I was just contemplating things we take for granted, or gospel, or maybe even just plain old "well it's always been done that way", and figured I should share some engineering experience with breaking things for a living that I used to do. And how that relates to the topic line.
Since I joined about a year ago, I have seen some discussions relating to how things go together, and how they may break. I've got a few observations regarding a lot of things we use that is beyond the intended design. For example: old Ford tie rod ends, the 11/16 threaded ones, that are used for split wishbones. The same logic applies to spherical rod ends.
Consider this: in the original application, tie rod ends and spherical rod ends (Heims for ease of typing), were designed for push-pull situations. Yes the stud on a tie rod end is in bending as well as shear, but that is a good safety valve to let the stud bend if the tire hit a pot hole or kerb. The body of the tie rod end was not intended to take bending loads however. Actual Ford tie rod ends were made from good steel alloys, but that doesn't make that much difference when applying a bending load. So we take the wishbones or hairpins and what happens?
The threaded part ends up in bending as well as taking end loads and shear loads. The same is true for Heim joints: they were invented by the Germans for Aircraft use to control moving wing surfaces, allowing some rotation on the stud, whereas the clevis that formerly was used tended to bind on the pivot bolts.
We found out about them from a recovered German fighter plane that the English were able to dissassemble and analyze. The Rose company in England started making them for the British aircraft industry, and they shared the design with us during the runup to WW2. By the way clevises fall in this same category, not originally intended to absorb bending loads.
So you say "Dave G, what's yer point?" Just that we should thank the engineer who specced the originals to have a decent safety factor for the intended usage. I tend to go over sizewhen using these, like using the Ford truck 3/4-16 thread tie rod end, part number something like ES150 I think. The stud is larger, and the threaded body is much larger. Meaning when I use it in an application it wasn't designed for, that there will be a little more margin of safety before failure.
As an aside, the place to keep an eye on for fatigue cracks to start is at the point where the threads just exit whatever it is threaded into. This is the junction point between the bung welded into a split bone for instance and a locknut if used, or right at the end of the threaded bung where the male thread exits the female thread. The root of threads are the highest stressed part of a fastener, and in bending is a good (well bad for us!) place for a crack to start and propagate over time. And since we tend to use a lot of the vehicles we build with these adaptations in a spirited manner and seasonal, there is also the potential for stress corrosion cracking, which can lower the failure stress limit to a fraction of the ultimate strength of the part.
Concerning bolts thru Heims, double shear is preferable if you can design bracketry to accomodate it. Single shear puts the boltin both shear and bending, whereas double shear is pure shear. Shear strength is typically 1/2 of ultimate strength, so a bolt with a 90,000 psi ultimate strength can handle approximately 45,000 psi in shear. To determine the actual load, or force it can handle, we need to know the actual cross sectional area of the bolt where its in shear. So if part of the threads are in the shear zone use the root diameter for calculating the stressed area, otherwise its the area of the bolt body.
An example of the area of a bolt is a 1/2 inch bolt has a cross sectional area of 0.196 square inches, and assuming a bolt ultimate strenght of say 180,000 psi ( a really good bolt) the shear strength is about 90,000 psi, resulting in a load of about 17,670 pounds in shear per shear zone, or for double shear, approximately 35,000 pounds of load. But I'm going to throw a monkey wrench in the works.
Most materials are load rate sensitive, meaning under impact loading they fail at lower stress levels. In Applied Mechanics terms this is called dynamic fracture toughness. Also, many materials (steel can be one) also have a transition temperature, where above that temperature failure tends to be ductile, and below failure tends to be brittle. Anyone who rode or raced snowmobiles back in the late 60s-early 70s may have experienced this, with odd things breaking unexpectedly, like springs, motor mount bolts.
I could go on forever, but I won't. Ask Dan, I used to have diarrhea of the brain frequently when we were bench racing... It comes from having the "Knack"
So please ask me questions if interested, as a retired mechanical engineer with both a graduate degree in Applied Mechanics and a lot of experience in Failure, it helps keep my mind active.
