TORQUE Program Example Calculation - Metric Units Presented below are the results from TORQUE for a M12 Grade 8.8 bolt with a nylon patch type locking device that creates a prevailing torque. (These calculations are in metric units, the TORQUE program can also work in units of inches and pounds and work with the unified thread form.) TORQUE TIGHTENING ANALYSIS RESULTS Example calculation for a M12 bolt. Torque tightening analysis for a M12 bolt. FASTENER DETAILS Fastener Diameter = 12.00 mm Fastener Shank Diameter = 12.00 mm Thread Pitch = 1.75 mm Included angle between the thread flanks = 60.00 degrees Thread Pitch Diameter = 10.863 mm Thread Root Diameter = 9.853 mm Diameter related to the Thread Stress Area = 10.358 mm Thread Stress Area = 84.264 mm² Thread Root Area = 76.248 mm² Bearing Area under Nut/Bolt Head = 99.620 mm² Fastener Outer Bearing Diameter = 17.20 mm Fastener Inner Bearing Diameter = 13.00 mm Fastener Clearance Hole Diameter = 13.00 mm Effective friction diameter of nut/bolt = 15.20 mm Fastener Yield Strength = 640.00 N/mm² JOINT ASSEMBLY DETAILS Black oxide steel external thread, no finish on steel internal thread, no lubricant. Black oxide steel nut or bolt, no oil, machined steel bearing surface. Prevailing torque caused by a nylon/polyester patch on the threads. Thread Friction Value = 0.120 Nut/Bolt Head Friction Value = 0.120 TORQUE TIGHTENING ANALYSIS RESULTS Yield Point Tightening Factor specified = 0.90 Total Tightening Torque = 83.64 Nm
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Transcript
TORQUE Program
Example Calculation - Metric Units
Presented below are the results from TORQUE for a M12 Grade 8.8 bolt with a nylon patch type locking device that creates a prevailing torque. (These calculations are in metric units, the TORQUE program can also work in units of inches and pounds and work with the unified thread form.)
TORQUE TIGHTENING ANALYSIS RESULTS Example calculation for a M12 bolt. Torque tightening analysis for a M12 bolt. FASTENER DETAILSFastener Diameter = 12.00 mmFastener Shank Diameter = 12.00 mmThread Pitch = 1.75 mmIncluded angle between the thread flanks = 60.00 degreesThread Pitch Diameter = 10.863 mmThread Root Diameter = 9.853 mmDiameter related to the Thread Stress Area = 10.358 mmThread Stress Area = 84.264 mm²Thread Root Area = 76.248 mm²Bearing Area under Nut/Bolt Head = 99.620 mm²Fastener Outer Bearing Diameter = 17.20 mmFastener Inner Bearing Diameter = 13.00 mmFastener Clearance Hole Diameter = 13.00 mmEffective friction diameter of nut/bolt = 15.20 mmFastener Yield Strength = 640.00 N/mm² JOINT ASSEMBLY DETAILSBlack oxide steel external thread, no finish on steelinternal thread, no lubricant. Black oxide steel nutor bolt, no oil, machined steel bearing surface. Prevailingtorque caused by a nylon/polyester patch on the threads.Thread Friction Value = 0.120Nut/Bolt Head Friction Value = 0.120 TORQUE TIGHTENING ANALYSIS RESULTSYield Point Tightening Factor specified = 0.90Total Tightening Torque = 83.64 NmThis torque is composed from:Torque needed to extend the fastener = 8.98 NmTorque needed to overcome thread friction = 24.26 NmTorque needed to overcome nutface friction = 29.40 NmPrevailing Torque Value = 21.00 Nm FORCE ANALYSIS RESULTSFastener Preload = 32239.37 NDirect Force that would Yield the Fastener = 53928.91 NPreload as a percentage of Yield Force = 59.78 %
MAXIMUM STRESSES INDUCED INTO THE FASTENERPercentage of the yield strength utilised = 90.00 %Von-Mises Equivalent Stress = 576.00 N/mm²Tensile Stress due to Preload = 382.60 N/mm²Torsional Stress due to the applied torque = 248.59 N/mm²Surface Pressure under the Nut Face = 323.62 N/mm²
Bolt Preload Calculation
Question: How is bolt installation preload calculated?
