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figures and diff

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gtheler 5 vuotta sitten
vanhempi
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8 muutettua tiedostoa jossa 280 lisäystä ja 388 poistoa
Jaettu näkymä Näytä diff tilastot
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@@ -9,6 +9,9 @@ hash.yaml
*.out
*.msh
*.vtk
diff.pdf
diff.tex
old.tex
refs
nafems4.html
nafems4.tex

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if [ ! -e old.tex ]; then
echo "error: old.tex does not exist"
exit 1
fi

./make.sh tex
latexdiff old.tex nafems4.tex > diff.tex
sed -i s/svg/pdf/ diff.tex
xelatex diff
xelatex diff

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nafems4.md Näytä tiedosto

@@ -103,7 +103,7 @@ For the case study, as the loads come principally from operational loads, the AS

![A fatigue or $S$-$N$ curve for two steels.](SN.svg){#fig:SN width=100%}

It should be noted that the fatigue curves are obtained in a particular load case, namely purely-periodic and one-dimensional, which cannot be directly generalised to other three-dimensional cases. Also, any real-life case will be subject to a mixture of complex cycles given by a stress time history and not to pure periodic conditions. The application of the curve data implies a set of simplifications and assumptions that are translated into different possible “rules” for composing real-life cycles. The ASME $S$-$N$ curve also adopt tow safety factors which increase the stress amplitude and reduce the number of cycles respectively. All these intermediate steps render the analysis of fatigue into a conservative computation scheme. Therefore, when a fatigue assessment performed using the fatigue curve method arrives at the conclusion that “fatigue is expected to occur after ten thousand cycles” what it actually means is “we are sure fatigue will not occur before ten thousand cycles, yet it may not occur before one hundred thousand or even more.”
It should be noted that the fatigue curves are obtained in a particular load case, namely purely-periodic and one-dimensional, which cannot be directly generalised to other three-dimensional cases. Also, any real-life case will be subject to a mixture of complex cycles given by a stress time history and not to pure periodic conditions. The application of the curve data implies a set of simplifications and assumptions that are translated into different possible “rules” for composing real-life cycles. The ASME $S$-$N$ curves also adopt two safety factors which increase the stress amplitude and reduce the number of cycles respectively. All these intermediate steps render the analysis of fatigue into a conservative computation scheme. Therefore, when a fatigue assessment performed using the fatigue curve method arrives at the conclusion that “fatigue is expected to occur after ten thousand cycles” what it actually means is “we are sure fatigue will not occur before ten thousand cycles, yet it may not occur before one hundred thousand or even more.”

## Solid mechanics, or what we are taught at college

@@ -197,7 +197,7 @@ Remember that when any solid body is subject to external forces, it has to react
* There are no azimuthal displacements\ $u_\theta$ because the pipe is fully symmetric around the axis.
* There are only radial displacements\ $u_r$ and they depend only on the radial coordinate\ $r$ and not on the axial position\ $z$ or on the azimuthal angle\ $\theta$.

What does this mean? Well, that overall the whole pipe expands a little bit radially with the inner face being displaced more than the external surface (use your imagination or wait until we get to [@fig:ur]). How does it expand?
What does [@eq:ur] mean? Well, that overall the whole pipe expands a little bit radially with the inner face being displaced more than the external surface (use your imagination or wait until we get to [@fig:ur]). How does it expand?

1. linearly with the pressure, i.e. twice the pressure means twice the displacement, and
2. inversely proportional to the Young’s Modulus\ $E$ divided by $1+\nu$, i.e. the more resistant the material, the less radial displacements.
@@ -356,7 +356,7 @@ dnl As I wanted to illustrate in [@sec:five], it is very important to decide wha

Since we already agreed there is no way to obtain analytical expressions for the stresses in this general case, we need to employ a numerical scheme to solve the equations. Of course we are choosing the finite element methods, but keep in mind that there are a lot of other methods for solving partial differential equations: finite differences, finite volumes, modal methods, etc. In particular, finite elements compute nodal values (i.e. displacements and stresses at discrete points in space) and then provide a way to interpolate back the results into any other arbitrary point of the domain. If the method is applied correctly, a mesh refinement will lead to improved results---at the cost of needing an exponentially-increasing computing power. In the limit of infinite number of nodes, the FEM results converge to the actual solution of the original PDEs.

