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PostPosted: October 3rd, 2008, 7:27 pm
by grumpyvette
youll generally find that 4000 FPM (FEET PER MINUTE)of piston speed with stock parts or 4500fpm with VERY good quality forged parts is the reasonable limit on the lower end ROTATING ASSEMBLY stress, but the valve control issues tend to become the limiting factor before the lower assembly causes problems,assuming you've used good quality parts and balanced the rotating assembly,and your avoiding lubrication and detonation issues,as a general rule you'll have sufficient strength in a 2 bolt block to run at about 4000 ft per minute in piston speed,over a long life expectancy, exceed that stress level by much and your increase stress will eventually cause durability issues
its commonly the valve train,control issues, the lubrication and cooling systems failures or detonation in the cylinders that are common factors in engine failures
keep in mind stress on the rods and main caps goes up rapidly with increased rpms
the aftermarket forged 4340 steel connecting rods with the 7/16" ARP cap screw connecting rod bolts, has about a 20% larger cross sectional area , and in many cases a 100% or greater strength advantage more than the factory connecting rods with 3/8" rod bolts, its generally not needed till you exceed about 4200fpm in piston speed, but it seldom is a bad idea to use 4340 forged connecting rods with the 7/16" ARP rod bolts, as they are not that expensive, and your engine durability depends on good rods

a good set of SCAT FORGED 4340 forged connecting rods costs less than $400 and they are 150%-200% stronger than MOST OEM chevy SBC rods
it will cost you almost that much to replace the bolts with ARP wave lock bolts, balance and polish and resize stock rods and you have far weaker rods when your done
keep in mind most posted info is meant to give good guide lines and related info,theres no way of pin pointing the the exact rpm that your particular components will fail, but experience has shown that having a decent cushion in the component strength is a very good idea.
rarely will I get specific,s to one particular build unless it states that, now your combo, your building may have forged rods and 3/8" ARP rod bolts that are significantly stronger than stock OEM 3/8" rod bolts, and forged rods that are much stronger than the OEM powdered metal connecting rods,and it may do fine up into the 6500rpm range for very brief and fairly infrequent use,and Id think that anything under about 6300rpm for only brief periods should be safe with a decent forged rod with ARP rod bolts in a 383 combo, but it certainly won,t allow you to spin the engine at near 6500 rpm for prolonged duration, if only because the valve train most guys use in a street build 383 is unlikely to work well under those conditions, especially with any hydraulic lifter valve train..
a 7/16" rod bolt is about 20% larger in cross sectional area than a 3/8" rod bolt and most 7/16" rod bolt rods are significantly thicker in cross section making them much stiffer and less likely to fail under loads


ARP posted these pictures of failed bolts


Image ... ewall.html


lets do a bit of math with a high rpm 383 combo, it might help here

lets take this connecting rod (645 grams)
this piston (527 grams)
and just temporarily ignore the rings,and bearing weight

thats about 18210 grains at 4500 fpm in piston speed thats 75 ft per second
6588 inertial pounds the piston weight per piston at just over 7000rpm, and your looking to reverse its direction of travel , at over 116 times PER SECOND at 7000 rpm, effectively doubling even that load of the stress on the exhaust stroke ,if you don,t think thats absolutely amazing that its potentially possible to do without instantly self destructing you have zero grasp on the potential levels of stress, then we add the fact that theres potentially 600 psi of pressure on the power stroke over a piston or about 7700 pounds resisting the piston on the power stroke but not on the next exhaust stroke and it mind boggling it holds together for even a second or two if we throw in the rings and bearing weights

4 bolt blocks using all forged rotating assembly's can usually be safely operated for brief periods at up to 4500 ft per minute in piston speeds, but the stress levels are cumulative and the higher the average rpm rage the lower the life expectancy
hitting the engines red line 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.

the 4000fpm rule is a general V8 rule with STOCK components,and it takes into account several factors like lower end stress and likely valve train components, if you upgrade to ALL FORGED & balanced components, and aftermarket valve train, you can usually go up to 4500fpm in piston speed, keep in mind 4000fpm=48,000 inches per minute and you need to use twice the stroke per revolution in calculations ... speed.html



generally hydraulic 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 ... ILEID=9290 ... re=related







a reasonable limit on cast pistons usually falls near 4000 feet per minute in piston speeds
so your stroke is a factor not just rpms

a balanced set of quality forged pistons can probably handle 4500fpm, or a bit more

keep in mind that's max PEAK engine rpms, that should only rarely be reached ,your engine will NEVER stay together if subjected to those rpms consistently 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