Answer: Bolt pretension, also called preload or prestress, comes from the installation torque T you apply when you install the bolt. The inclined plane of the bolt thread helix converts torque to bolt pretension. Bolt preload is computed as follows.
Pi = T/(K D) (Eq. 1)
where Pi = bolt preload (called Fi in Shigley). T = bolt installation torque. K = torque coefficient. D = bolt nominal shank diameter (i.e., bolt nominal size).
Torque coefficient K is a function of thread geometry, thread coefficient of friction t, and collar coefficient of friction c. Look up K for your specific thread interface and collar (bolt head or nut annulus) interface materials, surface condition, and lubricant (if any). ("Torque specs for screws," Shigley, and various other sources discuss various K value estimates.) If you cannot find or obtain K from credible references or sources for your specific interfaces, then you would need to research to try to find the coefficients of friction for your specific interfaces, then calculate K yourself using one of the following two formulas listed below (Shigley, Mechanical Engineering Design, 5 ed., McGraw-Hill, 1989, p. 346, Eq. 8-19, and MIL-HDBK-60, 1990, Sect. 100.5.1, p. 26, Eq. 100.5.1, respectively), the latter being far simpler.
K = {[(0.5 dp)(tan + t sec )/(1 – t tan sec )] + [0.625 c D]}/D (Eq. 2)
K = {[0.5 p/] + [0.5 t (D – 0.75 p sin )/sin ] + [0.625 c D]}/D (Eq. 3)
where D = bolt nominal shank diameter. p = thread pitch (bolt longitudinal distance per thread). = thread profile angle = 60° (for M, MJ, UN, UNR, and UNJ thread profiles). = thread profile half angle = 60°/2 = 30°. tan = thread helix angle tan = p/( dp). dp = bolt pitch diameter. t = thread coefficient of friction. c = collar coefficient of friction.
D and p can be obtained from bolt tables such as Standard Metric and USA Bolt Shank Dimensions.
The three terms in Eq. 3 are axial load component (coefficient) of torque resistance due to (1) thread helix inclined plane normal force, (2) thread helix inclined plane tangential (thread friction) force, and (3) bolt head or nut washer face friction force, respectively.
However, whether you look up K in references or calculate it yourself, the engineer must understand that using theoretical equations and typical values for K and coefficients of friction merely gives a preload estimate. Coefficient of friction data in published tables vary widely, are often tenuous, and are often not specific to your specific interface combinations and lubricants. Such things as unacknowledged surface condition variations and ignored dirt in the internal thread can skew the results and produce a false indication of preload.
The engineer and technician must understand that published K values apply to perfectly clean interfaces and lubricants (if any). If, for example, the threads of a steel, zinc-plated, K = 0.22, "dry" installation fastener were not clean, this might cause K to increase to a value of 0.32 or even higher. One should also note that published K values are intended to be used when applying the torque to the nut. The K values will change in relation to fastener length and assembly running torque if the torque is being read from the bolt head.
One should measure the nut or assembly "running" torque with an accurate, small-scale torque wrench. ("Running" torque, also called prevailing torque, is defined as the torque when all threads are fully engaged, fastener is in motion, and washer face has not yet made contact.) The only torque that generates bolt preload is the torque you apply above running torque.
A few more things to be aware of are as follows. Bolt proof strength Sp is the maximum tensile stress the bolt material can withstand without encountering permanent deformation. Published bolt yield strengths are determined at room temperature. Heat will lower the yield strength (and proof strength) of a fastener. Especially in critical situations, you should never reuse a fastener unless you are certain the fastener has never been yielded.