For instance, [fig:ur] shows how some results obtained with the finite element method using different number and order of elements compare to the analytical solution from the previous section. The bullets correspond to the nodal values of the radial displacement obtained with the finite element method. A detailed analysis of the analytical solution and how results obtained with finite elements converge with respect to the mesh size can be found in the report “On convergence of linearised stresses in an infinite pipe computed using the finite element method.” See [@sec:online].
For instance, [@fig:ur] shows how some results obtained with the finite element method using different number and order of elements compare to the analytical solution from the previous section. The bullets correspond to the nodal values of the radial displacement obtained with the finite element method. A detailed analysis of the analytical solution and how results obtained with finite elements converge with respect to the mesh size can be found in the report “On convergence of linearised stresses in an infinite pipe computed using the finite element method.” See [@sec:online].

![Comparison of $u_r(t)$ computed with finite elements and\ [@eq:ur].](ur.svg){#fig:ur width=90%}

@@ -371,9 +371,9 @@ First of all, we should note that we need to solve
iii. the elastic problem to obtain the stress tensor needed to compute the alternating stress to enter into the fatigue curve.

::::: {#fig:pt}
![Case #1](pt-1.svg){#fig:pt1 width=95%}
![Case #1: postulated transient data.](pt-1.svg){#fig:pt1 width=95%}

![Case #2](pt-2.svg){#fig:pt2 width=95%}
![Case #2: actual plant data from a real transient.](pt-2.svg){#fig:pt2 width=95%}

Typical transient pressure and temperature conditions for the pipes.
:::::
@@ -387,7 +387,7 @@ So for each time\ $t$ of the operational (or incidental) transient such as the o

### Thermal transient {#sec:thermal}

Let us invoke our imagination once again. Assume in one part of the transients the temperature of the water inside the pipes falls from say 300ºC down to 100ºC in a couple of minutes, stays at 100ºC for another couple of minutes and then gets back to 300ºC (similarly to [@fig:pt]). The temperature within the bulk of the pipes changes as times evolves. The internal wall of the pipes follow the transient temperature (it might be exactly equal or close to it through the [Newton’s law of cooling](https://en.wikipedia.org/wiki/Newton%27s_law_of_cooling)). If the pipe was in a state of uniform temperature, the ramp in the internal wall will start cooling the bulk of the pipe creating a transient thermal gradient. Due to thermal inertia effects, the temperature can have a non-trivial dependence when the ramps start or end (think about it!). So we need to compute a real transient heat transfer problem with convective boundary conditions because any other usual tricks like computing a sequence of steady-state computations for different times would not be able to recover these non-trivial distributions.
Let us invoke our imagination once again. Assume in one part of the transients the temperature of the water inside the pipes falls from say 250ºC down to room temperature in a couple of minutes, stays cool for another couple of minutes and then gets back to 250ºC (similarly to [@fig:pt]). The temperature within the bulk of the pipes changes as times evolves. The internal wall of the pipes follow the transient temperature (it might be exactly equal or close to it through the [Newton’s law of cooling](https://en.wikipedia.org/wiki/Newton%27s_law_of_cooling)). If the pipe was in a state of uniform temperature, the ramp in the internal wall will start cooling the bulk of the pipe creating a transient thermal gradient. Due to thermal inertia effects, the temperature can have a non-trivial dependence when the ramps start or end (think about it!). So we need to compute a real transient heat transfer problem with convective boundary conditions because any other usual tricks like computing a sequence of steady-state computations for different times would not be able to recover these non-trivial distributions.

Remember the main issue of the fatigue analysis in these systems is to analyse what happens around the location of changes of piping classes where different materials (i.e. different expansion coefficients) are present, potentially causing high stresses due to differential thermal expansion (or contraction) under transient conditions. Therefore, even though we are dealing with pipes we cannot use beam or circular shell elements, because we need to take into account the three-dimensional effects of the temperature distribution along the pipe thickness. And even if we could, there are some tees that connect pipes with different nominal diameters that have a non-trivial geometry, such as the weldolet-type junction shown in\ [@fig:weldolet-cad;@fig:weldolet-mesh]. In this case, there are a number of SCLs (Stress Classification Lines) that go through the pipe’s thickness at both sides of the material interface as illustrated in\ [@fig:weldolet-scls]. It is in these locations that fatigue is to be evaluated.


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@@ -1,31 +1,4 @@
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