压气机中间级转子叶片强迫振动响应快速分析

韩乐; 王延荣; 魏大盛; 杨唯超; 李迪; 邹正平

doi:10.13224/j.cnki.jasp.20220314

压气机中间级转子叶片强迫振动响应快速分析

doi: 10.13224/j.cnki.jasp.20220314

韩乐^{1, 2,},
王延荣²,
魏大盛^2,*, ,,
杨唯超²,
李迪³,
邹正平¹

1.
北京航空航天大学航空发动机研究院，北京 100191
2.
北京航空航天大学能源与动力工程学院，北京 100191
3.
中国航发商用航空发动机有限责任公司，上海 200241

基金项目: 国家科技重大专项(2017-Ⅳ-0002-0039，J2019-Ⅳ-0012-0080）

详细信息

作者简介:
韩乐(1989-)，男，助理研究员，博士，从事叶轮机械流致振动、结构强度和流动稳定性研究。E-mail: hanle@buaa.edu.cn

通讯作者:
魏大盛(1978-)，男，教授、博士生导师，博士，从事航空发动机结构强度、振动与疲劳寿命研究。E-mail：dasheng.w@163.com

中图分类号: V231.1
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出版历程
- 收稿日期: 2022-05-07
- 网络出版日期: 2022-10-10

Fast analysis of forced vibration response on intermediate stage compressor rotor blade

HAN Le^{1, 2
,},
WANG Yanrong²,
WEI Dasheng^{2,*
, ,},
YANG Weichao²,
LI Di³,
ZOU Zhengping¹

1.
Research Institute of Aero-Engine，Beihang University，Beijing 100191，China
2.
School of Energy and Power Engineering，Beihang University，Beijing 100191，China
3.
Commercial Aircraft Engine Company Limited，Aero Engine Corporation of China，Shanghai 200241，China

摘要

摘要:
以多级压气机第二级转子叶片为例，对其共振时的振动特征进行了快速分析。一方面通过进口边界条件修正对压气机模型进行减缩，并结合谐波方法提高了非定常气动力计算效率，另一方面利用系统在模态空间的频响关系和扫频技术进一步提高了求解效率，特别是当工程上亟需开展非定常气动力作用下的强迫振动响应分析时，可高效评估共振时叶片振动特征。结果表明：第二级转子叶片主要受上游静子激励，下游静子的势扰动影响有限，在共振转速附近易激起第15阶模态共振，叶片尖区有较高的振动应力，经瞬态响应分析，在给定的阻尼和工况下，所考查位置和方向的振动应力约71 MPa，利用扫频等方式评估共振时非定常气动力引起的转子叶片振动应力约为92 MPa。所形成的分析方法与流程有一定的普适性。
- 多级压气机 /
- 转子叶片 /
- 强迫振动响应 /
- 共振应力 /
- 快速分析
Abstract:
The second stage rotor blade of multi-stage compressor was investigated to efficiently calculate the response caused by aerodynamic excitation. The compressor was reduced through the inlet boundary correction, and combined with the harmonic method to improve the calculation efficiency of unsteady aerodynamic force. The relationship between frequency and response in modal space and the harmonic analysis method were used to further improve the solution efficiency. When the forced vibration response analysis under aerodynamic excitation was urgently required in engineering, the blade vibration features at resonance can be evaluated efficiently. The results showed that the excitations of the second rotor blade was mainly subject to the upstream stator passing frequencies, and the influence of the downstream stator was limited. The 15th modal shape was excited near the resonance condition and there was significant vibration stress in the blade tip area. Under the given damping and simulated working conditions, the vibration stress was about 71 MPa in the examined position and direction obtained by the transient response analysis, and it was about 92 MPa under the resonance condition evaluated by the harmonic analysis method.
- multi-stage compressor /
- rotor blade /
- forced vibration response /
- resonance stress /
- fast analysis

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图 1 强迫振动响应解耦法分析流程

Figure 1. Analysis process of forced response by using the weak decoupling method

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图 2 数值模型计算域

Figure 2. Computing domain of the numerical model

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图 3 非线性谐波法与时间推进法结果对比

Figure 3. Comparison between nonlinear harmonic method and time marching method^[17]

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图 4 R2的结构网格、监测点以及材料参数

Figure 4. Solid Grids, monitors and material parameters of R2

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图 5 压气机特性图

Figure 5. Performance of the compressor

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图 6 简化模型与原模型不同叶高的压力相似度

Figure 6. Pressure similarity of different blade spans between the simplified and the original models

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图 7 99%叶高压力分布

Figure 7. Distribution of pressure at 99% blade span

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图 8 99%叶高Ma分布

Figure 8. Distribution of Ma at 99% blade span

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图 9 叶片表面非定常压力各阶谐波分布

Figure 9. Distribution of pressure harmonics on the blade surface

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图 10 90%叶高不同弦向位置压力谱

Figure 10. Pressure spectrum of 90% blade span at different chord lengths

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图 11 第二级转子叶片Campbell图

Figure 11. Campbell diagram of R2

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图 12 R2叶片模态结果

Figure 12. Modal analysis results of R2

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图 13 强迫振动响应收敛情况

Figure 13. Convergence of forced response

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图 14 各监测点振动应力

Figure 14. Vibration stresses at monitors

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图 15 共振点附近振动应力分布

Figure 15. Distribution of vibration stresses near the resonance point

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图 16 共振点附近监测点1585的振动应力

Figure 16. Vibration stress of Monitor 1585 near the resonance point

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图 17 共振点振动应力与误差分布

Figure 17. Vibration stress and error distribution at the resonance point

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表 1 S1、R2和S2无量纲设计参数

Table 1. Normalized design parameters of S1, R2 and S2

参数	S1	R2	S2
叶片数	46	37	62
叶尖前缘半径	1.014	1	0.982
叶尖尾缘半径	1.004	0.986	0.971
叶根前缘半径	0.628	0.653	0.7
叶根尾缘半径	0.650	0.697	0.718
叶尖弦长C	0.191	0.264	0.157
叶根弦长	0.123	0.224	0.101
叶尖间隙	0	0.28% C	0

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表 2 不同转速下气动激励与100%转速下的偏差

Table 2. Deviation of aerodynamic excitation between different rotating speeds and 100% rotating speed

归一化转速	模态力/N	误差/%
0.95	32.23	−1.6
0.98	31.24	−4.5
1.0	32.74	0
1.02	34.58	5.6
1.05	30.86	−5.7

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表 3 转子叶片表面监测点

Table 3. Monitors on blade surface of R2

模态阶次	最大主应力	监测点编号	主应力方向
模态阶次	最大主应力	监测点编号	n_x	n_y	n_z
1	S1	1326	0.97	−0.17	−0.16
2	S1	874	0.96	0.15	−0.22
3	S1	1375	1.00	−0.02	−0.00
15	S1	1585	−0.16	0.93	0.33
16	S1	1585	0.06	0.91	0.41

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表 4 第二级转子叶片100%转速各阶模态频率和阻尼比

Table 4. Modal frequencies and damping ratios of R2 at 100%rotating speed

模态阶次	频率/Hz	ζ_i /%
1	709.46	0.030
2	1701.2	0.030
3	2116.5	0.033
15	11013	0.138
16	11817	0.148

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表 5 不同方法的计算耗时

Table 5. Calculation time of different methods

方法	归一化时间
共振模态频响关系分析	1
扫频方法	10
瞬态分析	10³～10⁴

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