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<blockquote data-quote="NesikaChad" data-source="post: 327128" data-attributes="member: 7449"><p>Here's how I go about the process of tuning up a receiver. </p><p></p><p><a href="http://longriflesinc.com/process.html" target="_blank">LongRifles, Inc.</a></p><p></p><p></p><p>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.</p><p></p><p>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.</p><p></p><p>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.</p><p></p><p>Enjoy.</p><p></p><p></p><p>Hi Chad,</p><p>Regarding our discussion about tensile stress of your ¼-28 fastener, I have outlined some of</p><p>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.</p><p>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.</p><p>Anyway, let me go back over some of the questions we discussed on the phone:</p><p></p><p>Q1: What tensile load is being applied to the connection.</p><p></p><p>Q2: On a 1/4-28 socket head screw with 82° Csk head made of 15-5 PH CRES Steel, how</p><p>much clamping force is generated by 40 inch pounds of torque on the screw.</p><p></p><p>Q3: How much shear load exists between the screw head and the thread at 40 in lbs.</p><p>First, let me mention a few common characteristics of nuts and bolts that you might find useful.</p><p>A lot of this you probably already know, but it's worth stating so we can approach this from the</p><p>same perspective.</p><p></p><p>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</p><p>threads and developing more contact area (more friction area). </p><p></p><p>2) Calculations and tables for nut and bolt connections are based on new pieces</p><p>with manufacturing tolerances (per class of fit) figured in to the equation. Unless</p><p>it is noted otherwise, the torque references are usually for dry or lightly oiled screw</p><p>threads — not greased or plated components, and definitely not dry-lubed or for</p><p>parts that have anti-seize type compounds on them.</p><p></p><p>In fact Chad, I don't consider a connection to be secure (from the forces of friction) if</p><p>anti-seize compounds are used. It makes for a nice smooth fit, but that connection will</p><p>experience 30% higher tension in the screw from those highly lubricious compounds.</p><p></p><p></p><p></p><p>When a connection is made with anti-sieze on the threads and under the head of the bolt,</p><p>over time it can (and often do) come loose.</p><p></p><p>If any shock loading occurrs, or worse if vibrations are occurring in the joint, then the nut</p><p>will most likely unscrew itself in a short time if anti-seize is used.</p><p></p><p>3) Vibrations are particularly bad for fasteners. If constant (or even intermittant) vibrations</p><p>are present, then a mechanical locking component should be used.</p><p>Tang washers are a common type of mechanical lock, but lock-tite compounds might also</p><p>work as an anti-vibration method.</p><p></p><p>4) When choosing a fastener, use fine pitch threads for maximum strength of the joint. But</p><p>if you must assembly and reassembly the joint frequently, then a coarse thread is better.</p><p>There is less chance of cross threading a coarse thread, and if the joint has a chance of</p><p>becomming corroded, then a coarse threaded joint has a much better chance of being</p><p>disassembled without damaging the threads.</p><p></p><p>Your connection should be considered as being in a high vibration area. Of course, the vibrations</p><p>are very low frequency, but they are high in magnitude. I am not sure what effect this has on</p><p>threaded fasteners. I think that high frequency vibrations at any amplitude would be far worse.</p><p>The major interest in most cold fastener joints is tensile strength. There are other problem areas</p><p>such as joints that work at high temperatures, or in highly corrosive environments, or joints that</p><p>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.</p><p></p><p>If you want to determine the load carrying capability of a screw, you need to know the tensile</p><p>strength of the material that it is made of. A Rockwell hardness test will determine it's tensile</p><p>strength, or you can be more accurate by actually tensile testing a coupon of material being</p><p>used. Either way you do it, you will have a value to work with. Most engineers just use the bolt</p><p>grade to determine its tensile strength, but many aircraft (or highly stressed) bolts are tensile</p><p>tested to be certain.</p><p></p><p>Also there are tables that list the "commonly used" strength values. But I think it's good to know</p><p>how to figure the actual load carrying ability of the screw you choose. Here's how to do it:</p><p>Tensile strength is determined from the formula S = P/A</p><p>where S = tensile strength (psi)</p><p>P = tensile load (pounds)</p><p>A = tensile stress area (square inches)</p><p>Based on this, the minimum tensile load requirement for a fastener can be calculated in terms of</p><p>P = S A</p><p></p><p></p><p></p><p>(Øm/2)2 times pi</p><p></p><p>X 1.10 for coarse thd. , or</p><p>X 1.05 for fine thread.