rod bolts rpm vs stress



rod bolts rpm vs stress

Postby grumpyvette » October 3rd, 2008, 7:18 pm

many guys don,t realize that the rod bolt material and cross sectional area are critical to durability , especially in a high rpm range combo.
while the rods themselves occasionally fail, its much more likely that the rod bolts lost their clamping strength, stretched a bit first and that was a major contributing factor in the bearing failure or the rod failure process.
the cross sectional area of the two rod bolts is usually considerably less than the connecting rod forging in any other area
strength, obviously it depends on materials, design, care in manufacturing and which connecting rods are being compared properly prepared LS7 or L88 big block rods are a whole lot stronger than the stock 3/8" rod bolts big block rods, but many of the better aftermarket rods are significantly stronger that even the l88 rods
I beam rods typically have a balance pad and thats a good feature, typical H beam rods are SUPPOSED TO BE nearly identical in weight, as they are usually machined not castings (obviously they too occasionally need to be balanced)H beams generally have a bit more clearance, both rod designs are good , strength differs between manufacturers but the SCAT rods linked to are at least 50%-100% stronger than the stock chevy rods, Ive got zero problem using either design rod config if its a quality 4340 forged steel rod with 7/16" rod bolts that fits the application
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BOTH THE FORGED ,BIG BLOCK ,7/16" ROD BOLT RODS ARE GOOD CANDIDATES FOR A PERFORMANCE BUILD, AFTER BEING REWORKED AND POLISHED/RE-SIZED and TOP QUALITY ROD BOLTS INSTALLED, and MAGNA TESTED FOR FLAWS (REMEMBER THESE RODS MIGHT BE 40YEARS OLD
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viewtopic.php?f=53&t=510

viewtopic.php?f=53&t=1168

viewtopic.php?f=50&t=428&p=22981&hilit=stroker+tips#p22981

viewtopic.php?f=53&t=3540

viewtopic.php?f=53&t=5064&p=14370#p14370

http://www.hotrod.com/techarticles/hrdp ... ting_rods/

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stock

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better aftermarket

worth reading thru

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graph out the piston velocity


read these links

http://arp-bolts.com/pages/technical_failures.shtml

http://www.hotrod.com/techarticles/stee ... index.html


http://www.oliverconnectingrods.com/dow ... alog07.pdf

interesting info from ARP

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http://www.arp-bolts.com/Tech/TechWhy.html

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http://arp-bolts.mobi/p/tech.php?page=3

http://arp-bolts.mobi/p/tech.php?page=2

http://arp-bolts.com/

http://www.harborfreight.com/1-inch-tra ... r-623.html
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http://www.summitracing.com/parts/arp-1 ... /overview/ $248
http://www.summitracing.com/parts/arp-1 ... /overview/ $186
yes you can find non-name brand rod bolt stretch gauges from about $50-$80
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ARP rod bolts are set up to use a stretch gauge with both ends of the bolt pre-machined for the gauge the bolt packaging from ARP,comes with the correct length the bolts are supposed to reach under the correct pre-load tension, in the instructions OR its available on their web site
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most guys are familiar with use of a torque wrench to tighten rod bolts to the correct preload, but while this gets you very close its not as precise as a rod bolt stretch gauge,
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Other Stresses

It must be realized that the direct reciprocating load is not the only source of stresses in bolts. A secondary effect arises because of the flexibility of the journal end of the connecting rod. The reciprocating load causes bending deformation of the bolted joint (yes, even steel deforms under load). This deformation causes bending stresses in the bolt as well as in the rod itself. These bending stresses fluctuate from zero to their maximum level during each revolution of the crankshaft.

Fastener Load

The first step in the process of designing a connecting rod bolt is to determine the load that it must carry. This is accomplished by calculating the dynamic force caused by the oscillating piston and connecting rod. This force is determined from the classical concept that force equals mass times acceleration. The mass includes the mass of the piston plus a portion of the mass of the rod. This mass undergoes oscillating motion as the crankshaft rotates. The resulting acceleration, which is at its maximum value when the piston is at top dead center and bottom dead center, is proportional to the stroke and the square of the engine speed. The oscillating force is sometimes called the reciprocating weight. Its numerical value is proportional to:
It is seen that the design load, the reciprocating weight, depends on the square of the RPM speed. This means that if the speed is doubled, for example, the design load is increased by a factor of 4. This relationship is shown graphically below for one particular rod and piston


http://www.arp-bolts.com/Tech/TechWhy.html
[QUOTE=blueovalz;928783]This chart confuses me and (for me) indicates the smaller the bolt cross section, the higher the tensile strength?).[/QUOTE]


I did a quick DOUBLE TAKE on that graph the first time also....look closer at the edges of the graph, its points out the STRONGER the material USED the SMALLER the dia. necessary for a given tensile strength, your limited in clearance on rod bolt max size so the material needs to have higher yeild strength, and potential durrability, to increase the rod bolt strength

FROM ARP

"Metallurgy for the Non-Engineer

By Russell Sherman, PE

1. What is grain size and how important is it?

Metals freeze from the liquid state during melting from many origins (called allotropic) and each one of these origins grows until it bumps into another during freezing. Each of these is a grain and in castings, they are fairly large. Grains can be refined (made smaller); therefore, many more of them can occupy the same space, by first cold working and then by recrystallizing at high temperature. Alloy steels, like chrome moly, do not need any cold work; to do this – reheat treatment will refine the grain size. But austenitic steels and aluminum require cold work first. Grain size is very important for mechanical properties. High temperature creep properties are enhanced by large grains but good toughness and fatigue require fine grain size-the finer the better. (High temp creep occurs at elevated temperature and depending on material and load could be as much as .001 per inch/per hour.) All ARP bolts and studs are fine grain – usually ASTM 8 or finer. With 10 being the finest.

2. How do you get toughness vs. brittleness?

With steels, as the strength goes up, the toughness decreases. At too high a strength, the metal tends to be brittle. And threads accentuate the brittleness. A tool steel which can be heat-treated to 350,000 psi, would be a disaster as a bolt because of the threads."

http://www.arp-bolts.com/Tech/TechMetals.html

"are all rod stretch gauges created equal "

obviously no more than all girls are equally good looking
but most of the gauges are functional, some just have more features or more precise calibrations, some are adjustable in length ,over a wider range, some have digital read outs, ETC.

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http://store.summitracing.com/partdetail.asp?autofilter=1&part=SUM%2D900015&N=700+4294854225+115&autoview=sku

http://www.carcraft.com/techarticles/116_0609_using_rod_bolt_stretch_tool/index.html

http://www.chevyhiperformance.com/techarticles/0710ch_proper_engine_fasteners/index.html

youll generally find that 4000 FPM of piston speed with stock parts or 4500fpm with good quality forged parts is the reasonable limit on the lower end stress, but the valve control issues tend to become the limiting factor before the lower assembly causes problems, hitting the engines redline doesn,t mean the engines going to sustain damage, but it generally induces significant stress, stress that WILL eventually cause DAMAGE, it might happen instantly or require hundreds of repetitions BUT it will eventually happen if its exceeded regularly, because STRESS IS CUMULATIVE

generally hydrolic lifters max out at about 6500rpm or lower
and stock rockers and valve trains rarely control valves well even with solid lifters above 7000-7500 rpm

http://www.rustpuppy.org/rodstudy.htm

http://www.csgnetwork.com/pistonspeedcalc.html

keep in mind thats max PEAK engine rpms, that should only rarely be reached ,your engine will NEVER stay together if subjected to those rpms consistantly for more that brief moments before shifting, hold any engine at redline for more than a few seconds and bad and expensive things are likely to happen
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
grumpyvette

