Here's how I go about the process of tuning up a receiver.
Regarding guard screw torque and stuff. I have some very deep rooted opinions on the torque of an action's guard screw and barrel tennon. It stems from a situation I once faced. Years ago I ran into an issue with a rifle and the solution seemed to be cranking down harder on the screws. Through a mutual friend I met a guy who makes a very, very good living designing and testing fasteners for the aerospace industry. We began discussing things like shear loads, torsional loading, and friction coefficients as they apply to a 60 degree thread form. The conversation was interesting and when I hung up the phone I assumed that was that. A couple months later I received what I'm about to share with everyone. This guy had performed a bunch of testing on his own and shared it with me. Very, very cool of him to do.
Take from it what you want. I offer it as data. Who'd of thought a lowly screw could generate so much ruckus. Kinda cool.
This was sent to me in PDF format. It didn't copy and paste very well so I went through some of it and tried to clean it up so it'd be easier to read. Sorry if I missed spots.
Regarding our discussion about tensile stress of your ¼-28 fastener, I have outlined some of
the topics. I planned to get back with you much sooner, but discovered I had some of my tools missing. I wasn’t able to do actual load tests until I replaced them. Then we were swamped with a series of orders for our gauge kits and it hasn’t let up since. We have never been this busy. One company alone ordered over 50 kits, 20 of which were on UPS Red basis.
I have a frame and load cell device that can give some actual test results of a fastener. There are calculations for all of the conditions you described, but I like having some actual data.
Anyway, let me go back over some of the questions we discussed on the phone:
Q1: What tensile load is being applied to the connection.
Q2: On a 1/4-28 socket head screw with 82° Csk head made of 15-5 PH CRES Steel, how
much clamping force is generated by 40 inch pounds of torque on the screw.
Q3: How much shear load exists between the screw head and the thread at 40 in lbs.
First, let me mention a few common characteristics of nuts and bolts that you might find useful.
A lot of this you probably already know, but it’s worth stating so we can approach this from the
1) A nut and bolt (or screw) assembly will experience a lower preload after they are assembled as few as three to five times. This is due to the nut conforming to the screw
threads and developing more contact area (more friction area).
2) Calculations and tables for nut and bolt connections are based on new pieces
with manufacturing tolerances (per class of fit) figured in to the equation. Unless
it is noted otherwise, the torque references are usually for dry or lightly oiled screw
threads — not greased or plated components, and definitely not dry-lubed or for
parts that have anti-seize type compounds on them.
In fact Chad, I don’t consider a connection to be secure (from the forces of friction) if
anti-seize compounds are used. It makes for a nice smooth fit, but that connection will
experience 30% higher tension in the screw from those highly lubricious compounds.
When a connection is made with anti-sieze on the threads and under the head of the bolt,
over time it can (and often do) come loose.
If any shock loading occurrs, or worse if vibrations are occurring in the joint, then the nut
will most likely unscrew itself in a short time if anti-seize is used.
3) Vibrations are particularly bad for fasteners. If constant (or even intermittant) vibrations
are present, then a mechanical locking component should be used.
Tang washers are a common type of mechanical lock, but lock-tite compounds might also
work as an anti-vibration method.
4) When choosing a fastener, use fine pitch threads for maximum strength of the joint. But
if you must assembly and reassembly the joint frequently, then a coarse thread is better.
There is less chance of cross threading a coarse thread, and if the joint has a chance of
becomming corroded, then a coarse threaded joint has a much better chance of being
disassembled without damaging the threads.
Your connection should be considered as being in a high vibration area. Of course, the vibrations
are very low frequency, but they are high in magnitude. I am not sure what effect this has on
threaded fasteners. I think that high frequency vibrations at any amplitude would be far worse.
The major interest in most cold fastener joints is tensile strength. There are other problem areas
such as joints that work at high temperatures, or in highly corrosive environments, or joints that
work in shear rather than with tensile loads. But tensile strength is usually the first thing a person considers when selecting a screw or bolt.
If you want to determine the load carrying capability of a screw, you need to know the tensile
strength of the material that it is made of. A Rockwell hardness test will determine it’s tensile
strength, or you can be more accurate by actually tensile testing a coupon of material being
used. Either way you do it, you will have a value to work with. Most engineers just use the bolt
grade to determine its tensile strength, but many aircraft (or highly stressed) bolts are tensile
tested to be certain.