If a more accurate answer for bolt preload is needed than discussed above, the specific combination and lubricant would have to be measured instead of calculated. Measurement methods are generally involved, time-consuming, and expensive, and are beyond the scope of this article. But perhaps one of the simplest and least expensive methods, to test specific combinations and lubricants, is to measure the installed fastener with a micrometer, if possible, and compute torque coefficient K as follows, per Shigley, op. cit., p. 345, para. 2.
K = T L/(E A delta D) (Eq. 4)
Where T = bolt installation torque, L = bolt grip length, E = bolt modulus of elasticity, A = bolt cross-sectional area, D = bolt nominal shank diameter, and delta = measured bolt elongation in units of length.
Bolt Torque Chart
Suggested Starting Values
The below estimated torque calculations are only offered as a guide. Use of its content by anyone is the sole responsibility of that person and they assume all risk. Due to many variables that affect the torque-tension relationship like human error, surface texture, and lubrication the only way to determine the correct torque is through experimentation under actual joint and assembly conditions.
3. Torque has been converted into ft/lbs by dividing the result of the formula by 12
4. All calculations are for Coarse Thread Series (UNC).
5. Grade 2 calculations only cover fasteners 1/4"-3/4" in diameter up to 6" long; for
longer fasteners the torque is reduced significantly.
6. Clamp loads are based on 75% of the minimum proof loads for each grade and size.
7. Proof load, stress area, yield strength, and other data is based on IFI 7th Edition (2003)
Technical Data N-68, SAE J429, ASTM A307, A325, A354, A449, and A490.
Methods of Tightening Threaded FastenersWe have a web site dedicated to training, have a look at www.bolting.info - the material on this site provides additional information on this topic.
One of the major problems with the use of bolted joints is the precision, with regard to achieving an accurate preload, of the bolt tightening method selected. Insufficient preload, caused by an inaccurate tightening method, is a frequent cause of bolted joint failure. It is important for the Designer to appreciate the features and characteristics of the main methods employed to tighten bolts. Presented below is a brief summary of the major bolt tightening methods. Note however that whatever method is used to tighten a bolt, a degree of bolt preload scatter is to be expected.
There are six main methods used to control the preload of a threaded fastener. Specifically:
1. Torque control tightening.
2. Angle control tightening.
3. Yield controlled tightening.
4. Bolt stretch method.
5. Heat tightening.
6. Use of tension indicating methods.
Torque Control Tightening Controlling the torque which a fastener is tightened to is the most popular means of controlling preload. The nominal torque necessary to tighten the bolt to a given preload can be determined either from tables, or, by calculation using a relationship between torque and the resulting bolt tension.
When a bolt is tightened the shank sustains a direct stress, due to the elongation strain, together with a torsional stress, due to the torque acting on the threads. Most tables of bolt tightening torques ignore the torsional stress and assume a direct stress in the threads of some proportion of the bolts yield stress, usually 75%. For high frictional conditions the magnitude of the torsional stress can be such that when combined with the direct stress, an equivalent stress over yield can result, leading to failure. A more consistent approach is to determine the magnitude of the direct stress which, when combined with the torsional, will give an equivalent stress of some proportion of yield. The proportion commonly used with this approach is 90%.
Torque prevailing fasteners (such as Nyloc, Cleveloc nuts etc.) are often used where there exists a risk of vibration loosening. The prevailing torque has the effect of increasing the torsional stress in the bolt shank during tightening. This affects the conversion of the tightening torque into bolt preload and should be allowed for when determining the correct torque value for this type of fastener.
As can been seen by study of the above chart, a fundamental problem with torque tightening is that because the majority of the torque is used to overcome friction (usually between 85% and 95% of the applied torque), slight variations in the frictional conditions can lead to large changes in the bolt preload. This effect can be reduced by the use of so called friction stabilisers. These are substances which are coated onto the fasteners to reduce the frictional scatter. Other ways to improve the accuracy of the method are:
1. Do not use plain washers; their use can result in relative motion to change from the nut to washer, to washer to joint surface, during tightening. This as the effect of changing the friction radius and hence affects the torque-tension relationship. If, because of excessive bearing pressure, a larger bearing face is required, thought should be given to the use of flanged nuts and bolts.
