Latest / Mechanical Engineering Made Simple / (#149) The Fatal Disconnect Between CAD and Steel

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(#149) The Fatal Disconnect Between CAD and Steel · Mechanical Engineering Made Simple

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Welcome back, design junkies. School taught you to fear
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failure. The real world will reveal
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failure is your greatest teacher.
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This episode of Mechanical Engineering Made Simple comes to
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you by a listener request. If you didn't know, we do those.
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So if there's a topic you want us to break down, explain, or
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explore, just drop us a comment. We prioritize getting the
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content you want in front of you.
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So let's dive into today's requested topic.
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We are taking a massive stack of listener questions and
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condensing them into, well, a single brutal reality check.
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Yeah, a reality check is definitely the right word for
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it. Right, because right now,
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somewhere in the world there is a Florida State engineered
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mathematically perfect 500 ton pressure vessel sitting on a
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barge. I mean, the CAD model was
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flawless. The finite element analysis was
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entirely green. All the colors on the screen
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were exactly what the engineering manager wanted to
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see. Exactly.
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But because of a single like 1 inch miscalculation in the
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diameter of a lifting, or just a fundamental misunderstanding of
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how a wave through multiplies apparent gravity, that vessel is
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about to tear itself off the deck.
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It's going to sink the barge and bankrupt the project.
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And the scary part is, it happens far more often than
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anyone in this industry wants to it.
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Oh, absolutely. We spend thousands of hours
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analyzing the operational stresses of a fluid inside a
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machine, but we entirely gloss over the violently chaotic
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reality of how that machine is, you know, fabricated, inspected,
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and transported. Right, because it's easy to just
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look at the screen. Yeah, a perfect CAD model is a
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wonderful thing on a screen, but if that design cannot survive
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the welding arc or the radio graphic inspector's
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interpretation and the open ocean, I mean, it is nothing
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more than expensive digital abstract art.
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We are dealing with factual, truthful mechanics today.
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No pretty promises. OK, let's untack this.
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Because of the gap between digital perfection and physical
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reality, it really starts the moment you print the blueprint.
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I call it the Blueprint Illusion.
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I like that the blueprint illusion.
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Yeah, because before any massive structure can be built, it has
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to be drafted. And the biggest pain point I see
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young engineers running into is assuming a mathematically
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perfect 3D model translates directly to a perfect physical
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part on the shop floor. That is easily the most
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dangerous assumption you can make.
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The fundamental disconnect is that university mechanics
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courses they treat materials as mathematically continuous right
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and geometrically perfect. So they teach you that a one
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inch hole is exactly 1 inch. Exactly, but the physical shop
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floor operates in a world of entropy.
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A1 inch hole is actually a highly complex, totally
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irregular cylinder. Oh good, the drill bit chatters
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and stuff. Right.
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It's diameter varies from top to bottom.
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It might lean slightly to the left the location of its center
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shifts depending on the thermal expansion of the steel in a hot
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machine shop, or even just the microscopic wear on the drill
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bit over the course of an 8 hour shift.
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Which is exactly why the American Society of Mechanical
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Engineers, or ASME, created the Y 14.5 standard for geometric
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dimensioning and tolerance in GD and T.
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Yes, and it's not just a suggestion of how to draw lines
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on a paper. It's a rigorous, legally binding
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language. It defines exactly how much
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physical imperfection is legally allowed before you have to
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literally throw the part in the scrap bin.
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And yet it is a language that is routinely misunderstood by so
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many people. To even begin speaking GD and T,
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you have to establish a physical coordinate system in reality,
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not just on your computer screen.
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Right, the datum reference frame.
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Exactly the datum reference frame.
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Whenever you have a physical part floating in space, whether
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it is a tiny aerospace bracket or a massive engine block, it
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possesses 6° of freedom. OK, let's break those down for
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everyone. You have your three
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translational degrees, right? Movement along the XY and Z
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axis. Yep.
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So moving forward and backward, sliding left and right, and
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moving up and down. But what always trips people up
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are the three rotational degrees of freedom, pitch, Yaw and roll
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around those exact same XY and Z axis, usually denoted as UV and
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W in kinematics. Right, so consider what happens
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when a machinist or an inspector tries to measure a physical
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part. If you just set a block of
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machine steel on a flat granite surface plate, you haven't
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actually locked it down. You've only constrained three of
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those 6° of freedom. Precisely because the granite
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table which acts as the primary datum plane.
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It stops the block from falling through the table along the Z
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axis, and it stops it from pitching and rolling.
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But I can still take my hand and just slide the block around on
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the table in the X&Y directions, right?
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And I can spin it like a top around the Z axis.
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Which means your measurement origin is still floating.
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It's useless to fully constrain the part, you must push it
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against a secondary datum. Like a perfectly perpendicular
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metal fence bolted to the table. Exactly.
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That eliminates the ability to slide in the Y direction and
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stops the spinning. That's two more degrees of
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freedom gone. Finally, you slide the block
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along the fence until it hits a tertiary datum, like a physical
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stop at the end. And that constrains the final X
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axis translation. Right.
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Only when all six degrees of freedom are locked in the datum
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reference frame can you actually establish true position.
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But man, I've review drawings where the engineer referenced
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datums in a way that makes physical measurement literally
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impossible. Oh, I see it all the time.
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They'll call out the center axis of a theoretical hole as the
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primary datum, but you can't physically rest apart on a
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theoretical axis on a granite table.
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No, you cannot touch a theoretical axis, but let's dive
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into the concepts that cause the absolute most heated arguments
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between engineering and manufacturing Maximum material
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condition, or MMC and least material condition LMC.
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Oh boy. Yeah, the material condition
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modifiers. They're the most powerful tools
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in ASMEY 14.5, and honestly the most frequently abused.
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The maximum material condition is the condition where a
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geometric feature contains the absolute maximum amount of
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material permitted by its specified tolerance limits.
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So to make this tangible, if I am designing a steel shaft that
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has to slide into a hole, the MMC of the shaft is when the
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machinist turns it to its largest allowable diameter.
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Correct, it has the most physical metal left on it.
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Conversely, the MMC of the hole it goes into is when the hole is
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drilled at its smallest allowable diameter.
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Yeah, because a smaller hole means less metal was actually
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removed from the block. Yes, and that combination, the
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largest possible shaft attempting to mate with the
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smallest possible hole is your absolute worst case scenario for
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assembly. Because if it fits, then and
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it'll fit always. Exactly.
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If your design works at MMC, you guarantee that any other
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combination of shafts and holes within the tolerance band will
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also fit together seamlessly. And then Least material
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condition or LMC is simply the inverse.
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Right, it is the state where the feature contains the least
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amount of material, so the smallest shaft and the largest
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hole. I see LMC used a lot when
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structural integrity is the primary concern rather than
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assembly. Like if you were drilling a hole
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near the edge of a pressure containing wall.
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Calculate the stress base on the largest possible hole which
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leaves the thinnest possible wall.
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That is the exact application for LMC, yeah, but the true
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genius of why 14.5 is how these material conditions interact
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with positional tolerance. Oh, the bonus tolerance concept,
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Yes. This brings up my absolute
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favorite way to visualize positional tolerance at MMC.
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Think about parking a car in a standard home garage.
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OK, I love this analogy. If you were driving a massive
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lifted heavy duty SUV, that vehicle represents maximum
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material condition. It has the most material because
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it's so large. Your positional tolerance for
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parking it perfectly straight inside that garage is incredibly
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tight. It's almost 0.
