﻿ 法兰式无键液压联轴器安装及运行稳定性验证分析
 舰船科学技术  2020, Vol. 42 Issue (11): 59-62    DOI: 10.3404/j.issn.1672-7649.2020.11.012 PDF

Analysis on installation and operation stability of flange type hydraulic coupling with non key
WANG Rui, FAN Hua-tao, SONG Qiang
State Key Laboratory of Deep-sea Manned Vehicles, China Ship Scientific Research Center, Wuxi 214082, China
Abstract: Aiming at the problem of hydraulic control of hydraulic installation of flange non key coupling of ship shafting and the anti impact performance of hydraulic coupling. A new type flange type hydraulic coupling as the researchobject, buildingthe three-dimensional model of flange type keyless hydraulic couplings, finite element simulation calculating of contact equivalent stress、radial hydraulic pressure、axial thrust and coupling impact resistancebased on the requires of the theory calculation of steel ships of the classification, and the results of simulation analysis and theoretical analysis results are verified. The results show that when the Length of installation reaches a certain position, the axial thrust simulation value and the theoretical calculation value deviation increase due to the increase of the friction force at the contact edge position; In the case of stable operation of hydraulic coupling, the stress of the inner sleeve of the hydraulic coupling is less than that of the outer sleeve, and the weak link of the inner sleeve is in the two ends, After the installation of the hydraulic coupling to run the test, found that the stability of the stability of the coupling is very good.
Key words: the coupling     hydraulic installation     stability     length of installation     radial oil pressure
0 引　言

1 计算原理 1.1 推入量的计算公式

 $\begin{split} {S_1} = & \frac{1}{K}\left[ {47750 \times {{10}^4}\frac{{{N_{\rm{e}}}}}{{A{n_e}}}\left( {\frac{{{C_1}}}{{{E_1}}} + \frac{{{C_2}}}{{{E_2}}}} \right) + } \right.\\ & \left. {({a_2} - {a_1})(35 - t){d_1} + 0.03} \right]\text{，} \end{split}$ (1)
 ${S_2} = \frac{1}{K}\left[ {0.7{\sigma _s}d{}_1\frac{{K_2^2 - 1}}{{\sqrt {3K_2^4 + 1} }}(\frac{{{C_1}}}{{{E_1}}} + \frac{{{C_2}}}{{E{}_2}}) - ({a_2} - {a_1}){d_1}t} \right]\text{。}$ (2)

1.2 液压安装的基础力

 $P = \frac{{{S_F}T}}{{A{B_1}}}\left( { - \frac{{{S_F}K}}{2} + \sqrt {\mu _1^2 + {B_1}{{\left( {\frac{{{F_V}}}{T}} \right)}^2}} } \right)\text{，}$ (3)
 ${F_0} = A\left( {0.002 + \frac{K}{{20}}} \right) \cdot \left[ {{P_{35}} + \frac{{18}}{{{B_2}}}({\alpha _2} - {\alpha _1})} \right]\text{。}$ (4)

1.3 轴向推力和径向油压

 ${F_1} = Ap\left( {{\mu _2} + K/2} \right)\text{，}$ (5)

 $P = 1.1SK/({d_1}{B_2})\text{。}$ (6)
1.4 应力的计算

 $\begin{split} \sigma = & \sqrt {\frac{1}{2}\left[ {{{\left( {{\sigma _1} - {\sigma _2}} \right)}^2} + {{\left( {{\sigma _2} - {\sigma _3}} \right)}^2} + {{\left( {{\sigma _3} - {\sigma _2}} \right)}^2}} \right]} = \\ & 0.9p\left[ {\sqrt {3K_2^4 + 1} /\left( {K_2^4 - 1} \right)} \right]\text{。} \end{split}$ (7)

2 实例计算分析

2.1 模型的网格划分和约束的施加

 图 1 液压联轴器安装和抗冲击三维模型图 Fig. 1 Three dimensional model of hydraulic coupling installation and shock resistance

2.2 等效应力

 图 2 液压联轴器应力 Fig. 2 Stress of hydraulic coupling

2.3 径向油压

 图 3 径向油压的仿真值和理论值对比 Fig. 3 Comparison of simulation value and theoretical value of radial oil pressure

2.4 轴向推力

 图 4 轴向推力理论和仿真值的对比 Fig. 4 Comparison of theoretical and simulation values of axial thrust

2.5 无键联轴器液压安装方案

3 联轴器的稳定性分析

 图 5 最大扭矩下联轴器外套和内套的应力分布云图 Fig. 5 Stress distribution nephogram of coupling outer sleeve and inner sleeve under maximum torque

4 结　语

1）由于液压联轴器实际安装过程中边缘等效应力的奇异性，接触的边缘位置具有应力集中的现象，接触面的摩擦力异常增大。当推入量达到一定的值以后，轴向推力仿真的计算结果大于理论计算的值。为了保证液压联轴器正常的工作，安装后的极大等效应力值应该小于联轴器的材料屈服强度。

2）联轴器液压安装过程中径向油压和轴向推力的仿真值和理论值基本都是呈线性增大，当推入量达到某位置以后，仿真值和理论计算值偏差开始增大，并且径向油压的理论计算值几乎都小于有限元的仿真值。为了保证液压联轴器正常的工作，径向油压和轴向推力理论值与仿真值的最大误差要低于船级社钢制海船入级规范中的7%的要求。

3）在液压联轴器运行稳定的情况下，液压联轴器内套的应力小于外套，由于受到边缘等效应力集中现象的影响，内套的最大应力出现在边缘位置，联轴器承内套的薄弱环节出现在两端。

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