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Re: rod bolts rpm vs stress

Postby grumpyvette » October 6th, 2008, 3:09 pm

I see rods and rod bolt failures blamed frequently when engines self destruct at high rpms, but its NOT always what it at first might appear to be....are there any detailed pictures of the rods or rod bolts that failed??? in many cases the source of the problem can be seen with a careful detailed exam, if you don,t know the SOURCE of the problem your doomed to repeat the sequence... and keep in mind a good deal of what might appear to be rod/rodbolt failures, are ACTUALLY the result of over reving the valve train,and loss of valve train control, OR detonation, theres no way to compress a bent valve or broken piston ring land without potentially damaging the rods
rods may fail due to spun bearings, lack of oil flow, etc, but the most comon failures I see blamed on rods are usually valve control issues that resulted in the rod/pistons bending a valve, then trying to compress the broken valve, its hardly a rods fault if the valve fails to retract in time to get out of the way due to valve control issues at high rpms



If you ever get the idea that selecting high quality connecting rods with ARP 7/16" rod bolts is a waste of cash and effort, consider the results when a rod bolt snaps due to stress at high rpms. you might be able to save the rockers,valve covers and intake manifold and water pump

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USE A STRETCH GAUGE and ARP BOLTS, or at LEAST a TORQUE WRENCH

http://store.summitracing.com/partde...5&autoview=sku


or at a MINIMUM, READ THE INSTRUCTIONS CAREFULLY,at least tighten then loosen and re-tighten to spec EACH BOLT ,THREE SEPARATE TIMES with a QUALITY torque wrench


great
http://www.nationaltoolwarehouse.com/

ok

http://www.nationaltoolwarehouse.com/



http://www.hotrodshack.com/torque_settings.htm

http://www.arp-bolts.com/Tech/TechTorque.html

Ive got to ask why anyone would even think seriously of using used stock connecting rods, when you have zero idea as to the cycles they have been thru, in unknown condition with pressed pins that have a core cost of $150-$200
(common cost of a set of 7/16" bbc rods for example)
when NEW ARP 7/16" bolt cap screw 4340 forged rods, are so reasonably priced?
example
http://www.summitracing.com/parts/SCA-26135/

under $300 seems like a much better deal for new rods and 200,000 psi arp rod bolts

http://www.summitracing.com/parts/ESP-63853DL19/
even at $570 these are a bargain
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
grumpyvette

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Re: rod bolts rpm vs stress

Postby grumpyvette » October 22nd, 2011, 4:18 pm

rod bolts can fail for a couple dozen plus reasons
OVER tightening
UNDER tightening
lack of bearing lubrication
lack of rod to block clearance
piston rings locking in the bore when hot
failure to measure stretch or use a torque wrench
detonation damaged pistons
valve train failures
over revving the engine
lack of quench clearance
valve to piston contact
broken valve springs
lack of cam to rod clearance
lack of rod bearing to crank edge clearance
etc. ETC.ETC.

very few are directly related to the rod bolt strength limitations under designed operational conditions, itself failing, most are operator or engine assembly induced problems yet the
"DAMN ROD BOLTS ALWAYS SEEM TO GET THE BLAME"

but the results similar in most cases, below is the result of things coming apart at rpms regardless of the initial cause, it eventually cause the rod bolts to fail
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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
grumpyvette

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Re: rod bolts rpm vs stress

Postby grumpyvette » March 18th, 2012, 9:09 am

" hey grumpy?
Have a 468, BBC steel crank, JE 12:1 ..44cc dome pistons. When I bought these pistons(new) from a speed shop, they also came with a fresh set of 3/8 rods, bushed small ends, arp bolts. The only thing Im a little nervous about, even though they are weight matched, 5 are car rods, 2 are thumb rods and one dimple rod. Again, these are all weight matched and have fresh recondition, new bolts as well as new small end bushings. This will be a street strip car on E85, solid roller, 6500 rpm"



I would not run mixed rod sets of un- known history or mileage when decent low cost new forged rods with 7/16 bolts are so cheap

they may be weight matched for total weight but I can,t see how they are weight matched on both the small and big ends as there are very different in original weight and thickness, I pitch ALL 3/8" big block rods in the nearest dumpster , or sell them dirt cheap to guys that want them and use NEW forged 4340 7/16" rods that are easily 150% stronger in my engine builds simply because by the time you put the required work into them the cost is close too or exceeds the far stronger forged 7/16" rods that they can never equal in strength.
I had 3 complete sets of 454/ 3/8" bbc rods with factory flat top pistons I was never going to use ,I just swapped for an old dual plane intake



http://www.jegs.com/i/Scat/942/26135716/10002/-1

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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
grumpyvette

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Re: rod bolts rpm vs stress

Postby grumpyvette » March 18th, 2012, 9:30 am

ARP posted these pictures of failed bolts


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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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Re: rod bolts rpm vs stress

Postby grumpyvette » March 22nd, 2012, 1:09 pm

http://www.epi-eng.com/piston_engine_te ... basics.htm
WELL WORTH READING THRU

The crankshaft, connecting rods, wristpins and pistons in an engine comprise the mechanism which captures a portion of the energy released by combustion and transforms that energy into useful rotary motion. This page describes the characteristics of the reciprocating motion which the crankshaft and connecting rod assembly imparts to the pistons.

A crankshaft contains two or more centrally-located coaxial cylindrical ("main") journals and one or more offset cylindrical crankpin ("rod") journals. The V8 crankshaft pictured in Figure 1 has five main journals and four rod journals.

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The crankshaft main journals rotate in a set of supporting bearings ("main bearings"), causing the offset rod journals to rotate in a circular path around the main journal centers, the diameter of which is twice the offset of the rod journals. The diameter of that path is the engine "stroke": the distance the piston moves up and down in its cylinder. The big ends of the connecting rods ("conrods") contain bearings ("rod bearings") which ride on the offset rod journals. ( For details on the operation of crankshaft bearings, Click Here; For details on crankshaft design and implementation

The small end of the conrod is attached to the piston by means of a floating cylindrical pin ("wristpin", or in British, "gudgeon pin"). The rotation of the big end of the conrod on the rod journal causes the small end, which is constrained by the piston to be coincident with the cylinder axis, to move the piston up and down the cylinder axis.

Figure 2: TDC

The following description explains the not-so-obvious characteristics of the motion which the crankshaft / conrod mechanism imparts to the piston.

Figure 2 shows a sectional end-view of a crankshaft, connecting rod and piston (CCP) mechanism when the piston is at the furthest extent of its upward (away from the crankshaft) travel, which is known as the top dead center (TDC) position.

The furthest extent of the piston's downward (toward the crankshaft) travel is known as the bottom dead center (BDC) position.

In the CCP mechanism shown, the crankshaft has a 4.000 inch stroke and the center-to-center length of the conrod is 6.100 inches. The rod to stroke ratio (R / S) is the center-to-center length of the conrod divided by the stroke. In this example, the R/S ratio is 6.100 / 4.000 = 1.525.

This ratio is important because it has a large influence on piston motion asymmetry, and on the resulting vibration and balance characteristics, as well as certain performance characteristics, as explained below.

For purposes of this discussion, the extended centerline of the cylinder bore intersects the center of the crankshaft main bearing, and the wristpin is coincident with the cylinder centerline (defined as zero wristpin offset). Although the following descriptions apply strictly to configurations with zero wristpin offset, the general observations apply to nonzero offset configurations as well.