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Yeah, it's almost 0. If you veer an inch to the left
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or right you are ripping the side mirrors off on the garage
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door frame. A perfect constraint mechanism.
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But take that same garage instead of the SUV you pull in
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on a tiny dirt bike. The dirt bike represents a
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significant departure from maximum material condition.
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Because it's so much less material.
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Right. And because the bike is so
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small, your positional tolerance for parking it grows massively.
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You can park that by crooked, totally sideways, or shoved
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diagonally in the corner and it will still fit entirely inside
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the strict boundaries of the garage.
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The dynamic nature of that analogy perfectly illustrates
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the concept of bonus tolerance, which is the functional core of
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the MMC modifier. It really is.
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And it explains one of the most terrifying call outs a young
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engineer can put on a blueprint 0 Positional tolerance at MMC.
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Oh man, I've seen machinists completely lose their minds when
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they see a 0 tolerance call out. Because nothing in the physical
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universe is perfect. The laws of thermodynamics
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prohibit it. Oh, I.
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Get so. Angry.
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So they look at the print and say you're demanding infinite
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precision. This is unmanufacturable.
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It causes panic because they're reading it as an absolute zero
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rather than a conditional 0. The blueprint isn't demanding
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perfection under all circumstances.
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It is demanding perfection only when the part is exactly at its
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maximum material condition. OK, let's apply your garage
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analogy to a physical flange. Sure, we have a hole that must
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be located at a true position so a bolt can pass through it.
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The blueprint specifies the hole diameter as 10mm with a
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tolerance of plus 0.5mm -, 0. Therefore, the hole can legally
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range from 10.0 to 10.5mm. Right.
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And the MMC of that hole, the worst case scenario with the
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most metal is exactly 10.0mm. So if we apply a feature control
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frame that calls for a positional tolerance of 0 at
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MMC, we are stating A conditional rule.
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We are saying if and only if the machinist drills that hole at
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exactly 10,000 millimeters, its center must be located with
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absolute perfect 0 deviation precision from the true
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position. Which is physically impossible.
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It is entirely impossible, but here is the brilliant functional
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trick of the standard Machinists know they can't hit perfect
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position, and they know they can't hit exactly 10,000
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millimeters. They don't aim for the absolute
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edge of a tolerance band, they aim for the middle.
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So the machinist uses a slightly larger drill bit and creates a
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hole that is say 10.25mm in diameter.
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And because that hole is now .25mm larger than the MMC limit,
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it has departed from maximum material condition.
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Exactly, and the standard dictates that whatever amount
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you depart from MMC, you get to add that exact amount to your
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positional tolerance as a bonus. Oh wow, so the extra clearance
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created by the larger hole is transferred directly to the
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location tolerance? Yes, suddenly that verifying 0
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positional tolerance has dynamically grown into a .25mm
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tolerance zone. So if the machinist drills the
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hole all the way up to its maximum allowable size of
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10.5mm, the least material condition, they generate a
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massive .5mm bonus tolerance. Right, they can drill the hole
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significantly off center and the part is still legally compliant
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and perfectly functional. Because a larger, slightly
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misaligned hole will still allow the mating bolt to slide through
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just as well as a perfectly positioned small hole.
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The MMC modifier is basically a reward system for the
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manufacturer. It prevents an inspector from
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rigidly measuring a slightly off center hole and scrapping a
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$50,000 titanium component, when in reality the hole was drilled
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large enough that the mating part would have fit perfectly
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anyway. It acknowledges reality over
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theoretical rigidity. But to communicate all of this
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without ambiguity, you have to use the right language.
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You really do. This is where we run into the
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alphabet soup of engineering drawings, which is heavily
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governed by the ASMEY 14.38 standard.
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The Dictionary of drawing abbreviations.
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Exactly this standard is the legal dictionary.
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If you aren't meticulous, casually dropping acronyms onto
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a drawing can cause absolute nightmares.
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ASMEY 14.38 exists because an ambiguous contract is a failed
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contract and the blueprint is the ultimate manufacturing
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contract. Misinterpreting a standard
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abbreviation can completely alter the manufacturing process.
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Can you give an example? Sure.
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For example, in marine engineering FBT stands for
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Forward Ballast tank. In a different context it might
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mean something completely different.
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Or OOR, which is universally out of roundness in pressure vessel
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mechanics. I've actually seen a case where
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someone misinterpreted LMC. We just spent 10 minutes talking
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about least material condition. Oh no.
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What do they think it meant? I saw an assembly procedure
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where a junior tech thought LMC meant lower minimum clearance.
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You're kidding. Nope, they fundamentally flipped
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the math on a critical shaft to bearing fit assuming they needed
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less space instead of calculating for the least amount
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of metal. It ruined the entire sub
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assembly. That is a brutal lesson.
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Misunderstanding a single symbol in a feature control frame or
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mixing up an acronym from Y 14.38 isn't just a type, it is
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the literal difference between a complex mechanism sliding
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beautifully together and creating a very expensive, very
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heavy paperweight. The difference between
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engineering and guessing is precision in communication.
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So assuming your blueprint communication is perfect, your
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GD and T is airtight, your datums are physical and
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accessible, we have to take a step back and ask a more
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fundamental question. Who wrote the physical rules
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that dictate the thicknesses and materials on that blueprint in
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the first place? Right, because you didn't just
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guess that the wall needed to be 15 millimeters thick.
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You certainly shouldn't. Once you know how to dimension
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the part, you must ensure the part possesses the structural
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integrity to survive its operating parameters without
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catastrophic failure or fatigue. This moves us out of drafting
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and into the labyrinth of international design codes for
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heavy mechanical equipment, specifically pressure equipment.
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The two global heavyweights are ASME Section 8 and the European
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Pressure Equipment Directive, commonly known as the PED.
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And here's a huge misconception that I see constantly debated on
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engineering forums. People have this inherent belief
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that math is an absolute universal truth, right?
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They do. They assume physics is physics
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everywhere. Right.
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So if you have a steel cylinder and you put 1000 PSI of pressure
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inside it, all engineering codes around the world should yield
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the exact same answer for how thick that steel needs to be to
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not explode. But they don't.
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The answer is very wildly. They absolutely do, and it is
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often a very bitter pill for engineers transitioning between
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international markets to swallow.
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An engineer might run a calculation using a harmonized
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European standard under the PED and find that a vessel wall
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needs to be 12mm thick. OK, 12mm.
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Then they submit that design to a client in North America, who
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runs it through ASME section, the Vision one, and the code
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kicks it back, mandating a 15 millimeter wall for the exact
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same operating pressure, temperature and fluid.
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So the immediate reaction is well, one of these codes is
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wrong, one is unsafe, or the other is massively over
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engineered. Neither is inherently wrong and
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both are statistically safe. What you are witnessing is a
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difference in the underlying philosophical and legal
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assumptions baked into the mechanics.
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The standards take totally different approaches to material
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variance testing and inherent factors of safety.
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Let's look at the history of those safety factors because it
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explains so much. ASME was born out of the
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horrific boiler explosions of the late 19th and early 20th
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centuries. Steamboats and factories just
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blowing up constantly. Yeah, because metallurgy was so
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inconsistent back then, ASME adopted A heavily conservative
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approach for a long time. The design margin on the
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ultimate tensile strength of the material was 4 point O.