2. Determine the correct tightening torque by the completion of tests. Strain gauges can be attached to the bolt shank and tightening completed on the actual joint. A load cell
under the bolt head can be used, however it is not as accurate as strain gauging, since the joint characteristics have been changed.
3. If it is not feasible to establish by testwork the actual tightening torque, determine the tightening torque using the best information available i.e. fastener finish, nut head bearing surface size and prevailing torque characteristics, if applicable. (The computer program TORQUE developed by Bolt Science can allow for all these effects.)
4. Ensure that the tightening torque value is specified on the assembly drawing. Quotation of a plus or minus 5% tolerance is good practice. More unusually, quote that a calibrated torque wrench is to be used to check the torque after installation. The method used to tighten the bolt has a significant influence on the preload scatter (see below).
Angle Controlled Tightening This method, also known as turn of the nut method, was introduced for manual assembly shortly after the second World War when a certain tightening angle was specified. The method has been applied for use with power wrenches, the bolt being tightened to a predetermined angle beyond the elastic range and results in a small variation in the preload due, in part, to the yield stress tolerance. The main disadvantages of this method lie in the necessity for precise, and, if possible, experimental determination of the angle; also the fastener can only sustain a limited number of re-applications before it fails. Yield Controlled Tightening This method, developed by the SPS organisation, is also known under the proprietary name "Joint Control Method". Very accurate preloads can be achieved by this method by minimising the influence of friction and its scatter. The method has its roots in a craftsman's "sense of feel" on the wrench which allowed him to detect the yield point of the fastener with reasonable precision. With the electronic equivalent of this method, a control system is used which is sensitive to the torque gradient of the bolt being tightened. Rapid detection of the change in slope of this gradient indicates the yield point has been reached and stops the tightening process. This is achieved by incorporating sensors to read torque and angle during the tightening process. Since angle of rotation and torque are both measured by the control system, permissible values can be used to detect fasteners which lie outside their specification (having too low a yield for example).
A small degree of preload scatter still results from this method due to the influence of friction. The method detects the yield point of the fastener under the action of combined tension and torsion. The higher the thread friction, the higher the torsional stress, which, for a given yield value, results in a lower preload due to a lower direct stress.
The method has been used in critical applications, such as cylinder head and conn-rod bolts, in order that consistently high preloads can be achieved (which can allow smaller bolts to be used). However, because of the cost of the tools necessary to use this method (a hand wrench incorporating the control circuitry costs many times more than a conventional torque wrench), widespread adoption of this method is unlikely. (Although manufacturers may be able to invest in the equipment, unless service staff have similar
equipment, the Designer cannot depend upon high preloads being maintained in the field.)
Bolt Stretch Method A problem relating to the tightening of large bolts is that very high tightening torques are required. Although this can be partly overcome by the use of hydraulic torque wrenches (the reaction of the torque however can be a problem), the use of hydraulic tensioning devices is commonplace for bolts over 20mm in diameter. The method uses a small hydraulic ram which fits over the nut, the threaded portion of the bolt/stud protrudes well past the nut and a threaded puller is attached. Hydraulic oil from a small pump acts upon the hydraulic ram which in turn acts upon the puller. This is transmitted to the bolt resulting in extension occurring. The nut can then be rotated by hand with the aid of an integral socket aided by a tommy bar.
Control of the hydraulic pressure effectively controls the preload in the bolt. A small amount of preload reduction however does occur when the pressure is removed as the nut elastically deforms under the load. Removal of nuts corroded to the bolts can be a problem with this method.