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It is important to understand that the motion of the piston within 90° before and after TDC is not symmetric with the motion within 90° before and after BDC. The rotation of the crankshaft when the crankpin is within 90° of TDC moves the piston substantially more than half the stroke value. Conversely, the rotation of the crankshaft when the crankpin is within 90° of BDC moves the piston substantially less than half the stroke value. This asymmetry of motion is important because it is the source of several interesting properties relating to the operation, performance and longevity of a piston engine.



Figure 3 shows the subject CCP with the crankpin rotated 90° past TDC. Note that the piston has moved over 58% of its total stroke (2.337 inches). That is because in addition to the 2.000" (half-stroke) downward motion of the crankpin (motion projected onto the vertical plane), the crankpin has also moved horizontally outward by 2.000", putting the conrod at an angle with the vertical plane.

The cosine effect of that conrod angle functionally shortens the projected length of the conrod in the vertical plane by 0.337", from 6.100" to the 5.763" shown in the picture. This dynamic "shortening" of the conrod has the effect of adding 0.337" to the 2.000" of downward motion imparted by the crankpin rotation, as illustrated by the two vertical blue lines in Figure 3.

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Now, since the piston has already moved about 58% of the stroke during the first 90° of crank rotation, it stands to reason that during the next 90° of crank rotation (to BDC) the piston will only have to travel the remaining 42% of the stroke to reach BDC, as shown in Figure 4.

The reason is that as the crank rotates toward BDC, the crankpin also moves horizontally back toward the center of the cylinder and "restores" the effective length of the rod. That cosine-effect "lengthening" of the conrod opposes the downward movement of the piston, subtracting 0.337 from the half-stroke of vertical motion produced from 90° to BDC. That effect is illustrated by the lower two vertical blue lines in Figure 4.

Clearly then, when the crankshaft is in any position other than TDC or BDC, the axis of the connecting rod is no longer parallel to the centerline of the cylinder (the line along which the piston, wristpin and small end of the rod are constrained to move). Therefore, the "effective length" of the conrod at any point other than TDC or BDC is the actual conrod center-to-center length multiplied by the cosine of the angle between the rod and the cylinder centerline. It is clear that the dynamic change in the conrod effective length adds to and subtracts from the purely sinusoidal motion caused by crankpin rotation.
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Figure 5: Half-Stroke







Figure 5 shows that, with the R / S ratio in this CCP example (1.525), the half-stroke position of the piston occurs at about 81° crank rotation after TDC. The rapid change in volume of the combustion chamber after the TDC position has some interesting ramifications with respect to the P-V diagram and thermal efficiency (discussed on a different page).


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(Note: If you still believe that installing longer connecting rods will increase an engine's stroke, there's no need for you to go any further on this page, or on this entire site, for that matter.)

PISTON VELOCITY

It is obvious that as the piston moves from TDC to BDC and back, the velocity is constantly changing, and that it is zero at TDC and BDC. Velocity is, by definition, the first feriative of the motion curve, or simply a measurement of how rapidly the motion is changing with respect to the reference (usually time). The value and location of the maximum velocity (the maximum slope of the motion curve) varies directly with engine RPM and are strongly influenced by the R / S ratio.

Figure 6: Maximum Velocity









Figure 6 shows the location of the point of maximum piston velocity, in crankshaft degrees before and after TDC, for the subject CCP. At that position (73.9° before and after TDC), the piston has traveled only 43.9% (1.756") of the total stroke (4.000"). For the configuration used in this example (4-inch stroke, 6.100" rod length, R / S = 1.525), at 4000 RPM, the peak piston velocity is 4390 feet per minute.







Figure 7 shows graphs of piston position and of instantaneous velocity as a function of crankshaft rotation. The blue line ("position") shows piston travel (as a % of stroke) at any point during one rotation of the crankshaft. The blue line is oriented so as to show position in an intuitive sense (top, bottom), therefore the "-" signs shoud be ignored. The green velocity line shows the relative speed of the piston (as a % of maximum) at any point. Velocity with a "plus" sign is motion TOWARD the crankshaft; velocity with a "minus" sign is motion AWAY from the crankshaft.

Note again that at TDC and again at BDC, the piston velocity is zero, because the piston reverses direction at those points, and in order to change direction, the piston must be stopped at some point.

Note also that the position plot (blue) shows that, for this R / S ratio ( 1.525 ), the 50% stroke positions occur at approximately 81° before and after TDC (as illustrated in Figure 5 above). The velocity plot (green line) shows the maximum piston velocities occur at about 74° before and after TDC (as illustrated in Figure 6 above). The velocity line also shows that the piston velocity at any rotation point from TDC up to the maximum velocity is greater than at the same number of degrees before BDC. For example, compare the velocity at 30° after TDC (62%) with the velocity at 30° before BDC (34%).

Piston Travel and Velocity
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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
grumpyvette

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Re: rod bolts rpm vs stress

Postby grumpyvette » March 22nd, 2012, 1:14 pm

http://www.epi-eng.com/piston_engine_te ... basics.htm


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Thursday, March 22nd, 2012
- Piston Motion Basics -
Travel, Velocity, Acceleration, Vibration

The crankshaft, connecting rods, wristpins and pistons in an engine comprise the mechanism which captures a portion of the energy released by combustion and transforms that energy into useful rotary motion. This page describes the characteristics of the reciprocating motion which the crankshaft and connecting rod assembly imparts to the pistons.

A crankshaft contains two or more centrally-located coaxial cylindrical ("main") journals and one or more offset cylindrical crankpin ("rod") journals. The V8 crankshaft pictured in Figure 1 has five main journals and four rod journals.

Figure 1

The crankshaft main journals rotate in a set of supporting bearings ("main bearings"), causing the offset rod journals to rotate in a circular path around the main journal centers, the diameter of which is twice the offset of the rod journals. The diameter of that path is the engine "stroke": the distance the piston moves up and down in its cylinder. The big ends of the connecting rods ("conrods") contain bearings ("rod bearings") which ride on the offset rod journals. ( For details on the operation of crankshaft bearings, Click Here; For details on crankshaft design and implementation, Click Here )

The small end of the conrod is attached to the piston by means of a floating cylindrical pin ("wristpin", or in British, "gudgeon pin"). The rotation of the big end of the conrod on the rod journal causes the small end, which is constrained by the piston to be coincident with the cylinder axis, to move the piston up and down the cylinder axis.

Figure 2: TDC

The following description explains the not-so-obvious characteristics of the motion which the crankshaft / conrod mechanism imparts to the piston.

Figure 2 shows a sectional end-view of a crankshaft, connecting rod and piston (CCP) mechanism when the piston is at the furthest extent of its upward (away from the crankshaft) travel, which is known as the top dead center (TDC) position.

The furthest extent of the piston's downward (toward the crankshaft) travel is known as the bottom dead center (BDC) position.

In the CCP mechanism shown, the crankshaft has a 4.000 inch stroke and the center-to-center length of the conrod is 6.100 inches. The rod to stroke ratio (R / S) is the center-to-center length of the conrod divided by the stroke. In this example, the R/S ratio is 6.100 / 4.000 = 1.525.

This ratio is important because it has a large influence on piston motion asymmetry, and on the resulting vibration and balance characteristics, as well as certain performance characteristics, as explained below.