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You had to design the vessel to be 4 times stronger than
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theoretically necessary to account for invisible flaws in
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the steel. Exactly overtime as steelmaking
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improved, ASME reduced that margin to 3.5 in Division One
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and even lower in division two if you utilize design by
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analysis. The European codes are
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different. Very different, developing under
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a different industrial framework, European codes have
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historically relied far more heavily on advanced exhaustive
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material testing upfront at the steel mill.
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Ah, so they catch the flaws before they build the vessel.
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Right, because they demand tighter controls on the chemical
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composition and mechanical testing of the steel before it
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ever leaves the foundry. The PED allows engineers to use
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a slightly lower factor of safety in their calculations.
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The result is a mathematically thinner wall achieving A
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practically identical level of physical safety.
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But regardless of whether you use ASME or the PED, you have to
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deal with how these codes categorize the enemy of all
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materials stress. Yes, stress categorization is
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everything. The days of just running a
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simple wall thickness equation are over.
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When you run a modern finite element analysis in FEA, you get
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a beautiful 3D heat map of your part.
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But I tell people all the time can't just point to a glowing
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red spot on that screen and declare the part fails here.
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The software is stupid. It just shows you raw numbers.
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Exactly. You have to actively categorize
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that stress based on its physical origin.
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If we connect this to the bigger picture, stress categorization
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is the entire foundation of modern design by analysis.
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If you don't understand the origin of the force, you cannot
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predict the failure mechanism. The most critical division in
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engineering mechanics is distinguishing between primary
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stress and secondary stress. OK, let's break that down.
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Primary stress is developed by the imposed loading, which is
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necessary to satisfy the fundamental laws of equilibrium
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of external and internal forces and moments.
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Let's strip away the textbook definition.
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You're talking about the raw, undeniable physical forces
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pushing on the structure. Exactly.
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The most classic example being primary membrane stress.
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If I have a cylindrical pressure vessel and I fill it with high
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pressure gas, that gas is pushing outward equally in all
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directions trying to expand the cylinder.
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The stress is spread evenly across the entire thickness of
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the steel wall. It is uniform and pervasive and
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here is why primary stress is so lethal.
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It is not self limiting. Explain the physics of that.
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What do you mean by not self limiting?
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If the internal pressure in that cylinder creates a primary
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membrane stress that exceeds the yield strength of the steel, the
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steel will cross the elastic limit and begin to stretch
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plastically. Right, it permanently deforms.
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But the high pressure gas inside doesn't care that the steel is
18:18
yielding. The external force remains
18:20
constant, so the steel stretches.
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The wall becomes thinner, which means there is less metal to
18:25
hold the pressure, which causes the stress to increase further.
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Wow, so it's a runaway physical feedback loop that results in
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rapid, catastrophic rupture? It just blows up.
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The force won't redistribute to a stronger part of the vessel,
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it just keeps pushing until it pops.
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That is terrifying. Now contrast that terrifying
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reality with secondary stress. Secondary stress is not caused
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by an external relentless load. It is developed by the
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constraint of adjacent material or by the self constraint of the
18:54
structure itself. So it's fighting itself.
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Exactly. It is almost exclusively caused
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by geometric stiffness, mismatches or severe thermal
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gradients. A thermal gradient is a great
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way to visualize this. Imagine you have a really thick,
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heavy steel pipe in a chemical plant.
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It's sitting at room temperature.
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Suddenly a valve opens and fluid at 600°F violently rushes
19:18
through the inside of the pipe. A massive thermal shock.
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Right. The inner millimeter of that
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steel wall heats up instantly and wants to undergo thermal
19:25
expansion and tries to grow. But the outer layers of the pipe
19:29
wall are still at room temperature.
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They are cold, rigid and massive.
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So the outer wall physically restrains the inner wall from
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expanding. Which creates a massive
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localized compressive thermal stress on the inner wall and a
19:44
tensile stress on the outer wall.
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If you ran that in an FEA, the inner wall will be glowing
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bright red. So bright red.
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The thermal stress might technically exceed the yield
19:52
point of the material. A junior engineer would look at
19:55
that and say the pipe is going to fail, we need to make it
19:57
thicker. Which would actually make the
19:59
thermal gradient worse. Precisely because secondary
20:02
stress behaves differently. If the thermal stress exceeds
20:06
the yield point, the inner wall will plastically deform it.
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It will yield just a tiny microscopic amount.
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But because the force wasn't caused by relentless gas
20:15
pressure, but rather by the metal fighting against its own
20:18
geometry, that tiny bit of yielding actually relieves the
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thermal mismatch. The fight is over.
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Yes, the stress self relieves, local yielding occurs, the
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internal structure of the metal shifts slightly and the massive
20:32
spike in stress dissipates without the pipe rupturing.
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That is wild. This phenomenon is why the ASME
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code allows allowable limits for secondary stresses to be much,
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much higher than primary stresses.
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The code understands that the material will yield slightly,
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undergo what we call shakedown and return to a safe elastic
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state. But, and this is a massive butt
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that destroys pipelines, if that pipe yields back and forth
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repeatedly, we run into the real world failure mechanisms.
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The valve opens, the pipe heats up and yields in compression.
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The valve closes, the pipe cools down, and now it yields in
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tension because it was deformed. Heat up, cool down, expand
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contract. That cycle introduces high
21:08
strain, low cycle fatigue. If you take a metal paper clip
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and bend it slightly, it holds its shape.
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But if you bend it back and forth past its yield point
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repeatedly, the crystalline structure of the metal degrades
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and it snaps. That is, low cycle fatigue and
21:23
secondary stresses are the primary driver of it.
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You also have to design against plastic instability, which is an
21:30
incremental collapse or ratcheting over time with each
21:33
cycle. And if we are dealing with high
21:34
temperatures, you can't just worry about cycles, you have to
21:37
design against creep. Creep is a terrifying mechanism
21:40
because it happens while the equipment is just sitting there
21:42
operating normally. Creep is the insidious, slow,
21:45
continuous plastic deformation of a material under a constant
21:49
stress, specifically at elevated temperatures.
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Even if your primary membrane stress is perfectly safe, well
21:56
below the yield point of the steel, if the metal is hot
21:59
enough, the thermal energy allows the atoms within the
22:02
crystal lattice to slowly slide past one another.
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The dislocations within the metal glide and climb.
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So the vessel literally swells over time.
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Slowly expanding like a balloon over years or decades until the
22:13
walls thin out so much that it triggers a primary stress
22:16
rupture. Furthermore, we face stress
22:19
corrosion cracking, where the combined effects of constant
22:21
tensile stress and a corrosive chemical environment cause
22:24
microscopic cracks to rapidly propagate through the grain
22:27
boundaries of the steel. To put some hard numbers on how
22:30
temperature completely rewrites the rules of these codes, let's
22:34
look at standard flange pressure ratings under ASME B 16 point.
22:38
50 That's a perfect example. The codes dictate what a
22:41
standard pipe flange can legally handle based on its material,
22:44
group and pressure class. For example, if you have a Group
22:48
1.1 material which covers your standard carbon steels like
22:51
SA105A Class 150 flange at ambient room temperature is
22:55
rated to hold roughly 450PSI. OK, 450PSI.
22:59
If you upgrade to a massive class 900 flange of the exact
23:03
same carbon steel, it is rated for 2225 PSI.
