Deformation coordination relationship between two plates loose (tight but bolts are not stressed) bolts are supported by the original length of the two plates (excluding nut deformation): L = 2D set bolts and the stress level of the two plates in the material Within the elastic range, the deformation increment of the bolt operating state relative to the pre-tightening state is: $L=LEbAb$F where: the elastic modulus of the Eb bolt material; the effective bearing area of ​​the Ab bolt. The compression increment (rebound amount) of the operating state of the Steel Plate relative to the pre-tightening state is: $D=DEsAs$N where: the elastic modulus of the Es steel plate material; As) the pressed surface area of ​​the steel plate. The elastic modulus of the bolt material is equal to the elastic modulus of the Steel Sheet material (the difference is extremely small), that is, Eb=Es. Then: $F=AbAsF1+AbAs and the bolt operation state pulling force is: Fw=$F+Fp= Fp+AbAsF1+AbAs (the effective bearing area Ab of the bolt is very small compared with the area A of the pressed surface of the steel plate, so the incremental force of the bolt operating state is less than that of the preloaded state. In the general bolt arrangement principle, when the group bolts act, the center distance between the bolt and the bolt is not less than 3 times the diameter of the bolt itself. If the bolt diameter is d and the bolt center distance is 3d, the average area of ​​the steel plate under which each bolt is pressed is approximately As=P(3d)2Ab/4, and the bolt stress area is approximately Ab=Pd/42. Therefore: Ab/AsT1/8. Therefore: When $F(F/8)/(1+1/8)=F/9T0.11 is bolted, the following important facts cannot be ignored: According to the high-strength bolt tensile test results, When the applied external tension exceeds the pre-tightening force (ie, F>Fp), after the pulling force is removed, the bolt will have a loosening phenomenon in which the pre-tightening force is reduced. Only when the applied external pulling force meets the condition: F<0.9Fp, the relaxation phenomenon with reduced preloading force does not occur. For safety reasons, the ratio of the external tension of each bolt to its pre-tensioning tension should be controlled at: F/Fp0.8, the maximum tension of the bolt operation state is: Fw=Fp+$F=Fp+0.11F=Fp+0 .=1.088Fp, that is, the bolt operating state tension is only 8.8% larger than the preload tension. In order to control the section stress level under the tension state of the bolt operation within 80% of the yield strength of the material, that is, control: Fw/Abe<0.8Re(16) where: Abe))) effective sectional area of ​​the bolt; yield strength of the Re material. This is not the case. It is a very important situation that the flange to which the bolt is connected will actually bend. The consequence of this situation is that the tension of the operating state bolt will be much greater than the results of the above mechanical analysis. In order to illustrate the nature of the problem, the two-bolt compression of two sections of the same size is discussed. For example, the bending deformation is exaggerated to the limit, and only the end points of the two beams are in contact. At this time, the elongation of the bolt under the action of F is $l1, and $F is: $F=EbAb$l1/l. The stress increment is: $R=Eb$l1/l. Generally, Carbon Steel Eb=200@103MPa, it can be seen that even if $l1/l is only 1/1000, the stress increment of the bolt is as high as 200MPa. Therefore, pay special attention to the rigidity of the flange when designing the flange bolt connection. Flanges (steel plates) with sufficient rigidity not only contribute to the safety of the bolts, but are also critical to the sealing performance. This point needs special emphasis. At present, the calculation method of the bolt connection of the flange in the design standard is only the check of the strength, and must not only be satisfied by the calculation of the strength, but pay special attention to the importance of the rigidity of the flange rather than the strength. In the presence of a gasket, the core problem is the sealing of the flange bolts in the container. It is therefore inevitable that a seal is present in the joint. The presence of the gasket also causes a significant deviation from the actual mechanical model described above. First, it should be corrected to: L = 2D + Dg (in the formula: Dg) the free-state thickness of the gasket. It should be corrected to: $L=2$D+$Dg where: $Dg is the amount of rebound of the seal under F, and: $Dg=DgEgAg$N where: Eg) the modulus of elasticity of the gasket material ;Ag) The surface area of ​​the gasket corresponding to the basic sealing width (full tightness in the pre-tightened state). The formula should also be corrected to: $D=DEsAg$N Although the content has changed, the expression can still be maintained. $F=FEgAgEbAb+DgEgAg2DEbAbEgEs+Dg2D+1 From the formula: Because 2D is ten or even dozens of times larger than Dg, the pressing surface is a toroidal surface, not the simplest mechanical model 0, Ag is larger than Ab The multiple is always less than 8, and Eg is much smaller than Es or Eb. Namely: EgAgEbAb<1, EbEs<1, Dg2D<1, and DgEgAg2DEbAb<1. Because AgAb>1, it can be seen that due to the presence of the gasket, the tension of the bolt operating state is greater than that of the pre-tightening state, but it does not exceed the external tension of the bolt. half. The specific numerical size is subject to the characteristic data Eg and Ag of the gasket. It can also be seen from the formula that the wider the gasket (the larger the Ag), the smaller the tension increment $F. To ensure a seal, the wider the gasket, the greater the required bolt pretension. There is a huge difficulty in the presence of a gasket: the characteristic data Eg of the gasket lacks accurate data. Fortunately, there is a standard check calculation method. Although the standard/compression method and the pressure and gasket coefficient are not accurate, the concept is not clear, and it is not reasonable, but it is a safe calculation after many years of practice. law. The difference between the actual applied load and the simplest mechanical model is that the flange on the container is bolted, and the load is not the concentrated force but the uniform internal pressure. The difference from the former is that the latter causes the flange to generate internal pressure expansion, which has an influence on the magnitude and distribution of the flange stress, and the coupling bolt is of course also implicated. This is pointed out by many scholars. Other issues that should be considered The factors that should be considered above are far from all the factors that should be considered in the design of the bolted joint structure, as well as methods such as bolt preload control, material selection, performance grade selection, nut material and strength matching, Problems such as the accuracy of the joints and the like need to be carefully considered. 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