For purposes of this discussion, the extended centerline of the cylinder bore intersects the center of the crankshaft main bearing, and the wristpin is coincident with the cylinder centerline (defined as zero wristpin offset). Although the following descriptions apply strictly to configurations with zero wristpin offset, the general observations apply to nonzero offset configurations as well.

Figure 3: 90° After TDC

It is important to understand that the motion of the piston within 90° before and after TDC is not symmetric with the motion within 90° before and after BDC. The rotation of the crankshaft when the crankpin is within 90° of TDC moves the piston substantially more than half the stroke value. Conversely, the rotation of the crankshaft when the crankpin is within 90° of BDC moves the piston substantially less than half the stroke value. This asymmetry of motion is important because it is the source of several interesting properties relating to the operation, performance and longevity of a piston engine.



Figure 3 shows the subject CCP with the crankpin rotated 90° past TDC. Note that the piston has moved over 58% of its total stroke (2.337 inches). That is because in addition to the 2.000" (half-stroke) downward motion of the crankpin (motion projected onto the vertical plane), the crankpin has also moved horizontally outward by 2.000", putting the conrod at an angle with the vertical plane.

The cosine effect of that conrod angle functionally shortens the projected length of the conrod in the vertical plane by 0.337", from 6.100" to the 5.763" shown in the picture. This dynamic "shortening" of the conrod has the effect of adding 0.337" to the 2.000" of downward motion imparted by the crankpin rotation, as illustrated by the two vertical blue lines in Figure 3.

Figure 4: 180° After TDC

Now, since the piston has already moved about 58% of the stroke during the first 90° of crank rotation, it stands to reason that during the next 90° of crank rotation (to BDC) the piston will only have to travel the remaining 42% of the stroke to reach BDC, as shown in Figure 4.

The reason is that as the crank rotates toward BDC, the crankpin also moves horizontally back toward the center of the cylinder and "restores" the effective length of the rod. That cosine-effect "lengthening" of the conrod opposes the downward movement of the piston, subtracting 0.337 from the half-stroke of vertical motion produced from 90° to BDC. That effect is illustrated by the lower two vertical blue lines in Figure 4.

Clearly then, when the crankshaft is in any position other than TDC or BDC, the axis of the connecting rod is no longer parallel to the centerline of the cylinder (the line along which the piston, wristpin and small end of the rod are constrained to move). Therefore, the "effective length" of the conrod at any point other than TDC or BDC is the actual conrod center-to-center length multiplied by the cosine of the angle between the rod and the cylinder centerline. It is clear that the dynamic change in the conrod effective length adds to and subtracts from the purely sinusoidal motion caused by crankpin rotation.

Figure 5: Half-Stroke







Figure 5 shows that, with the R / S ratio in this CCP example (1.525), the half-stroke position of the piston occurs at about 81° crank rotation after TDC. The rapid change in volume of the combustion chamber after the TDC position has some interesting ramifications with respect to the P-V diagram and thermal efficiency (discussed on a different page).





(Note: If you still believe that installing longer connecting rods will increase an engine's stroke, there's no need for you to go any further on this page, or on this entire site, for that matter.)






PISTON VELOCITY

It is obvious that as the piston moves from TDC to BDC and back, the velocity is constantly changing, and that it is zero at TDC and BDC. Velocity is, by definition, the first feriative of the motion curve, or simply a measurement of how rapidly the motion is changing with respect to the reference (usually time). The value and location of the maximum velocity (the maximum slope of the motion curve) varies directly with engine RPM and are strongly influenced by the R / S ratio.

Figure 6: Maximum Velocity









Figure 6 shows the location of the point of maximum piston velocity, in crankshaft degrees before and after TDC, for the subject CCP. At that position (73.9° before and after TDC), the piston has traveled only 43.9% (1.756") of the total stroke (4.000"). For the configuration used in this example (4-inch stroke, 6.100" rod length, R / S = 1.525), at 4000 RPM, the peak piston velocity is 4390 feet per minute.







Figure 7 shows graphs of piston position and of instantaneous velocity as a function of crankshaft rotation. The blue line ("position") shows piston travel (as a % of stroke) at any point during one rotation of the crankshaft. The blue line is oriented so as to show position in an intuitive sense (top, bottom), therefore the "-" signs shoud be ignored. The green velocity line shows the relative speed of the piston (as a % of maximum) at any point. Velocity with a "plus" sign is motion TOWARD the crankshaft; velocity with a "minus" sign is motion AWAY from the crankshaft.

Note again that at TDC and again at BDC, the piston velocity is zero, because the piston reverses direction at those points, and in order to change direction, the piston must be stopped at some point.

Note also that the position plot (blue) shows that, for this R / S ratio ( 1.525 ), the 50% stroke positions occur at approximately 81° before and after TDC (as illustrated in Figure 5 above). The velocity plot (green line) shows the maximum piston velocities occur at about 74° before and after TDC (as illustrated in Figure 6 above). The velocity line also shows that the piston velocity at any rotation point from TDC up to the maximum velocity is greater than at the same number of degrees before BDC. For example, compare the velocity at 30° after TDC (62%) with the velocity at 30° before BDC (34%).

Piston Travel and Velocity

Figure 7

The profile of the velocity curve, and therefore the location of the maximum velocity, are influenced by the R / S ratio. As the rod gets shorter with respect to stroke (a smaller R / S ratio), two interesting things happen which can have important effects on cylinder filling: (1) the point of maximum piston velocity moves closer to TDC, and (2) the piston moves away from TDC faster, creating a stronger intake pulse. The location of maximum piston velocity influences the design of camshaft lobe profiles (especially intake) in order to optimize the intake event in a particular speed range, and can have an influence on the intake characteristics with regard to the strength and shaping of the intake pulse for ram tuning.
MEAN PISTON SPEED

There is another piston velocity which is used more as a "rule-of-thumb" in engine evaluations. It is called "mean piston speed", which is a calculated value showing the average velocity of a piston at a known RPM in an engine having a known stroke length.

Keeping in mind that every crankshaft revolution, the piston travels a distance equal to twice the stroke length, then Mean Piston Speed (MPS) is calculated by:

MPS (ft per minute) = RPM x 2 x stroke (inches) / 12 (inches per foot) = RPM x stroke / 6

The Mean Piston Speed at 4000 RPM for the example 4.000 inch stroke engine is:

MPS (ft per minute) = 4000 x 4 / 6 = 2667 feet per minute.

For purposes of rules of thumb, it is generally agreed that for an engine in aircraft service, 3000 fpm is a comfortable maximum MPS and experience has shown that engines having an MPS substantially exceeding that value have experienced reliability issues. Note that R / S has no influence on MPS, although it strongly affects peak piston speed (4390 fpm for the example engine {R / S = 1.525} at 4000 RPM).
PISTON ACCELERATION

The force it takes to accelerate an object is proportional to the weight of the object times the acceleration. From that it is clear that piston acceleration is important because many of the significant forces exerted on the pistons, wristpins, connecting rods, crankshaft, bearings, and block are directly related to piston acceleration. Piston acceleration is also the main source of external vibration produced by an engine. (Torsional vibration is discussed separately on another page.)

Acceleration is, by definition, the first derivative of the velocity curve, or in other words, the slope of the velocity curve at any given point along the reference. More simply, it is a measure of how rapidly velocity is changing, usually expressed with reference to time. If velocity does not change with respect to the reference, there is no acceleration. Conversely, if velocity changes very rapidly with respect to the reference, there is a large acceleration. (See Velocity and Acceleration for a more thorough explanation.)