23:06
And if you go up to the extreme class 2500, you are looking at
23:09
managing over 9000 PSI. It's incredible, but the
23:13
critical lesson is how drastically those legal ratings
23:16
plummet as the operating temperature increases.
23:19
That 2225 PSI rating for the class 900 flange is only legally
23:25
valid near room temperature. If you run hot oil through that
23:29
system and heat it up to 600°F, the yield strength of the carbon
23:34
steel drops. Creep becomes a mathematical
23:36
factor, and the code aggressively deviates the
23:39
flange. The rating drops by hundreds of
23:41
pounds of pressure just because it got hot.
23:43
The physics dictate the legality.
23:45
But I want to pivot slightly because we keep talking about
23:47
these stresses happening uniformly.
23:49
But vessels aren't just straight pipes.
23:51
The codes make a strict distinction between gross
23:54
structural discontinuities and local structural
23:56
discontinuities. Absolutely do.
23:58
Instead of the explain it like I'm 5 river analogy people
24:01
always use for stress concentrations.
24:02
We are really talking about stiffness mismatches here,
24:05
right? That is exactly the mechanism.
24:07
A discontinuity is anywhere the geometry of the structure
24:11
changes, which inherently changes its localized stiffness.
24:15
Let's look at a gross structural discontinuity.
24:18
The classic example is the circumferential Weld line, where
24:21
the straight cylindrical shell of a vessel meets the curved
24:24
hemispherical head. Under internal pressure, the
24:27
straight cylinder wants to expand radially outward,
24:30
proportional to the pressure in its radius.
24:33
It wants to get fatter. Right, but the hemispherical
24:35
head is a completely different geometric shape.
24:38
Because it's a sphere, it is inherently twice as stiff as a
24:41
cylinder of the same thickness. It resists expanding outward.
24:45
Because they are welded together, they are physically
24:47
forced to share a boundary. The cylinder tries to expand,
24:50
the head refuses to expand as much and they fight each other.
24:53
This creates a massive bending moment right at the Weld joint.
24:57
The membrane theory of uniform stress completely breaks down
25:00
here and bending theory takes over.
25:02
This is a gross discontinuity because it affects a large
25:05
portion of the structure. The entire circumference of the
25:09
vessel is experiencing this bending moment.
25:11
It fundamentally alters the overall stress pattern.
25:14
So then, what constitutes a local structural discontinuity?
25:18
A local discontinuity effects a relatively small volume of
25:21
material and does not drastically change the overall
25:25
global structural behavior of the vessel.
25:27
Imagine a small 2 inch instrumentation nozzle
25:31
penetrating the side of a massive 10 foot diameter tank.
25:34
The general cylindrical shell doesn't even know that small
25:36
hole is there, does it? The overall expansion of the
25:38
pank is unaffected. Exactly.
25:41
But right at the microscopic edge of that drilled hole, the
25:44
stress flow is completely disrupted.
25:46
The stress lines have to detour around the hole, bunching up at
25:49
the edges. Right at the edge of the hole,
25:51
the stress might peak to three times the normal primary
25:53
membrane stress of the surrounding shell.
25:55
Understanding the mathematical difference between gross and
25:58
local discontinuities dictates whether you are designing
26:01
against a global catastrophic collapse of the vessel or just
26:05
localized fatigue cracking at a nozzle corner.
26:08
The transition from the mathematics of the code to the
26:10
physical reality of the steel brings us to the shop slower
26:13
reality check. You applied ASME Section 8.
26:17
You've categorized your primary and secondary stresses.
26:20
You've mapped your discontinuities.
26:22
The design is mathematically bulletproof on your computer.
26:25
But now it hits the fabrication shop.
26:27
And as we established earlier, university taught you that
26:30
materials are perfectly uniform, isotropic and well behaved.
26:34
The shop floor will quickly teach you that thousands of tons
26:37
of steel will warp, bend and aggressively fight back.
26:40
The transition from paper to steel is rarely graceful.
26:43
We primarily see two manufacturing methods for heavy
26:46
pressure equipment, fabrication and forging.
26:49
Forging or extrusion is where seamless pipe work is literally
26:52
pulled or pressed out of a solid billet of red hot steel.
26:56
It is structurally superior because there are no
26:58
longitudinal Weld seams. But forging is severely limited
27:02
by size. You cannot forge a distillation
27:05
column that is 20 feet in diameter and 100 feet tall.
27:08
No, No machine on earth is big enough for that.
27:11
To build anything massive, you must rely on fabrication.
27:15
You have to take flat steel plates, push them through
27:18
massive hydraulic rollers to bend them into cylinders and
27:21
Weld the scenes shut. And this is the biggest pain
27:24
point in the industry, dealing with plate thickness during the
27:27
rolling process. Plate rolling is an art form
27:30
masquerading as industrial manufacturing.
27:32
The specific thickness of the plate dictates the exact
27:35
nightmare the fabricator will face.
27:37
Let's look at the extremes. If you are dealing with thin
27:40
plate, say a massive diameter vessel, but the pressure is low
27:44
so the wall is only 8mm thick. Your primary enemy is a total
27:48
lack of rigidity. Heat distortion will ruin you.
27:51
When you bring the two edges of that rolled cylinder together
27:54
and strike a welding arc to close the seam, you are pumping
27:56
an immense amount of localized heat into the steel.
27:59
The molten Weld pool expands rapidly, pushing the parent
28:02
metal away. But as the Weld cools, it
28:05
shrinks with incredible force. Because the 8mm plate lacks the
28:09
physical mass and structural rigidity to resist those thermal
28:12
contraction forces, the entire cylinder will warp.
28:15
It will suck inward at the Weld seam, creating a localized flat
28:19
spot or a sharp peak along the joint.
28:22
So the immediate engineering thought is thin plate is floppy
28:25
and warps under welding heat. Let's just specify thick plate.
28:28
Let's build it out of 30mm thick steel.
28:32
But thick plate introduces the exact opposite metallurgical
28:35
problem. A 30mm thick steel plate is
28:38
incredibly stiff. When you force it into the pinch
28:41
rollers and apply hundreds of tons of hydraulic pressure to
28:44
curve it, you are forcing the outer fibers of the steel to
28:47
yield in tension while the inner fibers yield in compression.
28:51
And the steel hates that. It fights back constantly.
28:54
When the curve plate finally emerges from the rollers and the
28:56
hydraulic pressure is released, it experiences spring back.
29:00
A massive amount of elastic energy is stored in the core of
29:03
the plate that didn't yield perfectly, causing the curve to
29:06
and curl slightly. Anticipating the exact
29:08
mathematical amount of spring back to hit a precise 10 foot
29:11
diameter is notoriously difficult.
29:14
Furthermore, heavily working thick plates cold causes work
29:18
hardening which makes it even stiffer.
29:20
Often rolling thick plates requires the material to be
29:23
heated red hot to drop its yield strength, which introduces
29:26
complex metallurgical phase changes and dimensional scaling
29:30
that totally ruin your tolerances.
29:32
So the metal comes out of the rollers.
29:34
It's never perfectly round. They tack Weld it up, make the
29:37
final longitudinal Weld, and then the mandatory visual and
29:41
dimensional examinations happen. The third party inspectors walk
29:45
out onto the shop floor with their clipboards and
29:47
flashlights. What are they actively hunting
29:49
for? They're hunting for the physical
29:51
evidence of the fabricator struggle.