Heat Tightening Heat tightening utilises the thermal expansion characteristics of the bolt. The bolt is heated and expands: the nut is indexed (using the angle of turn method) and the system allowed to cool. As the bolt attempts to contract it is constrained longitudinally by the clamped material and a preload results. Methods of heating include direct flame, sheathed heating coil and carbon resistance elements. The process is slow, especially if the strain in the bolt is to be measured, since the system must return to ambient temperature for each measurement. This is not a widely used method and is generally used only on very large bolts.
Tension Indicating Methods This category includes the use of special load indicating bolts, load indicating washers and the use of methods which determine the length change of the fastener. There are a wide number of ways bolt tension can be indirectly measured and the discussion presented here is not exhaustive.
Special bolts have been designed which will give an indication of the force in the bolt. One such fastener is the Rotabolt which measures bolt extension by the use of a central gauge pin which passes down a centrally drilled hole in the bolt. Underneath the head of the gauge pin, a rota is retained which is free to spin in a very accurately set gap. The fastener stretches elastically, whereas the gauge pin does not move since it experiences no load. As tightening continues, the bolt will stretch sufficiently to eliminate the gap and prevent the rota from being able to be rotated. This is the indication that the bolt is correctly loaded. Another proprietary fastener uses a similar method. The HiBolt uses a pin located centrally down the bolt as does the Rotabolt except the pin is gripped by the slight contraction of the bolt diameter; the pin being locked when the correct preload is reached.
The use of load indicating washers is widespread in structural engineering. Such washers have small raised pips on their surface which plastically deform under load. The correct preload is achieved when a predetermined gap is present between the washer and the underhead of the bolt. This is measured using feeler gauges. Generally they are not used in mechanical engineering, but are, extensively, in civil engineering.
The extension which a bolt experiences can be measured either using a micrometer or by a more sophisticated means such as using ultrasonics. The extension can be related to preload either directly, by calibration, or indirectly, by calculation. If ultrasonic measurement is used then the end of the bolt shank and the head may require surface grinding to give a good acoustic reflector.
To assist the Engineer in overcoming the problems associated with the use of threaded fasteners and bolted joints, Bolt Science has developed a number of computer programs. These programs are designed to be easy to use so that an engineer without detailed knowledge in this field can solve problems related to this subject.
Suggested Tightening Torque Values to Produce Corresponding Bolt Clamping Loads
1. Tightening torque values are calculated from the formula T = KDP, where T= tightening torque. lb-in. K=torque-friction coefficient; D = nominal bolt diameter. in; and P = bolt clamp load developed by tightening. lb.
2. Clamp load is also known as preload or initial load in tension on bolt. Clamp load (lb) is calculated by arbitrarily assuming usable bolt strength is 75% of bolt proof load(psi) times tensile stress area(sq in.) of threaded section of each bolt size. Higher or lower values of clamp load can be used depending on the application requirements and the judgement of the designer.
3. Tensile strength (min psi) of all Grade 7 bolts is 133,000. Proof load is 105,000 psi.
4. Tensile strength (min psi) of all Grade 8 bolts is 150,000 psi. Proof load is 120,000 psi. Ref.:Fastening Reference, Machine Design, Nov 1977.
What is the Proper Torque to Use on a Given Boltby Joe Greenslade
"What torque should I use to tighten my bolts?" is a question suppliers of bolts are frequently asked by end user customers. Many times I have been asked if a chart is published on the recommended tightening torque for various bolt grades and sizes. I do not know of any. This article provides such a chart for "Initial Target Tightening Torque. It See Figure 1. The formula for generating these values is explained below.