It is clear from Figure 7 that the piston velocity is constantly changing with respect to a constant value of crankshaft rotation. Therefore, In order to move from the zero-velocity point (TDC) to the maximum velocity point, the piston must be subjected to a large acceleration function which varies with the angular rotation of the crankshaft.

Figure 8 shows the acceleration, velocity and position plots for the example CCP under discussion. (All numeric values presented are for the R / S in this example.)

Image

The maximum positive value of acceleration (100%) occurs at TDC. Between TDC and maximum piston velocitty (74° in this case), acceleration is positive but decreasing toward zero (the piston velocity is still increasing but less rapidly). At maximum piston velocity (74° at this R / S ), the piston stops speeding up and begins to slow down. At that point, the acceleration changes direction (from a "plus" number to a "minus" number), and in so doing, momentarily passes through zero.

At this R / S, the maximum negative acceleration does not occur at BDC, but about 40° either side of BDC. The value of this maximum negative acceleration is only about 53% of the maximum positive acceleration seen at TDC. The acceleration at BDC is only 49% of the TDC maximum. The acceleration from max piston velocity (74°) to BDC is negative, and that acceleration is slowing the piston to zero velocity. Therefore, it might be (incorrectly) called deceleration. However, that same negative acceleration is applied to the piston after BDC and is causing its velocity to increase.

The zero acceleration point occurs (by definition) at the point of maximum piston velocity (74° B/A TDC), where velocity is reversing direction, but the rate of change of velocity (the slope of the curve) is zero.

The somewhat odd shape at the bottom of the total piston acceleration (magenta) curve is the result of the fact that the total piston acceleration is the sum of several orders of acceleration, the first two being the most significant. The two major orders which combine to produce this total acceleration profile are important because they can produce significant vibration challenges to the engine designer (covered in Crankshafts).

Figure 8 shows the same total piston acceleration curve (magenta line) shown in Figure 7, along with the two significant orders of piston accelerations which combine to produce that curve. The total piston acceleration curve (magenta) is the sum of the two separate accelerationorders: primary (blue) and secondary (green).
Image

As explained in Piston Motion above, the piston motion in the first 90° of rotation consists of the sum of the effect of the half-stroke motion of the crankpin projected onto the vertical plane (2.000") and the effect of the apparent 0.337" "shortening" of the rod length projected onto the vertical plane. The second 90° of rotation also produces a half-stroke motion in the vertical plane, but the cosine-effect lengthening of the conrod in the vertical plane produces 0.337" motion which subtracts from the half-stroke.

The primary acceleration (blue line) is the result of the piston motion produced by the component of crankpin movement projected onto the vertical plane. This curve is a sinusoid which repeats once per revolution of the crankshaft (first order) and comprises the majority of the acceleration. Note that the primary acceleration curve crosses zero at the 90° rotation points and peaks at TDC and BDC.

The secondary acceleration (green line) is the result of the additional piston motion caused by the cosine-effect dynamic length-change of the conrod. This motion adds to the piston movement between TDC and the max velocity point and subtracts from the piston motion between the max velocity point and BDC. This curve is also sinusoidal and repeats twice per crankshaft rotation (second order) and crosses zero at the 45°, 135°, 225° and 315° rotation points. The total piston acceleration at any point is the sum of the values of the primary and secondary acceleration curves.

Contemporary piston engines tend to have R / S ratios in an approximate range of 1.5 to 2.0. Note that a rod / stroke ratio less than 1.3 is, for practical applications, not possible due to physical constraints such as the need for piston rings and a wristpin, sufficient piston skirt length, and the inconvenience of having the piston contact the crankshaft counterweight, not to mention the excessive side load such a small ratio would produce.

Here are two practical examples comparing the effects of R / S on acceleration and velocity. In a Lycoming IO-360 (and IO-540) the rod length is 6.75" and the stroke is 4.375", for a ratio of 1.543, close to the low end of the spectrum in contemporary design. At the other end of that spectrum, the connecting rod on a typical (circa 2007) 2.4-liter Formula-1 V8 engine is about 4.010" long, what your average race-engine mechanic would call a "very short rod". The stroke in the F1 engine is in the vicinity of 1.566", which produces a very large R / S ratio of 2.56. The following graph (Figure 9) clearly shows the effect of large and small R / S ratios.

Image

t is clear that the engine with the very small R / S ratio of 1.543 (the "long" 6.75 inch conrod, the blue velocity and acceleration curves) has a substantially higher peak acceleration (10%), a higher secondary acceleration, a higher peak velocity (3%), an earlier velocity peak (5 crank degrees) and the distinct acceleration reversal around BDC, confirming the substantial secondary vibration component.

Compare that to the large 2.56 R / S ratio (the "short" 4.01 inch conrod) magenta curves, showing a substantially lower peak acceleration (10%), a lower secondary acceleration, a later and slightly lower (3%) peak velocity, and the total acceleration curve is closer to symmetric, confirming the substantially-reduced secondary vibration component. Figure 9 also clearly demonstrates the absurdity of discussing conrod length as an absolute.

Figure 10 is a chart listing the main effects of R / S ratios varying from 1.40 to 2.55, with the reference point for Vmax %, PPA max-pos %, and PPA max-neg % being based on an R / S ratio of 2.00, since that ratio is the first at which the maximum negative acceleration occurs at BDC. Notice that at ratios beyond 2.00, the acceleration curve becomes more symmetric, but the peak velocity does not change much at all.
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Re: rod bolts rpm vs stress

Postby grumpyvette » March 1st, 2013, 7:12 pm

540 RAT posted this info

Default Rod bolt strength, what do we REALLY NEED?
This write-up is not intended to be a chapter out of an Engineering Design Book. That would be way too long, way too involved, and way too boring for most folks here to have any interest in. Instead, this is just a general overview of how connecting rod bolts compare, and what we REALLY NEED in our motors.

Yield Strength = the stress at which a material begins to deform plastically. Prior to the yield point the material will deform elastically and will return to its original shape and size when the applied stress is removed. Once the yield point is passed, the deformation will be permanent, which is considered a “failed” condition for a bolt. So, the bolt must be discarded.

Tensile Strength = the maximum stress that a material can withstand while being stretched or pulled, without starting to neck down and ultimately breaking.

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First let’s look at some typical strength values of various bolts, to get a general feel for how they compare.


Grade 2 hardware store general purpose bolt:
Yield strength = 55,000 psi
Tensile strength = 74,000 psi
Cost = a few cents each
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Grade 5 hardware store general purpose bolt:
Yield strength = 85,000 psi
Tensile strength = 120,000 psi
Cost = a few cents each
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Grade 8 hardware store general purpose bolt:
Yield strength = 120,000 psi
Tensile strength = 150,000 psi
Cost = a few cents each
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ARP 8740 chrome moly “connecting rod” bolt:
Yield strength = 180,000 psi
Tensile strength = 200,000 psi
Cost = $120.00 per set of 16 at Summit Racing Equipment, or about $8.00 each.
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ARP 2000 “connecting rod” bolt:
Yield strength = 180,000 psi
Tensile strength = 220,000 psi
ARP 2000 rod bolt material has twice the fatigue life of 8740 chrome moly rod bolt material.
Cost = $200.00 per set of 16 at Summit Racing Equipment, or about $13.00 each.
---------
ARP L19 “connecting rod” bolt:
Yield strength = 200,000 psi
Tensile strength = 260,000 psi
ARP L19 rod bolt material is subject to hydrogen embrittlement, and stress corrosion. It also cannot be exposed to any moisture, including sweat and/or condensation.
Cost = $200.00 per set of 16 at Summit Racing Equipment, or about $13.00 each.
---------
ARP Custom Age 625+ “connecting rod” bolt:
Yield strength = 235,000 psi
Tensile strength = 260,000 psi
ARP Custom Age 625+ rod bolt material has nearly 3 ½ times the fatigue life of the ARP 3.5 rod bolt material.
Cost = $600.00 per set of 16 at Summit Racing Equipment, or about $38.00 "EACH".
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ARP 3.5 “connecting rod” bolt:
Yield strength = 220,000 psi
Tensile strength = 260,000 psi
Cost = $855.00 per set of 16 at Summit Racing Equipment, or about $53.00 "EACH"!!!

