# Self Tightening Bolts, Self Locking Bolts

• posted on November 7, 2005, 3:58 pm
http://square.cjb.cc/bolts.htm
"Self Tightening Bolts theory. Warning: this page is only a theory, not a fact."
That's a good description.
Could someone please explain what self-tightening and self-locking bolts are and give examples. The author may have the latter in mind.
"Figure 4.1 This picture explains the great inertia and centrifugal force"
"When ever there is a difference in inertial force (as pointed out with the arrows) the pulley will move. Not 180-ft-lb torque can hold the pulley still."
I wonder what this is about.
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• posted on November 8, 2005, 2:13 am
karl wrote:

As a complete aside, this reminds me of a former high-school physics teacher's pet rant: "centrifual" force, or the observed outward-from-center force on a spinning object, he would always insist, is not a real force. The real force in play is centripetal force, or the tendency of the point on the object wanting to continue in a straight direction on tangent to the spin, is the ACTUAL force at work. "Centrifugal" force is only an imaginary thing.
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• posted on November 8, 2005, 4:09 pm

What this mean is that the object in motion will move in the direction of movement. But when there is a force that tries to change that motion usually from the engine or transmission the pulley will move when it's not intended to. Scroll down to Figure 4.2. It may explain more about centrifugal force not centripetal force. Centripetal is moving or directed toward a center or axis. The theory is that the centrifugal force can effect the bolt's movement in some way or just simply tighten up bolt.
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• posted on November 8, 2005, 5:27 pm
Burt S. wrote:

http://en.wikipedia.org/wiki/Centrifugal_force
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• posted on November 8, 2005, 11:42 pm

I am unconvinced by this theory.
1) If microscopic ratchet teeth are created to cause the bolt to self-tighten, wouldn't they be destroyed when the god-awful tight bolt is broken loose? The bolt at least should be specified as a "use once" item, regardless of how the mating threads in the crank fare.
2) In order to tighten, the bolt will have to move with respect to the pulley. That means the washer must have similar ratcheting action, and on a similar microscopic level to allow the ratchet to occur with miniscule motion. That means if the washer is less than pristine and is reused the bolt won't self-tighten.
3) The forces are downright outrageous. In round numbers, if the washer diameter is 1/2 inch and the bolt thread diameter is 1/4 inch, to tighten past the 200 ft-lb mark the bolt head has to experience 5000 pounds force from one side to the other, or 10000 pounds force on one side relative to the center. The equivalent force on the thread is double that.
4) If there is significant motion of the pulley relative to the crank, the mating surfaces will wallow out. We see it often enough with splined drive axles that are insufficiently torqued.
Altogether, it doesn't add up. Torsional forces between the pulley and crank must act unidirectionally on the bolt, with several tons of force being transferred through both sides of the washer and without damaging the pulley or crank mating surfaces, with enough movement to materially tighten the bolt. The theorized ratchet mechanism has to operate on a microscopic basis, not be damaged in removal, and to allow effortless unthreading when the bolt is broken loose. It must work over a wide range of lubrication, including a penetrant oil film or being cleaned with brake cleaner. I'm glad I haven't been asked to design something like that, particularly if I could just specify tightening to a different torque in the first place.
Mike
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<%-name%>
• posted on November 9, 2005, 12:58 am

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It's the stresses in the bolt, not the forces acting on the side of it, that matter. Specifically, torquing down on a bolt is the equivalent of stretching it until it holds two things together. The torquing causes the threads to act against each other so as to place the bolt in tension (as opposed to compression).
For correlating torque to the axial load it produces, one finds somewhat crude estimates like that given at the bottom of http://www.engineersedge.com/torque.htm . But of course, this formula will require tweaking depending on conditions. E.g. fine thread vs. coarse thread.
Anyway, it's really about 200 ft-lbs. divided over the six edges of the roughly 1.7/2 cm (= about .33 inch = about 0.028 foot) radius bolt head (for a 91 Civic, for one), anyway. (This Civic's pulley bolt has a 17 mm head and 14 mm nominal diameter.) So something like 200/6/(0.028) = about 1200 pounds is applied to each bolt head edge. Key word being "edge." Then one has to think about what it means to "apply" this force to the whole edge. It's distributed over the surface of the edge, for one thing. If one took 1200 lbs. and set it on a bar of steel with a cross-sectional area of about 1/8 inch by 1/8 inch = 1/64 inch (conservative for this back-of-the-envelope calculation), the stress would still be only 1200*64 = 77000 psi, far below the yield strength of typical steels. And it's not being applied perpendicularly to each face, but more in shear, besides.

Which mating surfaces?

crank
If the above is supposed to relate to your earlier calculation, then I think there's a conceptual error here.