29:53
They look for arc strikes outside the Weld zone which
29:56
create localized brittle spots. They look for embedded mill
29:59
scale that wasn't shot blasted off which will contaminate the
30:02
Weld chemistry. But most critically, they are
30:04
looking for bulging and peaking. Explain bulging because I've
30:08
seen vessels that look like the steel is literally pregnant.
30:11
That is the perfect visual. Remember the gross structural
30:14
discontinuity? We just analyzed the head to
30:17
shell junction. Think about the physical reality
30:20
of assembling that. The fabricator rolled a massive
30:23
cylinder and despite their best efforts, it's slightly Oval.
30:27
It is not perfectly round, and the dished head was pressed in a
30:31
massive dye. And it's not perfectly round
30:33
either. So when the crane lowers the
30:34
head onto the shell, they don't slide together perfectly.
30:37
There are massive gaps, a severe mismatch.
30:40
To close those gaps and get the metal edges touching so they can
30:43
be welded, fabricators will Weld heavy steel brackets called dogs
30:47
to the shell and the head. They then use massive hydraulic
30:51
jacks, steel wedges, and turnbuckles to physically force
30:54
twist and crush the shell until it aligns with the head.
30:57
They force it into submission and heavily tack Weld it while
31:00
it is under immense screaming mechanical stress.
31:03
They are essentially locking in thousands of pounds of residual
31:06
stress before they even strike the arc for the final structural
31:09
Weld. Exactly, and when they finally
31:11
perform that heavy multi pass Weld, the localized area where
31:15
they force the metal together will buckle.
31:17
It will bulge outward or suck violently inward, leaving a
31:20
highly visible permanent distortion.
31:23
A bulge is not just a cosmetic flaw, it is the physical
31:26
manifestation of severe, locked in residual stress that will
31:30
aggressively lower the fatigue life of the vessel.
31:33
And what happens when the fabricator knows the inspector
31:36
is coming to look for that mismatch?
31:38
If the Weld seam looks ugly and bumpy, the fabricator's primal
31:42
instinct is to grab an angle grinder and make it look
31:44
perfectly smooth. Right.
31:45
That is one of the most insidious dangers in pressure
31:48
vessel manufacturing. Fabricators will sometimes
31:50
heavily grind a poorly aligned, bumpy or peaked seam.
31:54
They blend it out over several inches so it looks completely
31:57
smooth and flushed to the naked eye.
31:59
The visual inspector might walk by, shine a flashlight on it and
32:02
say that is a beautiful, seamless transition.
32:05
But the grinder didn't add metal to fix the low spots, it removed
32:08
base metal from the high spots. Precisely by aggressively
32:13
grinding the parent metal to hide the mismatch, they have
32:15
actively reduced the effective wall thickness of the vessel
32:19
below the legal code allowances we just spent hours calculating
32:22
in the engineering office. They've traded a visible
32:25
geometric flaw for a hidden, catastrophic lack of structural
32:29
material. This is why a simple visual
32:31
check is never enough. The ASME code mandates rigorous
32:35
dimensional checks on five specific tolerance features of a
32:39
constructed vessel. Surface alignment to ensure the
32:41
plates meet flush without excessive offset.
32:44
Straightness of the overall shell circumference to ensure
32:47
the volume is correct. The general geometric profile
32:50
and critically, circularity, which the industry heavily
32:53
refers to as out of roundness or OOR.
32:57
Out of roundness is a critical mathematical parameter.
32:59
If a cylinder is perfectly mathematically round, internal
33:02
pressure creates pure uniform membrane stress in the wall.
33:06
But if the cylinder is slightly Oval shaped, meaning it is out
33:09
of round, the physics change. When you pressurize an Oval, the
33:12
internal force actively tries to push it back into a perfect
33:15
circle. Because a circle is the most
33:16
volume efficient shape for a given perimeter.
33:19
Yes, but as the internal pressure pushes the Oval walls
33:22
outward to make them circular, that physical pushing creates
33:25
severe bending moments within the wall of the vessel.
33:28
It acts like a beam bending rather than a balloon
33:31
stretching. And as we discussed with
33:33
secondary stresses, bending stresses are highly localized
33:36
and much more damaging. If the vessel is heavily out of
33:38
round and the pressure cycles up and down, those bending walls
33:41
will rapidly succumb to fatigue failure.
33:44
The code strictly limits OOR to typically 1% of the nominal
33:48
diameter to prevent this. So here is a practical reality
33:51
check. If I have a giant distillation
33:53
column sitting horizontally on roller beds on the shop floor,
33:57
50 feet long and 12 feet in diameter, how do I actually
34:00
measure its straightness or its datums?
34:03
I can't just hook a flimsy tape measure to the jagged edge of a
34:05
50 foot pipe. You absolutely cannot.
34:08
The physical edges are often beveled at complex angles for
34:10
future field welding, or they were just rough cut and
34:12
irregular. Furthermore, you can never use
34:15
the longitudinal or circumferential Weld seams as
34:17
your measurement datums. Welds are inherently lumpy,
34:20
distorted by heat, and inconsistent.
34:23
So how do you establish A coordinate system on a giant
34:26
featureless steel tube? To establish a true transverse
34:30
datum plane, the inspectors use deep center punch marks located
34:33
50 to 100mm inward from the final Weld seam directly on the
34:37
smooth, unaffected parent metal. They strayed A precise,
34:41
perfectly circumferential line all the way around the vessel
34:44
using a laser level or a wrap around band.
34:47
That punch line becomes the rigid physical coordinate
34:49
system. Every subsequent measurement,
34:51
nozzle elevations, support saddle locations.
34:54
Overall straightness is pulled strictly from those punch marks,
34:57
completely ignoring the irregularly physical ends of the
35:00
cylinder. So let's assume the fabricator
35:02
did a stellar job. They control the heat, they
35:04
didn't over grind the vessels perfectly round, and the visual
35:07
inspector signs off on the dimensions.
35:10
The welder swears on his life that the Weld is a perfect full
35:12
penetration joint, but we still don't truly know what is
35:15
happening inside the microscopic atomic structure of that steel.
35:18
The welder says it's perfect. The grinder made it shiny.
35:21
How do we actually know it won't unzip and explode when we fill
35:24
it with 2000 PSI of lethal gas? We have to trust but verify.
35:29
Which brings us to the most complex, legally fraught and
35:32
heavily debated aspect of pressure equipment construction,
35:34
not non destructive examination or NDE.
35:38
Governed almost entirely by ASME section V.
35:40
This is the rule book for how we look inside solid steel without
35:43
breaking it. Open Radiography, which is
35:45
X-rays. Ultrasonic testing, magnetic
35:47
particle liquid penetrant. But the actual physics of these
35:50
tests are only half the story. The history and the philosophy
35:53
of how we apply them is fascinating.
35:55
I was reading through the historical transcripts of the
35:57
ASME Subcommittee V 50th Anniversary panel discussion.
36:01
Hearing guys like EC Miller and Doctor Gerald Tenney, The
36:03
Pioneers of industrial radiography, argue back in 1970
36:06
give you such an immense appreciation for how hard fought
36:09
these regulations were. The 1970 panel is required
36:12
reading for anyone in quality control.
36:14
The core philosophy they fought to establish, and which still
36:17
dictates global NDE today, is a strict separation of powers.
