stable running when it starts. Pump mass can adjust theresonance peak position and reduce the transmitted power flow above the resonancefrequency, and lower the machine’s gravitational center to ensure safe operation. Increasesurface density is an economic and practical method to reduce transmitted power flow.
Fourthly, based on the actual measured vibration data, apply ANSYS finite elementanalysis software and Virtual Lab Acoustics acoustic simulation software to simulate thevibration-acoustic coupling sound field generated by the machines’vibration before andafter vibration isolation. Analyze the contribution of the heat pump vibration, the pumpvibration and the pipe vibration to the vibration-acoustic coupling sound field. Point outthe machines most important vibration isolation band. For heat pump’s vibration energy isthe largest and its location is the nearest to the center of the foundation, so its contributionto the sound field is the whole band. Since the pump is installed near to the corner of the foundation and its vibration energy is much smaller than heat pump’s, so it is difficult tostimulate low-frequency vibration. Contribution to the noise sound field is mainlyconcentrated above 300Hz. Though having many pipes’isolation fulcrums, the totalvibrational energy is concentrated below 200Hz and it is consistent with the noisespectrum characteristic.
Finally, propose a reference of power flow transmission ratio to evaluate vibrationisolation and noise reduction effect. Draw vibration isolation and noise reduction system’scurves of speed vibration level difference and power flow transmission ratio, and comparewith the band noise reduction spectrogram, results show that power flow transmission ratiocurve is more consistent with the band noise reduction spectrogram, it can be used as anevaluation criteria for the judgment of the noise reduction effect.
引文
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