Thanks for taking this for what its worth...
Since I joined about a year ago, I have seen some discussions relating to how things go together, and how they may break. I've got a few observations regarding a lot of things we use that is beyond the intended design. For example: old Ford tie rod ends, the 11/16 threaded ones, that are used for split wishbones. The same logic applies to spherical rod ends.
Consider this: in the original application, tie rod ends and spherical rod ends (Heims for ease of typing), were designed for push-pull situations. Yes the stud on a tie rod end is in bending as well as shear, but that is a good safety valve to let the stud bend if the tire hit a pot hole or kerb. The body of the tie rod end was not intended to take bending loads however. Actual Ford tie rod ends were made from good steel alloys, but that doesn't make that much difference when applying a bending load. So we take the wishbones or hairpins and what happens?
The threaded part ends up in bending as well as taking end loads and shear loads. The same is true for Heim joints: they were invented by the Germans for Aircraft use to control moving wing surfaces, allowing some rotation on the stud, whereas the clevis that formerly was used tended to bind on the pivot bolts.
We found out about them from a recovered German fighter plane that the English were able to dissassemble and analyze. The Rose company in England started making them for the British aircraft industry, and they shared the design with us during the runup to WW2. By the way clevises fall in this same category, not originally intended to absorb bending loads.
So you say "Dave G, what's yer point?" Just that we should thank the engineer who specced the originals to have a decent safety factor for the intended usage. I tend to go over sizewhen using these, like using the Ford truck 3/4-16 thread tie rod end, part number something like ES150 I think. The stud is larger, and the threaded body is much larger. Meaning when I use it in an application it wasn't designed for, that there will be a little more margin of safety before failure.
As an aside, the place to keep an eye on for fatigue cracks to start is at the point where the threads just exit whatever it is threaded into. This is the junction point between the bung welded into a split bone for instance and a locknut if used, or right at the end of the threaded bung where the male thread exits the female thread. The root of threads are the highest stressed part of a fastener, and in bending is a good (well bad for us!) place for a crack to start and propagate over time. And since we tend to use a lot of the vehicles we build with these adaptations in a spirited manner and seasonal, there is also the potential for stress corrosion cracking, which can lower the failure stress limit to a fraction of the ultimate strength of the part.
Concerning bolts thru Heims, double shear is preferable if you can design bracketry to accomodate it. Single shear puts the boltin both shear and bending, whereas double shear is pure shear. Shear strength is typically 1/2 of ultimate strength, so a bolt with a 90,000 psi ultimate strength can handle approximately 45,000 psi in shear. To determine the actual load, or force it can handle, we need to know the actual cross sectional area of the bolt where its in shear. So if part of the threads are in the shear zone use the root diameter for calculating the stressed area, otherwise its the area of the bolt body.
An example of the area of a bolt is a 1/2 inch bolt has a cross sectional area of 0.196 square inches, and assuming a bolt ultimate strenght of say 180,000 psi ( a really good bolt) the shear strength is about 90,000 psi, resulting in a load of about 17,670 pounds in shear per shear zone, or for double shear, approximately 35,000 pounds of load. But I'm going to throw a monkey wrench in the works.
Most materials are load rate sensitive, meaning under impact loading they fail at lower stress levels. In Applied Mechanics terms this is called dynamic fracture toughness. Also, many materials (steel can be one) also have a transition temperature, where above that temperature failure tends to be ductile, and below failure tends to be brittle. Anyone who rode or raced snowmobiles back in the late 60s-early 70s may have experienced this, with odd things breaking unexpectedly, like springs, motor mount bolts.
I could go on forever, but I won't. Ask Dan, I used to have diarrhea of the brain frequently when we were bench racing... It comes from having the "Knack"
So please ask me questions if interested, as a retired mechanical engineer with both a graduate degree in Applied Mechanics and a lot of experience in Failure, it helps keep my mind active.
Thanks for taking this for what its worth...
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