36:21
ASME Section V provides the fundamental methods, but the
36:25
manufacturer must develop the specific written procedures.
36:29
Section V acts like a physics textbook.
36:32
It tells you how gamma radiation attenuates as it passes through
36:35
carbon steel in a broad, generic sense.
36:37
It gives you the theory. Exactly.
36:39
But theory doesn't inspect welds.
36:41
The manufacturer's engineering team must write a rigid
36:43
procedure that says for this specific 15 millimeter thick SA
36:47
516 grade 70 carbon steel Weld using this specific Iridium 192
36:52
isotope at a source to film distance of exactly 24 inches,
36:55
exposed for precisely 4.5 minutes.
36:57
Here are the exact steps our technician will follow to prove
37:00
the Weld is sound. Improving the film is sound
37:03
brings up the biggest technical debate in the history of
37:05
radiography. The whole size versus plaque
37:08
thickness debate regarding radio graphic sensitivity.
37:12
OK, I need to break down the physical mechanics of this for
37:14
everyone so they understand why this is such a massive industry
37:17
pain point. When a technician takes an X-ray
37:20
of a Weld, they don't just shoot the Weld blindly.
37:23
They place a tiny precision machine piece of metal right
37:26
next to the Weld on the surface of the steel.
37:29
This is called a penetrameter or an image quality indicator, an
37:32
IQI. The IQI is the absolute
37:35
foundation of radio graphic trust.
37:37
This little metal plaque has tiny microscopic holes drilled
37:40
completely through it. When you develop the X-ray film,
37:43
that film was clearly show the image of those tiny holes.
37:46
The logic is simple. If the X-ray technique wasn't
37:48
sharp enough or sensitive enough to see the artificial, perfectly
37:51
drilled hole in the penetrameter, then it definitely
37:53
lacks the resolution to see a jagged, microscopic real world
37:57
crack hidden deep inside the Weld.
37:59
The panettremeter proves the baseline quality of the
38:01
radiograph. Correct.
38:02
If the inspector cannot clearly see the required hole on the
38:05
film, the entire radiograph is legally void.
38:08
It must be shot again. So here is where the
38:10
misconception occurs and it causes massive fights between
38:14
clients and fabricators. A client looks at a
38:16
specification and says we want to ensure absolute safety, we
38:19
want a tighter spec. Let's demand the fabricator use
38:22
a penetrometer with a much smaller hole.
38:24
So instead of the standard .062 inch diameter hole, the client
38:28
demands of the technician must prove they can see a tiny .040
38:32
inch hole. And the client smugly believes
38:34
they have just doubled the rigor of the inspection.
38:36
But from a purely logical optical standpoint, if I force
38:39
the NDE technician to find a .040 inch hole instead of a .062
38:44
inch hole, I am proving the film is sharper, right?
38:46
I am demanding high resolution of a smaller detail.
38:49
You are demanding higher resolution of A2 dimensional
38:51
plane, but radiography is not a photograph, it is a volumetric
38:56
density measurement. Shrinking the diameter of the
38:59
penetrometer hole does not automatically mean the
39:01
specification is tighter if you allow the thickness of the
39:04
penetrometer plaque to change. And this is exactly the
39:08
mathematical trap EC Miller was aggressively warning about in
39:11
1970. Explain the trap.
39:14
Why does the thickness of the little metal plaque matter if
39:17
I'm just looking at the hole diameter?
39:19
Because X-rays don't see holes, they see a difference in the
39:21
density of mass. When radiation passes through
39:24
solid steel, a certain amount is absorbed.
39:27
When it passes through the air gap of the drilled hole, less is
39:30
absorbed, so that spot on the film becomes darker.
39:33
It creates radio graphic contrast.
39:35
The contrast is entirely dictated by the volume of the
39:38
missing metal. The volume of the cylinder of
39:40
air created by the hole precisely.
39:42
If a client demands a smaller .040 inch hole but allows the
39:47
technician to use a pentetrometer plaque that is
39:49
twice as thick, say jumping from .0075 inches thick to .010
39:55
inches thick, the physical volume of missing metal might
39:58
actually be greater than the original wider hole in the thin
40:01
plaque. A deep, narrow hole in the thick
40:04
piece of metal creates a much darker, much more obvious shadow
40:07
on the X-ray film and a wide shallow hole in a paper thin
40:10
piece of metal. So by demanding a smaller hole
40:13
diameter but ignoring the plaque thickness, the client might
40:16
actually be making the test easier to pass.
40:19
The deep hole shows up bright as day on the film, the technician
40:22
passes the test easily, and the client feels falsely secure,
40:25
claiming they tighten the spec. It is a fundamental
40:27
misunderstanding of physics that catches novice quality managers
40:30
all the time. Which is why you must understand
40:32
the geometry of the pentetrometer as a whole 3
40:34
dimensional system. And this leads directly into the
40:36
broader engineering approach to NDE, which was another massive
40:39
topic in that 1970 panel heavily championed by Mr. Jackson.
40:43
The engineering approach is basically the antidote to over
40:46
specifying. Jackson's argument was that we
40:49
don't build every single pressure vessel out of exotic
40:51
aerospace grade in carinal. If a vessel is just holding
40:55
compressed air for shop tools, we build it out of standard
40:57
carbon steel because carbon steel is perfectly adequate for
41:00
that service. And Jackson argued that the
41:02
exact same logic must apply to inspection.
41:05
We should not mandate the absolute highest, most brutally
41:08
expensive level of 100% volumetric radiography for every
41:12
single Weld we make. It is not economically viable
41:15
and it is not structurally necessary.
41:17
Level of NDE must match the severity of the design service.
41:21
If you were designing A vessel that holds highly pressurized
41:23
lethal hydrogen sulfide gas in the middle of a populated city
41:26
center, yes, that vessel gets 100% radiography, 100%
41:31
ultrasonic testing, and magnetic particle on every single nozzle.
41:35
You spare no expense. But if you are building a simple
41:37
low pressure water storage tank, sitting in the middle of an
41:40
empty field, mandating 100% radiography is a profound waste
41:45
of resources. A visual exam, a hydrostatic
41:47
pressure test, and perhaps some localized dye penetrant are more
41:51
than sufficient to guarantee safety.
41:53
NDE must be a calculated engineering decision based on
41:56
risk, not just a blind blanket adherence to the heightest
42:00
possible specification to satisfy a lawyer.
42:02
But. Relying on human beings to make
42:04
those calculated decisions raises an incredibly spicy
42:07
question, a fundamental tension in the industry that exists
42:10
today just as much as it did 50 years ago.
42:13
It's the inherent conflict between the inspector and the
42:15
examiner, the men and women sitting in a dark room staring
42:18
at an X-ray film on a light box, deciding if a faint Gray
42:21
squiggle is an acceptable scratch on the film or a crack
42:24
that requires them to scrap $1,000,000 vessel.
42:26
We have to talk about the massive debate regarding Level 3
42:29
personnel qualifications. This is the ethical minefield of
42:32
non destructive testing. To understand the tension, you
42:35
have to understand the hierarchy.
42:37
In the NDE world, a Level 3 examiner is the absolute top of
42:41
the food chain. They're the masters of the
42:43
physics. They write the complex testing
42:45
procedures, they train and certify the Level 2 technicians
42:48
who shoot the film, and crucially, they make the final
42:51
binding call on highly complex flaw interpretations.