The widely recognized engineering formula, T= K x D x P (to be explained later in this article), was used to provide the chart's values, but it must be understood that every bolted joint is unique and the optimum tightening torque should be determined for each application by careful experimentation. A properly tightened bolt is one that is stretched such that it acts like a very ridged spring pulling mating surfaces together. The rotation of a bolt (torque) at some point causes it to stretch (tension). Several factors affect how much tension occurs when a given amount of tightening torque is applied. The first factor is the bolt's diameter. It takes more force to tighten a 3/4-10 bolt than to tighten a 318-16 bolt because it is larger in diameter. The second factor is the bolt's grade. It takes more force to stretch an SAE Grade 8 bolt than it does to stretch an SAE Grade 5 bolt because of the greater material strength. The third factor is the coefficient of friction, frequently referred to as the "nut factor." The value of this factor indicates that harder, smoother, and/or slicker bolting surfaces, such as threads and bearing surfaces, require less rotational force (torque) to stretch (tension) a bolt than do softer, rougher, and stickier surfaces. The basic formula T = K x D x P stated earlier takes these factors into account and provides users with a starting point for establishing an initial target tightening torque.
• T Target tighten torque (the result of this formula is in inch pounds, dividing by 12 yields foot pounds
• K Coefficient of friction (nut factor), always an estimation in this formula
• D Bolts nominal diameter in inches
• P Bolt's desired tensile load in pounds (generally 75% of yield strength)
The reason all applications should be evaluated to determine the optimum tightening torque is that the K factor in this formula is always an estimate. The most commonly used bolting K factors arc 0.20 for plain finished bolts, 0.22 for zinc plated bolts, and 0.10 for waxed or highly lubricated bolts. .
The only way to properly determine the optimum tightening torque for a given application is to simulate the exact application. This should be done with a tension indicating device of some type on the bolt in the application. The bolt is tightened until the desired P (load) is indicated by the tension indicating device. The tightening torque required to achieve the desired tension is the actual tightening torque that should be used for
that given application. It is extremely important to realize that this tightening value is valid only so long as all of the aspects of the application remain constant Bolt suppliers sometimes have customers say that their bolts are no good because they have started breaking while being installed. Thorough investigation commonly reveals that the customer has started lubricating the bolts to make assembly easier, but maintained to same torque as was used when the were plain finished
The table in this article shows that by using this formula a 1/2-13 Grade 5 plain bolt should be tightened to 82 foot pounds, but the same bolt that is waxed only requires 41 foot pounds to tighten the same tension. A perfect 1/2-13 Grade 5 waxed bolt will break if it is tightened to 81 foot pounds because the K factor is drastically lower. The bolts are fine, but the application changed. Suppliers need to understand this and be able to educate their customers to resolve this common customer complaint about breaking bolts.
The chart is provided for quick reference by fastener suppliers and users for selecting an initial target tightening torque. This chart was derived by using the formula shown earlier. An example of the calculation is as follows:
Product: 3/4-10 Grade 5 zinc plated boltFormula: T= K x D x P
Hopefully the chart will help suppliers with an initial answer to the customer's question, "What torque should I use to tighten my bolts?" Keep in mind this is only an estimated value. It may provide satisfactory performance, but it also may not. Every application should be evaluated on its own to determine the optimum torque value for each application. Major bolt suppliers should have tension indicating equipment necessary to help their customers determine the appropriate tightening values for their specific applications. Keep in mind that if the lubricant on a bolt and nut combination is changed, the tightening torque value must be altered to achieve the desired amount of bolt tension.
Bolt Torque - Screw Torque Data
Suggested maxium torque values for different material and grade bolts &
screws.
Torque, is the measurement of the turning or twisting force applied to an object. The desired result is to hold two
parts together with a tension or clamping force that is greater than any external force that could possibly seperate
them. The part then remains under constant stress and is immune to fatigue.
These charts apply to clean and dry parts. A lubricated bolt requires less torque to attain the same clamping force
as a non-lubricated bolt.
The values are stated in Inch pounds.
Bolt SizeThds Per
Inch
Low Carbon Steel
18-8 St. St.