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I have doubts that a cold bolt-pulley-crankshaft assembly would hold up to a hand application of 300 ft-lbs. of tightening torque. 'Cause crude estimators like the one I cite above indicate this would produce in the neighborhood of 300(12)/(.2*.55) = 32700 lbs. of axial load in the bolt, or 32700 / (Pi r^2) = about 137,000 psi of tensile stress in the bolt, which is mighty close to the yield strength (~ 130,000 to 150,000) of many steels. This is too close for engineering comfort.
Which is why I am led to believe galling, aggravated by extreme heat cycling and the high loads of that pulley working on an initially pretty tight bolt, plays at least some role and possibly all of it.
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<%-name%>
• posted on November 9, 2005, 1:49 am
wrote

I think we are talking about two separate things. I'm looking at what is required for force from the theorized pulley movement (in the original link) to tighten the bolt beyond 200 ft-lbs, rather than the tightening being from application of a socket. I don't see how that could be transmitted through the washer, even if pulley/crank movement occurred without wallowing out the mating surface between the crank and pulley.
Miscommunication aside, we seem to be on the same page. The bolt isn't turning to tighten itself, it's just sticking.
Mike
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<%-name%>
• posted on November 9, 2005, 5:59 am

Maybe so, but the page is too damn long. Trim your posts, ferchrysakes!
nb
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<%-name%>
• posted on November 9, 2005, 7:25 pm

from
Oh. That is different. Some of my comments still apply, but I think it's too much of a morass to sort out, under the circumstances.

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I don't claim the bolt sticks when it tightens in operation (in theory). I do propose that the crankshaft-pulley assembly moves relative to the bolt at times.
No big deal. Some time maybe we'll get some studies of whether the bolt does move relative to the shaft under some operating conditions.
Related aside: Does anyone know whether Honda specifies replacing this bolt after so many timing belt changes?
Someone here noted that dealer service shops apparently mark the bolt each time it has been removed. There could be a few reasons for this. I'm thinking one of them is to keep a record of how many times the bolt has been loaded yada a certain way.
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<%-name%>
• posted on November 21, 2005, 11:21 am

snip
| http://www.engineersedge.com/torque.htm | Design Considerations | | The first requirement in determining the amount of torque | to apply is a knowledge of the desired bolt stress. This | stress based on the yield strength of the bolt material. It | is recommended that the induced stress not be allowed to | exceed 80% of the yield strength. In the design of a | fastener application which will be subject to external | loading, whether static or dynamic, it will be necessary to | establish bolt size and allowable stress in accordance with | current engineering practice. | | The mathematical relationship between torque applied and | the resulting tension force in the bolt has been determined | to be as follows: | | T = Torque required (inch pounds) | F = Bolt tension desired (Axial Load) (pounds). | D = Nominal bolt diameter. (major dia.) | EQUATION: T = .2 D F | | This relationship is based on the assumption that regular | series nuts and bolts with rolled threads are used, acting | on surfaces without lubrication.
What a rubbish! This formula is simply wrong, dead wrong! The bolt diameter is irrelevant, but the pitch, which is inversely proportional to the Force, is missing from this formula.
"The [CORRECT] mathematical relationship between torque applied and the resulting tension force in the bolt," ignoring friction, is:
T = Torque required F = Bolt tension or compression desired (Axial Load) P = Pitch
T = P*F/2*Pi, or F = T*2*Pi/P
It is likely that the constant ".2" in the wrong formula
T = .2 D F
is chosen such that for common threads reasonable results are obtained, but it is irresponsible not to point out the limitation of this formula.
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<%-name%>
• posted on November 21, 2005, 3:44 pm

===========================================================================> > TOPIC: Self Tightening Bolts, Self Locking Bolts