So, as you can see above, hardware store general purpose bolts are considerably weaker than “purpose built” connecting rod bolts. And we won’t even bother getting into the differences in fatigue life. Suffice it to say, we CANNOT use general purpose hardware store bolts in our connecting rods.

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A connecting rod bolt’s maximum tension loads are determined by the mass of the parts involved, the rod length, the stroke length, and the max rpm. That’s it. It has absolutely nothing what so ever to do with the amount of HP being made. The max tension loads on the rod bolts will never change, no matter if you add Nitrous, a Turbo, or a Blower to an engine, as long as the short block and redline don’t change. That max tension loading occurs at TDC on the exhaust stroke as the mass involved is brought to a dead stop, and has its direction reversed. In order to change the max tension loading on the rod bolts, you’d have to change the short block configuration and/or the redline. And vacuum pulling on the rod bolts when chopping the throttle at high rpm, is not a concern. Because those affects don't even begin to build until well past TDC, which of course is "AFTER" the mass of the parts involved has already been brought to a stop, and their direction reversed.

The rod’s big end “clamp-up preload” provided by stretching/torquing the rod bolts, must always be HIGHER than the “cyclic tension load” applied to the bolts at TDC exhaust, in order to prevent rod bolt failure. And the larger the difference between the preload and the cyclic load, the better. Precision detailed "Strength Analysis" calculations can be performed using sound Engineering principles, to determine the “Margin of Safety” (MOS) between the “cyclic tension loading” and the “clamp-up preload”, to make sure you have a sufficient MOS for the engine to be reliable. I’ll spare you all the involved and complicated math, and just show you the results.

Before we go on, first a comment on “cap screw” rod bolt sizes. Your rod bolts are NOT the size you think they are. If you run 3/8” rod bolts, only the threads are 3/8”. But, the part of the bolt that matters regarding the stretch, is the shank. And the main length of the shank is only 5/16”, not the 3/8” you might have thought. And if you run 7/16” rod bolts, the threads are 7/16”, but main length of the shank is only 3/8”. So, where you are most concerned, the bolts are one size SMALLER than you thought.

And if that isn’t enough detail, you must also consider, in addition to the main section of the shank, the other diameters involved which come from the radius transition between the threads and the shank, the radius transition between the shank and the shoulder right under the bolt head flange, and that shoulder itself right under the bolt head flange. The bolts stretch the whole length between the threads and the bolt head flange. And all those individual sections contribute to the total stretch by different amounts.

So, the rod bolt “Strength Analysis” must take into account all those various diameters, as well as the length of each of those diameters. Because the stretch has to be calculated for each individual section of the shank between the threads and the bolt head flange. If this is not done correctly, the “Strength Analysis” results will simply end up being wrong and worthless. But, for the results shown below, all those details were carefully worked out for those portions of the “Strength Analysis”. So, the answers below are all accurate.

Rod bolt "Strength Analysis" performed on known real world Street Hotrods, Street/Strip cars and Sportsman Drag cars, being operated at their typical maximum rpm, indicates the following:

• An engine with a max rpm rod bolt MOS of around 125% or higher, results in the engine being as safe and reliable as a stock grocery getter, or in other words essentially bullet proof. This is our design target when planning a new build. Having a MOS higher than this can’t hurt of course, but in terms of strength requirements, there is really no added value for doing that. However, a higher MOS can help with rod bolt fatigue life, if that is critical for a particular application. More on fatigue life later.

• If you are a little more aggressive, and run a maximum rpm rod bolt MOS between 100% and 125% only “on occasion”, which limits the number of cycles at this higher stress level, you will still generally be able to keep the engine together.

• But, if you were to run a typical maximum rpm rod bolt MOS under 100%, your rod bolts will be expected to fail prematurely.

As mentioned above in the definition of Yield Strength, we CANNOT stretch our rod bolts beyond the yield point. Because once the yield point is passed, it is considered a “failed” condition for a bolt, and the bolt must be discarded. So, a typical conservative Engineering approach in most general applications is to use a preload clamp-up of about 75% of yield. That way you have a good range between the installed preload and the yield point, in case the bolts get stressed even more during operational use. However, typical engine connecting rod bolt preload clamp-up in most reliable engines, can vary from a low of about 60% of yield to a high of about 90% of yield, with 75% of yield, the sweet spot you might say, right in the middle.

Since rod bolt stretch specs have generally become the standard in High Performance engine builds, the stretch called for is more often around 90% of the yield point. Stretching to this higher percentage of yield, is used to maximize preload clamp-up, in an effort to try and help minimize rod big end distortion at high rpm, which can cause additional undesirable rod bolt bending that would add to the bolt stress.

So, this high level of stretch is a good idea from that standpoint, but at the same time, you are left with a smaller range between the installed preload clamp-up and the yield point. But, this common 90% of yield has worked out quite well in the real world for Hotrods, Street/Strip cars, and Sportsman Drag cars. Even though there is less range between the installed preload clamp-up and the yield point, the yield point in properly selected rod bolts is not typically reached in actual operation, so all is good.

You may also have noticed that through all this discussion of rod bolt strength, there has been no mention at all of rod bolt tensile strength. That’s because we CANNOT go beyond the yield strength which is reached well “BELOW” the tensile strength. So, what good is tensile strength then? For a large number of steels, there is a direct correlation between tensile strength and fatigue life. Normally, as tensile strength increases, the fatigue life increases. So, while tensile strength does not come into play during rod bolt "Strength Analysis", it is a factor in rod bolt fatigue life.

Rod bolt fatigue life is important to Road Racers because of the number of cycles they see. And rod bolt fatigue life is absolutely critical for Endurance Racers like NASCAR. And NASCAR teams do an incredible job managing the fatigue life of their rod bolts. But, for our Hotrods, Street/Strip cars and Sportsman Drag cars, rod bolt fatigue life isn’t typically a big concern, if the motors are built with the correct rod bolts in the first place. That is because these bolts won’t typically see enough cycles in their lifetime to cause a failure due to fatigue. But, with that said, it is still a good idea to keep fatigue life in the back of your mind, when it comes to choosing your rod bolts. It can be a tie breaker, in the event that multiple rod bolts are being considered for a certain build. More on that below.

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Even though there are various companies that offer rod bolts, below we will compare 5 different rod bolts offered by Industry leader ARP.

So, let’s take a look at a typical 540ci BBC motor, running steel rods with 7/16 “cap screw” rod bolts, and uses 7,500 rpm as its typical maximum, which results in a cyclic tension load on each rod bolt that = 7,280 lbs or about 3.6 tons:

• For general reference, let’s first take a look at rods installed the old school traditional way, here using ARP 2000 rod bolts that are torqued to about 75 ft lbs with original ARP moly lube.
Bolt stretch is about .005”, which = 76% of yield strength
Clamp-up preload on each rod bolt = 16,531 lbs or about 8.3 tons
Margin of Safety (MOS) for this setup = 127%, which meets our MOS design target for being safe, reliable and essentially bullet proof.