Yellow Brass
Silicon Bronze
Aluminum 2024-
T4
316 St. St.
Monel
0 80 1.0 --- --- --- --- --- ---
1 64 1.5 --- --- --- --- --- ---
72 2.0 --- --- --- --- --- ---
256 2.2 2.5 2.0 2.3 1.4 2.6 2.5
64 2.7 3.0 2.5 2.8 1.4 3.2 3.1
348 3.5 3.9 3.2 3.6 2.1 4.0 4.0
56 4.0 4.4 3.6 4.1 2.4 4.6 4.5
440 4.7 5.2 4.3 4.8 2.9 5.5 5.3
48 5.9 6.6 5.4 6.1 3.6 6.9 6.7
540 6.9 7.7 6.3 7.1 4.2 8.1 7.8
44 8.5 9.4 7.7 8.7 5.1 9.8 9.6
632 8.7 9.6 7.9 8.9 5.3 10.1 9.8
40 10.9 12.1 9.9 11.2 6.6 12.7 12.3
832 17.8 19.8 16.2 18.4 10.8 20.7 20.2
36 19.8 22.0 18.0 20.4 12.0 23.0 22.4
1024 20.8 22.8 18.6 21.2 13.8 23.8 25.9
32 29.7 31.7 25.9 29.3 19.2 33.1 34.9
1224 --- --- --- --- --- --- ---
28 --- --- --- --- --- --- ---
1/420 65.0 75.2 61.5 68.8 45.6 78.8 85.3
28 90.0 94.0 77.0 87.0 57.0 99.0 106.0
5/1618 129 132 107 123 80 138 149
24 139 142 116 131 86 147 160
3/8 16 212 236 192 219 143 247 266
24 232 259 212 240 157 271 297
7/1614 338 376 317 349 228 393 427
20 361 400 327 371 242 418 451
1/213 465 517 422 480 313 542 584
20 487 541 443 502 328 565 613
9/1612 613 682 558 632 413 713 774
18 668 752 615 697 456 787 855
5/811 1000 1110 907 1030 715 1160 1330
18 1140 1244 1016 1154 798 1301 1492
3/410 1259 1530 1249 1416 980 1582 1832
16 1230 1490 1220 1382 958 1558 1790
7/89 1919 2328 1905 2140 1495 2430 2775
14 1911 2318 1895 2130 1490 2420 2755
18 2832 3440 2815 3185 2205 3595 4130
14 2562 3110 2545 2885 1995 3250 3750
The values are stated in foot pounds.
Bolt Size (inches)
Thds Per Inchttc
SAE 0-1-274,000 psi
Low Carbon Steel
SAE Grade 3
100,000 psi Med Carbon Steel
SAE Grade 5
120,000 psi Med Carbon Heat T. Steel
SAE Grade 6
133,000 psi Med Carbon Temp. Steel
SAE Grade 7
133,000 psi Med Carbon
Alloy Steel
SAE Grade 8
150,000 psi Med Carbon
Alloy Steel
1/4 20 6 9 10 12.5 13 14
28 --- --- --- --- --- ---
5/1618 12 17 19 24 25 29
24 --- --- --- --- --- ---
3/816 20 30 33 43 44 47
24 --- --- --- --- --- ---
7/1614 32 47 54 69 71 78
20 --- --- --- --- --- ---
1/213 47 69 78 106 110 119
20 --- --- --- --- --- ---
9/16 12 69 103 114 150 154 169
18 --- --- --- --- --- ---
5/811 96 145 154 209 215 230
11 --- --- --- --- --- ---
3/410 155 234 257 350 360 380
10 --- --- --- --- --- ---
7/89 206 372 382 550 570 600
9 --- --- --- --- --- ---
18 310 551 587 825 840 700
8 --- --- --- --- --- ---
1-1/87 480 794 872 1304 1325 1430
7 --- --- --- --- --- ---
1-1/47 375 1105 1211 1815 1825 1975
7 --- --- --- --- --- ---
1-3/86 900 1500 1624 2434 2500 2650
6 --- --- --- --- --- ---
1-1/26 1100 1775 1943 2913 3000 3200
6 --- --- --- --- --- ---
1-5/85.5 1470 2425 2660 3985 4000 4400
5.5 --- --- --- --- --- ---
1-3/45 1900 3150 3463 5189 5300 5650
5 --- --- --- --- --- ---
1-7/85 2360 4200 4695 6980 7000 7600
5 --- --- --- --- --- ---
24.5 2750 4550 5427 7491 7500 8200
4.5 --- --- --- --- --- ---
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What Torque Should Be Used to Tighten Metric Machine Screws?