42:54
But here's the rub. Outside a very specific, tightly
42:58
controlled military, nuclear programs or certain aerospace
43:01
sectors, Level 3 examiners are not independent government
43:05
agents. They are certified and employed
43:07
by the manufacturer. The very company that built the
43:09
vessel. The company that stands to lose
43:11
millions of dollars, face massive schedule delays and
43:14
potentially incur liquidated damages.
43:17
If the Level 3 examiner looks at an X-ray and says this massive 5
43:20
inch thick Weld is cracked, you have to spend the next two weeks
43:23
gouging it out and welding it again.
43:25
You immediately see the inherent tension in the quality control
43:28
system. You have the authorized
43:30
inspector or AI. The AI is the independent third
43:34
party representative, usually employed by an insurance agency
43:37
or a government jurisdiction. The AI is there to verify that
43:41
the ASME code was strictly followed, but the code itself
43:45
often dictates that the acceptability of an imperfection
43:48
is the sole decision of the examiner.
43:50
And the Examiner draws their paycheck, their bonuses, and
43:53
their career advancement from the company that desperately
43:55
needs that vessel to ship by Friday.
43:58
It is the literal definition of the fox guarding the hen house.
44:02
Mr. Snyder, one of the ASME panelists, explicitly brought
44:05
this up in 1970. He noted the difficulty of
44:08
retaining objectivity when the code delegates the final call of
44:11
acceptable flaws to the manufacturer's payroll.
44:13
Who holds the ultimate moral and legal responsibility when an
44:16
examiner passes a borderline flaw that later propagates into
44:20
a catastrophic failure? It is a system that realized
44:22
perilously heavily on the ethics and unwavering professionalism
44:26
of the NDE community. And to their immense credit,
44:29
historically that community is incredibly strong and resistant
44:32
to pressure. But the systemic structural
44:35
tension is undeniably real. A production manager screaming
44:39
at a Level 3 to pass a Weld so they can hit a quarterly quota
44:42
happens. The authorized inspector can
44:44
review the radiographs, and if they spot something egregiously
44:47
obvious, they can challenge the Level 3.
44:50
But AIS are generalists. They are relying on the
44:53
manufacturer's specialized Level 3 to interpret the murky,
44:56
subjective Gray areas of the film.
44:58
It is an entire global safety system built on assumed trust,
45:02
verified by audit trails, but fundamentally reliant on the
45:05
unshakable integrity of a single individual looking at a film in
45:09
a dark trailer at 2:00 AM. Which is why the engineering
45:12
blueprint and the design code must be conservative enough to
45:15
survive the occasional human error on the shop floor.
45:18
OK, so let's assume the system works perfectly.
45:20
The vessel is designed mathematically.
45:22
Florida State. It is fabricated without
45:24
distortion. The NDE Level 3 gave the welds a
45:27
completely clean, ethical bill of health.
45:29
You are popping the champagne in the engineering office.
45:32
The project is a success. But it isn't.
45:34
Not yet. No, because now you were at
45:36
Section 5 getting it there. The brutal, unforgiving physics
45:40
of transportation and erection. You have a 500 ton piece of
45:43
solid steel sitting horizontally in a shop in Ohio, and it needs
45:48
to be installed vertically on a concrete foundation at a
45:51
refinery in Saudi Arabia. This is the great blind spot of
45:54
mechanical engineering. This is where so many flawless
45:56
designs meet their absolute violent doom.
46:00
It is staggering how often transportation is treated as an
46:03
afterthought. We spend thousands of computer
46:05
hours analyzing the operational stresses of hot fluid flowing
46:09
through the vessel, and we entirely forget that putting a
46:11
massive, rigid steel tower on a flat bottom barge and sending it
46:15
across the open ocean subjects it to dynamic, chaotic forces it
46:19
will never see in normal operation.
46:21
Let's breakdown the sheer brutality of barge shipping
46:23
forces. When that vessel is strapped
46:26
horizontally to the steel deck of a barge, gravity is pulling
46:30
it straight down at 1.0 WW, being the total shipping weight
46:33
of the vessel. If the ocean was a perfectly
46:36
flat stationary table, that would be the end of the
46:39
calculation. But the ocean is a dynamic,
46:41
fluid medium we must design for the marine forces of heave,
46:45
pitch, roll and collision. Let's start with heave, which is
46:49
the vertical rising and falling of the entire ship as it rides
46:52
over ocean swells. No wait, why is heave so
46:55
destructive? If the boat goes up and down,
46:57
the vessel just goes up and down with.
46:59
It it is about acceleration and apparent gravity.
47:02
When the barge rides up to the Crest of a massive wave and then
47:05
drops into the deep through, it is accelerating downward.
47:08
But when it hits the bottom of the through, it suddenly
47:10
decelerates as the buoyancy of the hull fights against the
47:13
water. This massive deceleration causes
47:16
the apparent gravity acting on the cargo to spike violently.
47:19
Furthermore, you have the hydrodynamic added Mass Effect.
47:22
The barge has to push thousands of tons of water out of its way
47:25
to move. So the vessel suddenly feels
47:26
much heavier. Vastly heavier.
47:29
The classic marine design codes require you to calculate the
47:32
base tie downs for a heat force of 1.2 times the total weight of
47:35
the vessel. For a brief moment in that wave
47:38
through, your 500 ton vessel effectively weighs 600 tons,
47:42
attempting to crush the steel saddles at rest Sun.
47:45
And then you have roll the boat tipping side to side as waves
47:48
hit a broadside. The standard lateral force
47:50
calculation for roll is ±0.45 W. Think about the magnitude of
47:56
that number. Nearly half the entire dead
47:58
weight of the vessel, 225 tons of force, is actively trying to
48:02
shear the tie down, bolt sideways and slide the vessel
48:05
straight off the deck into the ocean.
48:07
And if it doesn't slide off, you have to worry about collisions
48:09
or hard dockings when the barge hits the pier or tugboats
48:12
shunted aggressively. The longitudinal force during a
48:15
collision scenario is calculated at ±1.5 WA.
48:18
Force equal to 1 1/2 times the vessel's weight. 750 tons of
48:23
kinetic energy is trying to launch the vessel forward like a
48:26
massive steel missile. If the transport saddles and the
48:29
base plates welded to the vessel aren't specifically designed to
48:32
absorb a 750 ton horizontal shear load, the tie downs will
48:36
shear, the vessel will RIP its own base plates off, destroy the
48:39
deck, and sink the barge. The vessel must be designed as a
48:43
structural chassis first and a pressure container second.
48:46
Here's where it gets really interesting.
48:48
Assuming the tie downs hold, it survives the violent boat ride.
48:52
It arrives at the site in Saudi Arabia.
48:54
Now you have to pick it up. You have to take this 500 ton
48:57
cylinder from a horizontal resting position to a perfectly
49:00
vertical standing position. This requires massive cranes and
49:04
more importantly, it requires lifting lugs.
49:06
A top flange lifting lug is essentially a massive steel ear
49:09
welded to the very top edge of the vessel.
49:11
The crane drops its massive hook and cables down pins into this
49:15
lug and begins to hoist. I was looking at the math for
49:17
designing these lugs and the force vectors are incredibly
49:20
complex because as the crane pulls up, the vessel starts to
49:23
tilt. The angle of the lifting cable
49:25
is constantly changing relative to the lug as the vessel rotates
49:28
from 0° horizontal to 90° vertical.