</p><p></p><p>But here's one problem — calculating the tensile stress area of a screw isn't as easy as it looks!</p><p>Depending upon the industry you are in, the calculation can be taken as an area that is based</p><p>upon the pitch diameter (High Strength Aerospace Fasteners), or at the minor diameter (where</p><p>stress rupture strength or fatigue strength is important).</p><p>I only know of these two, but there might be more.</p><p></p><p>Which of these you use, there is still a serious problem in calculating the area of metal that will</p><p>become your tensile stress area. That is because you are working on a helix angle of the thread.</p><p>Another issue is that Aerospace Standards are based upon leaving only two threads exposed</p><p>from the nut and shank of the bolt. Commercial Standards allow up to 6 threads exposed. The</p><p>difference is that there will be higher tensile strength in the Aerospace configuration because for</p><p>some reason, shortening the exposed threads will reduce strain hardening and notch sensitivity</p><p>of the joint.</p><p></p><p>Of course, if you just determine the tensile strength of the material and calculate the stress area</p><p>of the bolt, then do the math and you will have your value. The only way you know it is different</p><p>is when you observe the actual tensile test results when breaking test coupons.</p><p>The easiest thing to do is look up in a chart what the tensile stress area is calculated to be. Or,</p><p>you can do what I do and just calculate the maximum strength of the bolt as the minor diameter</p><p>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:</p><p>Use the minor diameter of the thread, devide by 2, square this value, and multiply that</p><p>result by pi.</p><p></p><p>Øm = minor diameter</p><p>So, for example, if the minor diameter of a ½-13 thread = .4041</p><p>Devide this minor dia. by 2 = .2021 and then square this number = .0408</p><p>Multiply this number by pi .0408 x 3.14159 = .1283 square inches.</p><p>Multiply that number by 110% (coarse thd.) .1283 x 1.10 = .1411 square inches.</p><p>The actual tensile stress area is given as .1419 square inches from the chart. So you can see</p><p>that the above method is very close — within 1% of the far more difficult calculation!</p><p></p><p></p><p>Using the tensile stress area of .1419 square inches, we can simply multiply this</p><p>area by the tensile strength of the steel that the bolt is made of to find the maximum</p><p>load the bolt can handle.</p><p></p><p>For example, C1018 has an ultimate tensile strength of 125,000 pounds. So, just</p><p>multiply this material ultimate strength value by the tensile stress area of the bolt:</p><p>.1419 square inches x 125,000 pounds per square inch = 17,738 pounds maximum</p><p>But I never use the ultimate tensile strength when I design something. I use the yield strength</p><p>of the steel, because I want the maximum strength of the bolt without it going into the plastic</p><p>region and distorting. In other words, I want its maximum working strength, or proof load.</p><p>That value for C1018 is about 74,000 psi. Therefore the bolt's proof load is 10,500 pounds. If</p><p>you apply the safety factor to this (say 2 :1 factor) you can be pretty sure the bolt will work for a</p><p>long time at 5,250 pounds. Of course, safety factors vary greatly for the given circumstance.</p><p>So if you just use this method, you can find the tensile stress area of whatever screw you are</p><p>working with. The Machinists Handbook or handbook H28 has all the thread data listed. Just</p><p>use the minor diameter and work this formula.</p><p></p><p>If the screw or bolt you are using is in shear rather than in tension, then you can figure the</p><p>shear strength at about 75% of the tensile strength. This value works for most low to medium</p><p>strength steels. For high strength screws (125,000 UT or higher), this difference can be in the</p><p>54% to60% region.</p><p></p><p>Chad, I'm not done here but given the sorts of delays I run into, I'm going to send this much off</p><p>to you right now.</p><p></p><p>Later I will do some actual load tests on the ¼-28 threads you and I discussed. I will get back</p><p>to you after that.</p><p></p><p>If you have some questions regarding the above, please e-mail to me.</p><p>Again, sorry for this long delay. I'll try to be faster with the rest of the info.</p><p></p><p>Kindest regards,</p><p>Jerome</p></blockquote><p></p>
[QUOTE="NesikaChad, post: 327128, member: 7449"] Here's how I go about the process of tuning up a receiver. [url=http://longriflesinc.com/process.html]LongRifles, Inc.[/url] 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. Enjoy. Hi Chad, 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 same perspective. 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. Kindest regards, Jerome [/QUOTE]
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