American Fastener Journal, Mar/Apr 2004 by Greenslade, Joe
1 2 Next
A lot is written about bolt and nut tightening, but little is written about tightening machine screws. It is just as important to carefully select an appropriate tightening torque for securing machine screw joints as it is for securing bolt and nut joints. Properly secured joints are directly related to the quality of the end product assembly. The means of calculating the suggested tightening torque is the same for machine screws as it is for bolts. The values are just smaller.
The most widely used formula for calculating threaded fastener tightening torque is:
T = DKP
Where:
T = Torque (inch pounds and Newton meters; 1Nm = 9 in.Ib.)
D = Nominal thread diameter (expressed in inches; 1 mm = .03937 inches)
K = Nut factor (.22 for zinc electroplating)
P = Pounds of clamping force (75% of yield strength)
There are various strength levels of metric machine screws and each has a different recommended tightening value. ISO has two predominate machine screw strength levels: Property Class 4.8 (close to SAE 6OM) and Property Class 8.8 (close to SAE 12OM). Property Class 4.8 indicates a minimum tensile strength of 480 mega Pascal (MPa). This is equal to approximately 70,000 pounds per square inch (PSI). Property Class 8.8 indicates a minimum tensile strength of 880 mega Pascal (MPa). This is equal to approximately 127,000 pounds per square inch (PSI).
The chart below provides reasonable tightening values, but they are not the optimum tightening values for every application. A far better way to establish a tightening torque for a particular application is by conducting a simple study.
To determine the ideal tightening torque for any particular application joint, do the following:
* Make up 12 of the exact application joints being studied.
* Tighten the machine screws until something in the joint completely fails; then record every failure torque value.
The best failure is the twisting in two of the screw, but this does not always happen. The internal thread may strip; the components may crush or distort. It makes no difference what fails.
* Calculate the average torque value at which this particular joint fails.
* The optimum tightening value for the particular joint being studied is 60% of the average failure value.
CALCULATIONS ARE FINE, BUT TESTING IS SUPERIOR
The correct tightening of all threaded fasteners is critical to obtaining an end product of consistently high quality and dependability. Determining tightening torque by calculations or taking values from charts like the one provided in this article is better than just guessing at what a particular torque should be. The best approach to establishing the optimum tightening torque value for a particular joint is determined by performing the simple study described herein.
Joe Ctreimslade has been active in the fastener industry since 1970. he has held positions with major fastener producers in sales engineering, marketing, product design, manufacturing management, and research and development management.
Mr. Greenslade holds twelve U.S. patents on various fastener related products. he has authored over 136 trade journal articles on fastener applications, manufacturing and quality issues. he is one of the fastener industry 's most frequent speakers at trade association meetings and conferences. he is the youngest person ever inducted to the Fastener Industry Hall of Fame.
Mr. Greenslade is active in numerous fastener industry associations and societies holding office in several of them.
In addition to guiding the activities of Greenslade & Company, Mr. Greenslade works as a consultant with fastener suppliers and end users on product design, applications engineering, and quality issues. In this capacity he works to resolve fastener applications problems, ?? help select the best fastening approaches in new product designs, to assist in the standardization of fasteners used within an organization, and to provide training on various aspects of fastening technology and fastener quality assurance. he also serves as Expert Witness in litigation involving fastener related issues. he can be reached at: phone 817-870-8888, fax 817-870-9199 or email: greensladeandcompany@sbcglobal. net.