49:32
It is a dynamic stress analysis. You must calculate transverse
49:36
shear, the lateral force from the angled cables actively
49:39
trying to slice the lug off sideways.
49:41
You must calculate longitudinal shear, the raw vertical force
49:45
trying to RIP the lug straight up out of the Weld.
49:48
And most critically, you must calculate the massive bending
49:51
stress. As the crane pulls at a severe
49:54
angle, the cable acts like a giant mechanical lever, applying
49:57
A bending moment that tries to snap the lug off the flange like
50:00
a brittle twig. And.
50:01
There's a massive real world trap here that literally kills
50:04
riggers. It has nothing to do with the
50:06
thickness of the lug, it is entirely to do with the empty
50:09
space inside it. The pin diameter sizing.
50:11
You have a massive hole drilled through the lifting lug.
50:14
The crane drops a heavy steel pin through that hole, connects
50:17
the shackles and lifts. What is the fundamental non
50:20
negotiable rule regarding that pin?
50:22
The rule dictates that the outer diameter of the lifting pin can
50:25
be no less than 116 of an inch smaller than the inner diameter
50:29
of the hole. So if the hole in the lug is
50:31
exactly 4 inches in diameter, the steel pin sliding through it
50:34
must be at least three and 151516 inches in diameter on
50:38
your perfect fit. Why is the tolerance so absurdly
50:41
tight? What if the rigger is out in the
50:43
desert, he looks in the rigging box and he only has a 2 inch pin
50:46
handy. A solid 2 inch pin made of
50:49
forged high strength alloy steel can easily hold 500 tons without
50:53
shearing an F. Why can't he just drop a
50:55
perfectly strong but smaller pin into the big 4 inch hole?
50:59
Because of Hertzian contact stress, more commonly known in
51:02
this application as bearing stress, if you put a 2 inch pin
51:05
inside a four inch hole, the radius of the pin does not match
51:08
the radius of the hole. Therefore, the pin does not sit
51:11
flush against the curve of the metal.
51:13
It is resting on a tiny microscopic sliver of an arc at
51:16
the very top, dead center of the hole.
51:18
The contact area between the pin and the lug isn't a broad curve,
51:22
it is virtually a single infinitely thin geometric line.
51:26
Exactly. And let's look at the basic
51:28
physics equation for stress. Stress equals force divided by
51:31
area. We have a massive force, 500
51:34
tons of lifting weight, but because the pin doesn't fit, the
51:38
area of contact is approaching 0.
51:41
And as the denominator in that equation shrinks towards 0, the
51:44
resulting stress skyrockets toward Infinity.
51:47
Because the contact area is so incredibly small, when you apply
51:50
500 tons of lifting force to that single mathematical line of
51:53
contact, the bearing stress concentration becomes
51:56
astronomical. It massively instantly exceeds
51:59
the ultimate tensile strength of the lug material.
52:02
Even if the 2 inch pin itself is strong enough, if not to break,
52:05
it destroys the vessel. The small pin will literally act
52:08
like a hot wire through butter. It will yield the steel, slice
52:11
straight up through the heavy steel lug, RIP it wide open,
52:15
drop the 500 ton vessel to the concrete and kill anyone
52:18
standing underneath it. That is horrifying.
52:20
By mathematically mandating that the pin must be within 116 of an
52:24
inch of the hole diameter, you ensure that the curvature of the
52:27
pin almost perfectly matches the curvature of the hole.
52:30
This ensures the pins completely flush, spreading that 500 tons
52:34
of force over a massive surface area of steel, driving the
52:38
denominator up and keeping the localized bearing stress at a
52:41
safe elastic level. That blows my mind.
52:44
Not because the math is hard, but because of the vulnerability
52:47
it exposes. Think about the entire journey
52:49
we just took. This vessel was engineered using
52:52
advanced finite element analysis to hold thousands of pounds of
52:56
explosive high temperature pressure.
52:58
It's walls are inches thick, forged from certified alloys.
53:01
The welds were debated by Level 3 experts and X-ray to atomic
53:05
perfection. The GD and T was perfect.
53:07
And yet it could be utterly destroyed, bankrupt the company,
53:11
and take human lives just because a boat swayed
53:14
aggressively in a wave through, or because a tired rigor on a
53:17
Tuesday morning decided to use a lifting pin that was an inch too
53:20
small. It highlights exactly why
53:23
mechanical engineering requires holistic cradle to grave
53:26
thinking. You aren't just designing a
53:29
static pressure container in a vacuum.
53:31
You were designing an object that must survive the human,
53:34
physical and environmental chaos of the real world.
53:37
Which is why we also have to obsess over the trunnions.
53:40
Trunnions are the massive heavy walled pipes welded to the side
53:43
of the vessel lower down near the center of gravity.
53:46
We use them as tailing devices during the lift when the main
53:49
crane lifts the top flange. A secondary tailing crane holds
53:53
the trunyons to keep the bottom of the vessel from dragging on
53:56
the ground as it swings. And the forces there are just as
53:59
complex as the top look. Even more so when the vessel is
54:02
horizontal. The tailing crane lifting on
54:05
those side trunyons creates a massive torsional moment, a
54:09
twisting force that tries to ring the cylindrical shell like
54:12
a wet towel. As the vessel rotates and
54:15
becomes vertical, the force vector shifts.
54:18
It creates a severe longitudinal bending moment, acting like a
54:22
lever trying to RIP a gaping hole in the side of the vessel
54:25
wall. Which means the thin shell of
54:27
the vessel can't handle it alone.
54:28
Rarely, accommodating those temporary lifting loads requires
54:32
the engineering team to design massive external reinforcing
54:35
pads welded around the trunnions and often massive internal
54:40
structural steel rings inside the vessel just to prevent the
54:43
shell from buckling. We undergo meticulous Weld
54:46
analysis, fatigue calculations, and hundreds of thousands of
54:49
dollars in added material and fabrication time.
54:51
All for a temporary lifting operation that lasts, what, 20
54:54
minutes? Usually less. 20 minutes of
54:56
lifting literally dictates the structural architecture of the
54:59
entire internal vessel. Which leaves us with a
55:01
fascinating paradigm shift, a final provocative thought to
55:05
Mull over as we wrap up this master class.
55:07
What if the entire industry has this backward?
55:10
What have we stopped thinking of shipping barge dynamics and
55:13
crane erection as secondary transportation problems?
55:16
Yeah, that we just tack on to the end of the design process.
55:19
It is a vital question. What if from the very first day
55:23
of drawing the CAD model, we treated the violent boat ride
55:26
and the 500 ton crane lift as the vessel's primary, most
55:29
volatile operating environment? How would that paradigm shift
55:32
fundamentally change the way we design our base plates, our
55:35
internal skirts, and our transport saddles from day one?
55:38
If we prioritize engineering for the brutal journey first, does
55:41
the static destination simply take care of itself?
55:44
School taught you engineering, The shop floor will teach you
55:47
wisdom. Don't forget like follow any
55:48
act. If this helped you even a
55:50
little, please tell us. We're putting this out for free.
55:53
No paywall, no subscription, nothing locked behind a card
55:56
that only works if we know it's actually reaching and teaching.
55:59
If you're on a journey trying to figure this stuff out and this
56:02
gave you clarity you weren't getting anywhere else, drop a
56:05
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56:08
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56:10
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