Now, for the rest of the rod bolts we’ll be looking at, we’ll preload them to the more common higher percentage of yield strength, which is typical of the stretch called for these days.

• Using ARP 8740 chrome moly rod bolts (this has the same yield strength as ARP 2000)
Bolt stretch = .006” which = 90% of yield strength
Clamp-up preload on each rod bolt = 19,686 lbs or about 9.8 tons
Margin of Safety (MOS) = 170%

• Using ARP 2000 rod bolts (this has the same yield strength as 8740 chrome moly)
Bolt stretch = .006” which = 90% of yield strength
Clamp-up preload on each rod bolt = 19,686 lbs or about 9.8 tons
Margin of Safety (MOS) = 170%

• Using ARP L19 rod bolts
Bolt stretch = .0066” which = 90% of yield strength
Clamp-up preload on each rod bolt = 21,655 lbs or about 10.8 tons
Margin of Safety (MOS) = 197%

• Using ARP Custom Age 625+ rod bolts
Bolt stretch = .0078” which = 90% of yield strength
Clamp-up preload on each rod bolt = 25,445 lbs or about 12.7 tons
Margin of Safety (MOS) = 250%

• Using ARP 3.5 rod bolts
Bolt stretch = .0073” which = 90% of yield strength
Clamp-up preload on each rod bolt = 23,821 lbs or about 11.9 tons
Margin of Safety (MOS) = 227%

As you can see above in all 6 examples, whether torqued the traditional way to a lower stretch value, or stretched to the more recently called for higher percentage of yield value, all these rod bolts are above the minimum 125% MOS target for safety and reliability. Therefore, all these configurations would operate without issue, just like a stock grocery getter. So, if a builder chooses any of these bolts or stretch values between the 127% and the 250% "Margins of Safety" above, he could NOT go wrong, no matter how much HP the motor makes. Remember that HP has NOTHING to do with the max tension loads on rod bolts.

Since most Hotrods, Street/Strip cars, and Sportsman Drag cars, with their lower number of cycles, can live almost indefinitely with some of the more affordable mainstream rod bolts above, it’s rather hard to make a case for using the much more expensive and higher strength 625+ or 3.5 bolts, even if they do have higher fatigue life values.

BOTTOM LINE

So then, all we REALLY NEED, from a conservative Engineering standpoint, is to at least reach the 125% MOS target for safety and reliability, no matter how much HP is being made. And anything above that 125% is fine, but not necessary.

----------------

But, things aren’t always wine and roses, because some engines will NOT stay together and live like the well built configurations above. I've done "failed" rod bolt "Strength Analysis" on two smaller very high revving engines, after the fact, to take a look at why they failed. One blew-up catastrophically when a rod bolt broke, costing its owner 20 grand. And the other engine was found to have rod bolts stretched beyond the yield point, during a teardown for other reasons. So, its fuse had been lit, but fortunately it was caught just in the nick of time before they let go, saving its owner a ton of money and agony.

In both cases, the rod bolt "Strength Analysis" revealed that they had been built wrong, and that they were well BELOW 100% MOS, which predicts premature rod bolt failure. One had only a 67% MOS and the other had only an 86% MOS. If rod bolt "Strength Analysis" had been performed before these engines were built, during the planning stages, then all that grief and cost could have been avoided. They have since been rebuilt much stronger, with MOS values now well ABOVE that 125% safe target. And they have now been raced for some time without issue.

----------------

SUMMARY

• ARP 8740 chrome moly rod bolt - a strong affordable rod bolt, but it has only a moderate fatigue life, which makes the ARP 2000 rod bolt which is in the same general price range, a much better choice since it has twice the fatigue life.

• ARP 2000 rod bolt - considering how good its strength and fatigue life are, this rod bolt is an excellent choice for most Hotrods, Street/Strip cars, and Sportsman Drag cars.

• ARP L19 rod bolt - the strength and fatigue life increases this bolt provides over the ARP 2000 are not significant enough to overcome the concerns the L19 has with hydrogen embrittlement, stress corrosion, and the fact that it CANNOT be exposed to any moisture, including sweat and/or condensation. Don’t forget that every engine forms condensation inside, at every cold start-up. Plus, oil rises to the top of, and floats on water because of density differences, which can leave portions of the rod bolts exposed to water even after the engine is built. Therefore, it is best to avoid the L19 rod bolt altogether, especially since the ARP 2000 rod bolt already provides way more than enough strength and fatigue life than is typically required by most Hotrods, Street/Strip cars, and Sportsman Drag cars. So, there simply is no good reason to select the ARP L19 rod bolt. If you are currently running L19 bolts, I’d suggest you consider replacing them with different bolts the next time you have the motor apart.

• ARP Custom Age 625+ rod bolt - a very pricey bolt, but with its excellent strength and its impressive fatigue life, this bolt is one of the very best rod bolts on the market.

• ARP 3.5 rod bolt - this bolt has excellent strength, but its staggering cost is 43% HIGHER than the 625+ bolt, yet the 625+ bolt is superior to the 3.5 bolt in virtually every way. So, there is no good reason to select the 3.5 bolt either.

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CONCLUSION and RECOMMENDATION

Of the 5 rod bolts above:

• The ARP 2000 rod bolt is an excellent value, considering how good its strength and fatigue life are. And it should be considered the rod bolt of choice for most Hotrods, Street/Strip cars, and Sportsman Drag cars, no matter how much HP they make. And this is why you most often see quality aftermarket rods come with these bolts.

• ARP Custom Age 625+ rod bolt has a price that is not for the faint of wallet, but it should be considered the rod bolt of choice for very high revving engines, road race engines, and endurance engines, which require the utmost in rod bolt strength and/or fatigue life.

540 RAT

Member SAE (Society of Automotive Engineers) "
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Re: rod bolts rpm vs stress

Postby grumpyvette » March 2nd, 2013, 1:45 pm

MEASURING CONNECTING ROD BOLT STRETCH

http://www.carcraft.com/techarticles/cc ... ewall.html

http://www.eaglerod.com/mosmodule/bolt_torque.html

by Mike Mavrigian
http://www.precisionenginetech.com/tech ... ch-part-1/
Image

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Tightening connecting rod bolts while measuring bolt stretch provides a much more accurate method of achieving proper bolt preload and clamping force.

Professional performance and race engine builders have long realized that the correct approach to tightening connecting rod bolts is to stress the bolts into their “working” range of elasticity, but not beyond. Since OEM connecting rod bolts may vary in terms of their ideal torque by as much as 10 lbs./ft. from batch to batch due to variations in heat treating and materials, if the concern is to arrive at both peak bolt strength as well as maintain concentricity of the rod big-end, the rod bolts should be measured for stretch instead of simply tightening until the torque wrench hits its mark.

In simple terms, in order to measure bolt stretch, first measure the total rod bolt length (from the head surface to the tip of the shank) in the bolt’s relaxed state. Then monitor bolt length as you continue to tighten until the specified amount of bolt stretch has been achieved.

The difference in length indicates the amount of stretch the bolt experiences in its installed state. For the majority of production rod bolts, stretch will likely be in the 0.005″ to 0.006″ range (this is a generalization; always follow the specific stretch specified by the bolt maker). If it’s difficult to achieve enough stretch, the bolt is probably experiencing too much friction that is preventing the proper stretch (requiring lubricant on the threads). If stretch is excessive, the bolt may have been pulled beyond its yield point and is no longer serviceable.