===========================================================================> >

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Karl, it's a formula for approximating. Too many non-engineers operate under the illusion that engineering is an exact science. It's usually not. (Just as medicine is an inexact science.) Engineering computations are fraught with assumptions and of course limitations. Torque-axial load relationships for bolts are a great example of why engineering can't be an exact science per se and so approximating formulae are often appropriate. For one thing, as has been pointed out, friction effects vary a good deal and fairly unpredictably over the life of a bolt, and can drastically affect the torque-axial load relationship. For another, material manufacture means the strength of the material cannot be known precisely. For a third, geometries are inexact from the get-go. For a fourth, as materials are loaded and unloaded, their material properties may change, so over the life of, say, a bolt, the load at which it may fail can go down.
We can only approximate the torque-axial load relationship and build in factors of safety to anticipate worse case scenarios.
Diameter is relevant. The derivation of the formula is complicated. I could not do it off the top of my head, despite having quite a bit of experience teaching strength of materials subjects (that is, teaching the design of beams, pillars, fasteners, etc.; anything that has shear or axial stress in it upon angular or axial loading).
Marks Standard Handbook for Mechanical Engineers' has another formula which you might like more, assuming you could accept that figuring out how the geometry of threads "causes" torque to become axial load is not an easy task:
F = 2 Pi T / (L + kL sec b sec d cosec b + k ' D 3 Pi / 2)
L = the pitch b = thread angle k = the coefficient of friction d = sec (angle between faces of thread/2) k ' = coefficient of friction between nut and seat (bolt face and washer?)
(Hopefully I copied this correctly. It's probably on the net somewhere.)
As you can see, thread diameter still of course plays a role.
I'm sure we could find several more formulae, good for certain conditions and to a particular degree of certainty. Yet another appears below. Took about ten seconds of googling effort. I just pulled up the first site that came up in a google search for {torque bolt formula load}.

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Don't know where you got this, but its omission of diameter says a lot.
Here's another formula:
T = Fp * K * d
d here is diameter.

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Karl, you evidently missed the qualifier above, stating that the formula could be used as an approximation for /regular series/ nuts and bolts with rolled threads, etc.
Any competent engineer knows that formulae such as the one at the site above is an approximation and of course has limitations, at least some of which are stated at the site.
In sum, as much as I hate to be dismissive, the reality is that this is a complicated subject. Grasp of the precise nature of torque-load relationships requires study and high achievement in several college level engineering courses.
OTOH, my sense is that a lot of folks here do have a feel for how torque does cause axial load; the effects of friction, diameter, and pitch; etc. So some simple truths (or attempts to get at the truth) can be discussed and analyzed and even debated.
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<%-name%>
• posted on November 22, 2005, 6:27 am
Elle wrote:
snip

Elle, I wish you a nice Thanksgiving, and I will respond after that.
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<%-name%>
• posted on November 29, 2005, 7:23 am

The basic laws used in engineering are exact, just like in physics.

Yes, I was wrong stating, "The bolt diameter is irrelevant." This is only true without friction. I was too fast - didn't think through it.
When developing relationships one starts from simple systems and refines them as needed. In this case one starts assuming no friction:
F = force (axial load, tension) T = torque W = work s = distance traveled P = pitch
Example: Lifting 1 pound 1 foot:
W = F*s = 1 lb * 1 ft = 1 ftlb
Applying this to bolts and nuts: the axial work (with s=P) is equal to the rotational work:
W = F*P = 2*Pi*T* and F = 2*Pi*T / P
This is the basic relationship between the tension in the bolt and the applied torque, ignoring friction. It shows that the tension is proportional to the applied torque and inversely proportional to the pitch (as one would guess). The diameter is irrelevant in this ideal case assuming no friction.
In the real world, with friction present, both components of it - the friction at the face and the friction in the threads - are depending on the diameter.

Not I "missed the qualifier" - it is unequivocally stated,
> The mathematical relationship between torque applied > and the resulting tension force in the bolt has been > determined to be as follows:
Now, this is very clear. But this formula omits the pitch on which the tension is inversely proportional. Elle, you, it seems, missed my qualifier, that the formula I showed applies when "ignoring friction."

That applies to you, but I bet there are many people who, when they see the formula "T = .2 D F" trust it to be true and believe pitch is irrelevant - there is no place for it.

I don't think so. Even incorporating friction doesn't require special knowledge, just a little math and physical understanding.

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<%-name%>
• posted on November 29, 2005, 6:51 pm
Elle:

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The laws of physics are exact, and engineering design certainly does rely on laws of physics. But engineering design also takes into account the inability to ascertain quantities accurately. Above, I named several examples of quantities that cannot be measured accurately and how they figure into fastener design. Judgment based on experience and not without some subjectivity is essential to the engineering design process. Science by itself does not--and cannot--build car engines, etc. Engineering does.

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It's how much the bolt deflects under the axial load F, not pitch P, that should be used here.
For one thing, when tightening a bolt, the threads actually move a little farther apart from each other. So under axial load, the pitch changes.
The higher the axial load, the more the pitch will be off from its design value.
Bolt advance under no load (and due to the effects of rotation and the amount of pitch) is different from bolt deflection due to axially loading the bolt.

Omitting friction, it's more something like:
dW = F(x) dx = T(omega) domega
where x = deflection in the axial direction omega = angular deflection in radians F(x) = axial load, which is a function of deflection x T(omega) = torque, which is a function of angular deflection omega

bolt and the

This is a nice first attempt and certainly shows some understanding of the force and torque relationships in bolts, but it does have, for one, the blatant mistake I identify above.