Also there are tables that list the “commonly used” strength values. But I think it’s good to know
how to figure the actual load carrying ability of the screw you choose. Here’s how to do it:
Tensile strength is determined from the formula S = P/A
where S = tensile strength (psi)
P = tensile load (pounds)
A = tensile stress area (square inches)
Based on this, the minimum tensile load requirement for a fastener can be calculated in terms of
P = S A
(Øm/2)2 times pi
X 1.10 for coarse thd. , or
X 1.05 for fine thread.
But here’s one problem — calculating the tensile stress area of a screw isn’t as easy as it looks!
Depending upon the industry you are in, the calculation can be taken as an area that is based
upon the pitch diameter (High Strength Aerospace Fasteners), or at the minor diameter (where
stress rupture strength or fatigue strength is important).
I only know of these two, but there might be more.
Which of these you use, there is still a serious problem in calculating the area of metal that will
become your tensile stress area. That is because you are working on a helix angle of the thread.
Another issue is that Aerospace Standards are based upon leaving only two threads exposed
from the nut and shank of the bolt. Commercial Standards allow up to 6 threads exposed. The
difference is that there will be higher tensile strength in the Aerospace configuration because for
some reason, shortening the exposed threads will reduce strain hardening and notch sensitivity
of the joint.
Of course, if you just determine the tensile strength of the material and calculate the stress area
of the bolt, then do the math and you will have your value. The only way you know it is different
is when you observe the actual tensile test results when breaking test coupons.
The easiest thing to do is look up in a chart what the tensile stress area is calculated to be. Or,
you can do what I do and just calculate the maximum strength of the bolt as the minor diameter
of the screw thread that is used. I know this value is somewhat lower than actual real world tests would prove, but it is so much easier and because safety is always important, you have lower calculated tensile strengths than will actually exist, adding a small amount to your safety margin. Another thing, coarse threads are about 10% stronger than if you calculate based on the minor diameter. Fine pitch threads are about 5% stronger. Here is an easy way to get real close to the actual values, but without all the complicated math:
Use the minor diameter of the thread, devide by 2, square this value, and multiply that
result by pi.
Øm = minor diameter
So, for example, if the minor diameter of a ½-13 thread = .4041
Devide this minor dia. by 2 = .2021 and then square this number = .0408
Multiply this number by pi .0408 x 3.14159 = .1283 square inches.
Multiply that number by 110% (coarse thd.) .1283 x 1.10 = .1411 square inches.
The actual tensile stress area is given as .1419 square inches from the chart. So you can see
that the above method is very close — within 1% of the far more difficult calculation!
Using the tensile stress area of .1419 square inches, we can simply multiply this
area by the tensile strength of the steel that the bolt is made of to find the maximum
load the bolt can handle.
For example, C1018 has an ultimate tensile strength of 125,000 pounds. So, just
multiply this material ultimate strength value by the tensile stress area of the bolt:
.1419 square inches x 125,000 pounds per square inch = 17,738 pounds maximum
But I never use the ultimate tensile strength when I design something. I use the yield strength
of the steel, because I want the maximum strength of the bolt without it going into the plastic
region and distorting. In other words, I want its maximum working strength, or proof load.
That value for C1018 is about 74,000 psi. Therefore the bolt’s proof load is 10,500 pounds. If
you apply the safety factor to this (say 2 :1 factor) you can be pretty sure the bolt will work for a
long time at 5,250 pounds. Of course, safety factors vary greatly for the given circumstance.
So if you just use this method, you can find the tensile stress area of whatever screw you are
working with. The Machinists Handbook or handbook H28 has all the thread data listed. Just
use the minor diameter and work this formula.
If the screw or bolt you are using is in shear rather than in tension, then you can figure the
shear strength at about 75% of the tensile strength. This value works for most low to medium
strength steels. For high strength screws (125,000 UT or higher), this difference can be in the
54% to60% region.
Chad, I’m not done here but given the sorts of delays I run into, I’m going to send this much off
to you right now.
Later I will do some actual load tests on the ¼-28 threads you and I discussed. I will get back
to you after that.
If you have some questions regarding the above, please e-mail to me.
Again, sorry for this long delay. I’ll try to be faster with the rest of the info.