While an outside micrometer may be used to measure the rod bolt length, the most accurate method is to use a specialty fixture that is outfitted with a dial indicator. Excellent examples of this gauge include units from ARP, GearHead Tools and Goodson Shop Supplies. GearHead’s bolt stretch gauge features a heat-treated aluminum frame (with a very handy thumbhole) with a specially modified dial indicator with sufficient spring tension to hold the gauge firmly to the ends of the rod bolt. The indicator can be rotated for right- or left-hand operation and the lower anvil is adjustable to accommodate various bolt lengths.

Goodson Shop Supplies also offers a rod bolt stretch gauge (P/N RBG-4) featuring spherical points for consistent and repeatable readings, and can also be rotated for right or left hand operation. Also, ARP offers its own bolt stretch gauge (P/N 100-9941) designed with 0.0005″ increments, with a heavy spring and ball tips.
There is a debate among some engine builders regarding the validity of measuring rod bolt stretch, due to potential compression of the rod material as the rod cap is clamped to the rod. While this may be a consideration, the use of a stretch gauge remains the best and most practical method of accurately determining bolt load.

Connecting rod bolts can be viewed as high-tensile springs. The bolt must be stretched short of its yield point in order for accurate and, most importantly, repeatable clamping of the rod cap to the rod. Improper or unequal bolt clamping force can easily result in a non-round rod bore.

Stock, or production, rod bolts typically offer a tensile strength of approximately 150,000-160,000 psi. However, due to variances in bolt production, tolerances can be quite extreme with peak bolt stretch occurring anywhere from, say 0.003″ to 0.006″. If the installer uses only torque in the attempt to achieve bolt stretch, he runs the risk of unequal rod bolt clamping loads due to the potential inconsistencies between bolts.

High-performance rod bolts are manufactured to much tighter tensile strength tolerances. ARP, for instance, calculates each and every rod bolt for stretch, and the bolt packages include reference data to that effect. The instructions actually recommend that a specific amount of bolt stretch should be achieved on each bolt (ARP cites 190,000 psi as its nominal, or base tensile rating, with actual ratings much higher in some applications).

How can unequal/inadequate rod bolt tightness affect the connecting rod big end bore shape? Let’s cite an example: If one technician reconditions the connecting rods using torque value alone to tighten the rod bolts and another technician who is responsible for final assembly uses the bolt stretch method, the final result can be out-of-round bores. This is because of frictional variances that will be encountered. As a result, the assembler using the stretch method may achieve a higher clamping load on one or more bolts as compared to the loads imposed when the rod reconditioner torqued the nuts without regard to actual bolt stretch. When a bolt is tightened with dry threads, as much as 80 percent of the torque can be exerted because of friction as opposed to bolt stretch.

In a high-volume production rebuilding facility, technicians may not have the time to measure for bolt stretch. However, a slower-paced operation that is attempting to obtain maximum accuracy (for a race engine, as an example), is far better off using the stretch method instead of relying only on the torque method.

A set of connecting rod bolts’ instructions may list both a torque value and a stretch range, effectively giving you a choice of methods. Yes, tightening only to a specified torque value is quicker and measuring bolt stretch requires more time, but the best results will be achieved by measuring bolt stretch. So, unless you’re in a rush, take the time to measure stretch, tightening each rod bolt to the recommended stretch range. It’s all about the quest for precision.
RECOGNIZING AND UNDERSTANDING THE FRICTION FACTOR

Friction (underhead and thread area contact) is a variable that is difficult to control precisely. The best way to avoid friction variables is to use the stretch method when tightening rod bolts. The stretch method allows you to accurately control the all-important bolt preload, independent of friction.

Each time a bolt is tightened (to value) and loosened, the friction factor is reduced as the mating surfaces (threads and underhead) “wear” in. Eventually the friction becomes fairly constant during future tightening.

Considering this “evening-out” of friction, when installing a new rod bolt where the stretch method cannot be used (because of available space for the stretch gauge, etc.), the bolt should be tightened and loosened several times before trying to achieve final torque value. While the number of tightening/loosening cycles depends on the lubricant being used, ARP recommends when using its lubricant, five loosening/tightening cycles is sufficient.

If a bolt is tightened using straight torque, you may not necessarily achieve the desired pre-load due to the variable of friction. Since we can’t predict the frictional loss, measuring rod bolt stretch provides the most accurate method of ensuring that the clamping loads will be both sufficient for the task and that each pair of rod bolts will achieve equal loads.

Bolt stretch is generated by a number of factors, including tensile strength and mass (the length of the bolt being stretched). The effective diameter of the bolt contributes to this. For example, let’s consider two 3/8″ x 1″ bolts. One features a 1″ long shank with threads on the full 1″ of the shank length. The other bolt features 3/4″ of shank length that is full-diameter and smooth, and only 1/4″ of thread length at the tip. The bolt with partial thread will stretch less because the shank area between the head and nut engagement area has a thicker cross section. The partial-thread bolt will have a .375″ diameter shank, while the all-thread bolt will have only a .324″ shank (due to the smaller root diameter inside the thread path).

ARP, to cite one example, calculates the stretch number for every bolt. On the spec sheet that is included with every bolt set, this stretch goal is listed, in addition to a torque value, but the torque value should be used as a guide only. ARP does not want the installer to use a torque value as the final indication of bolt stretch. Rather, the bolts should be individually measured for stretch, to assure that each bolt is installed at its optimum strength.

While we cannot control the reaction of the connecting rod base material, at least consider the potential compression of the connecting rod material itself during bolt clamping. As the bolt is tightened, the head of the bolt will tend to embed itself into the rod, slightly compressing the stock material of the connecting rod under the bolt head. Production rods are typically softer, allowing the head of the rod bolt to sink into rod until the material under the bolt heads “work hardens” under compression. ARP recommends that the bolt stretch is based on the bolt itself and not on the compression of the rod since we can’t accurately predict what the rod does in this state.

Since too many variables exist in terms of rod bolts and connecting rods, we can’t draw any generalized conclusions regarding ideal connecting rod bolt stretch. However, to use the Chevy smallblock 350 as but one isolated example, ARP typically looks for an installed bolt stretch of .0063″. Since each engine/rod/bolt application differs, we cannot assume that ideal bolt stretch would be the same for any given application.
ROD BOLT STRESS RISERS

Fatigue failure is frequently caused by local stress risers, such as sharp corners. In bolts, this would correspond to the notch effect associated with the thread form. It is well known that maximum stress in an engaged bolt occurs in the last engaged thread. By removing the remaining non-engaged threads, the local notch effect can be reduced. This leads to the standard configuration used in most ARP rod bolts-a reduced diameter shank and full engagement for the remaining threads. Providing a local fillet radius at the location of the maximum stress further reduces the local notch effect.

The reduced shank diameter also reduces the bending stiffness of the bolt. When the bolt bends due to deformation of the connecting rod, the bending stresses are reduced below what they would otherwise be. This further increases the fatigue resistance of the bolt.

The direct reciprocating load is not the only source of stresses in bolts. A secondary effect arises because of the flexibility of the rod big end. The reciprocating load causes bending deformation of the bolted joint (rod to rod cap). This deformation causes bending stresses in the bolt as well as in the rod. These bending stresses fluctuate from zero to their maximum level with each crankshaft revolution.
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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