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No, its not a truss, elements have fixed connections. If I connect beams at their centroids, not only I am using more material than in reality (and more self weight in the model), but am I not considering more section strength also?

I want to model a connection between two CHS and one SHS in FEM beam model. How do I make the model to consider that the SHS does not reach the center of the CHSs, but only their exterior surface? I attach two possible model solutions. I believe the first one is unrealistic since the length...

Yes, that is what I believe causes the torque. Torque is moment Mx with the black axis represented as X axis.

I am calculating a padeye with a shackle, in which the force coming from a cable is deviated a certain angle in the both represented views (vertical and horizontal deviation), due to cable misalignment. Does the fixed surface in red have a torsion moment?. I can easily find a tension force...

Both supports are also bolts.

Yes, I know very well the statics and linear theory. But what happens in reality? Doesnt the bolt have plastic deformation on only one side? How do large deformations and nonlinear geometric analysis benefit one support over the other? Is there a nonlinear FEM example that proves it?

I have a rectangular steel plate with a bolt in the middle and pinned supports at each extreme. The bolt transmits a force to one side of the plate. Theoretically (following linear statics), both supports have the same reaction force value. But is this tru in reality? Doesnt large deformations...

Four meters. The impact is an increase in nozzle required nozzle thickness as per UG-37. The question is simple, should I use design or test pressure in UG-37 / UG-45 when test pressure is 3 times the design pressure as stated by client?

I read UG-99 and still don't known whether if I should use design or test pressure in UG-37 / UG-45 when test pressure is 3 times the design pressure as stated by client. Yes, UG-37 requirements of nozzle thickness are greater for test pressure than design pressure. Yes, that's the client...

But it doesnt say if I must use the test pressure for nozzle thickness and pad calculations.

I dont know that. But since test is performed right after assembly, there is no corrosion yet. I calculate nozzle pads and thicknesses with 3mm corrosion, so I believe Im on the safe side. Still, I would like to know if I have to justify the calculations with test pressure in ASME.

Client insists on hydro test pressure to be 3 times the design pressure

My design inputs of client just state: hydraulic test pressure of 0.15MPa

All right. Still, my test pressure exceeds 1,5 x DP.

1.5 x DP / MAWP 1.5 x 0.5 / 0.5 = 1.5

I also calculate for external pressure FV (1 bar). Why set the design internal pressure to 1 barg? What if DP=0.5 barg and MAWP=0.5 barg? Then, test pressure of 1.5 barg would make sense according to 1.5 x DP/MAWP.

It is a pressure vessel with design pressure of 0.05MPa internal. But client states that test pressure is 0.15MPa internal. Where does ASME state that nozzles should not be dimensioned for test pressure?

I am doing reinforcement pad calculations for nozzles with 0.05MPa design internal pressure with ASME VIII - Div. 1 2017. My input data from client specifies a test pressure requirement of 0.15MPa internal pressure (I know its hydraulic test). Should I consider 0.15MPa or 0.05MPa in UG-37...

It is an example picture, I havent started with the mesh yet. Yes you are right, but Im looking for a general answer and to understand why.

I wonder which is the most theoretically better and accurate mesh of a circular section subjected to perpendicular distributed load and with interest in stresses in the whole plane. I have seen several examples, using either triangular elements only or a mix between triangular and rectangular...