Re: redline

PostPosted: November 26th, 2008, 2:04 pm
by grumpyvette
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 engines red-line is the rpm limit where the stress is approaching the strength limitations of the weaker components in your engine, in most cases that's the rod bolts, rods, or a valve train control issues, on most engines.
Id also point out that valve springs tend to loose tension and load rate as they age and stress on valve train parts is cumulative , as is wear on the bearings, and stress on rods and pistons, etc.,so the rpm range you could safely maintain when the engine was new will be reduced as the engine ages.
on most American V8 engines that's reached at about 4000 fpm (feet per minute) in piston speed, with high quality forged components and a solid lifter valve train, that's at about 4500fpm,now Id point out Im not about to suggest that at 4100 fpm a stock engine self destructs,because theres a wide variation in components, and different engine designs but stress is cumulative, and those are reasonable limits as a guide,for a engine designed for long term street use durability now your engine may exceed that limit hundreds of times or only a few times before problems occur, but exceed the limits and eventually youll over stress rotating assembly ,components until they fail.
What I like is, building cost effective,easily duplicate-able engines, from off the shelf components if what I need is available, if not I have zero problem fabricating some components or modifying whats available if required,engines that use mostly built from well matched correctly fitted components. I like having total predictable control on a valve train components and long term durability, with low maintenance and in most cases large displacement high compression engine combos
I prefer to keep the piston speeds under 4300feet per minute , and Id prefer to keep the compression as high as I can , simply because its generally going to make for a more responsive combo with higher torque in the useable rpm range.
now in most cases thats engines of 400 cubic inches or larger displacement, and when I can 11:1 or higher compression, so 7000rpm is about where Im comfortable limiting valve train speeds, and I have zero issues building bigger displacement combos that might never see 6500rpm.
given the choice Id gladly give up that last extra 2%-5% of potential horse power in exchange for an added 20%-to-50% greater engine life expectancy, which is in many cases a choice you ARE forced too make

your fpm is found by multiplying the engine rpms x the stroke x 2
on a 383 sbc the stroke is 3.75", so 4000fpm on a 383 is reached at 4000rpm x 12"(inches)= 48,000 inches per minute of piston travel
so 48,000/7.5" of stroke =6400rpm, (7.5" is twice the engines 3.75" stroke)
exceeding that is a reasonably sure way to over stress the engine or (get into valve control issues ( which is a different can of worms)

this is a good place to point out that ARP makes much stronger rod bolts and main cap studs,and that the better aftermarket forged 4340 forged connecting rods are far stronger than most OEM connecting rods and that a 7/16" arp rod bolt is on average 200%-250% stronger than a stock 3/8" OEM rod bolt.
balanced components tend to put significantly lower stress on the engine components


obviously valve float and improper lash clearance can cause problems but in some cases, lash caps can reduce wear

in most cases when you see valve tip damage like this its the result of valve float or a weak valve spring , in many cases youll need to swap to a higher spring load rate and new springs to prevent or reduce this damage
interesting info from ARP

540 RAT posted this info
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
Grade 5 hardware store general purpose bolt:
Yield strength = 85,000 psi
Tensile strength = 120,000 psi
Cost = a few cents each
Grade 8 hardware store general purpose bolt:
Yield strength = 120,000 psi
Tensile strength = 150,000 psi
Cost = a few cents each
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.
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".
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
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.


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.


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.



• 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.



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"



Other Stresses & related info





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

"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.

Image ... index.html

a reasonable limit on cast pistons usually falls near 4000 feet per minute in piston speeds
so your stroke is a factor not just rpms

a balanced set of quality forged pistons can probably handle 4500fpm, or a bit more

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

but the results similar in most cases




Re: redline

PostPosted: November 14th, 2009, 4:32 pm
by grumpyvette
on many American V8 engines with FORGED and BALANCED COMPONENTS, its not the lower end strength, lack of oil pressure or the rotating assembly thats likely to be the CAUSE of a failure,but it is the other often over looked factors, Id bet serious money that most guys that find a bent connecting rod, blame the lower end rotating assembly, never thinking further into the cause,when in many cases thats the result not the cause of the failure.
detonation can destroy a piston or bearings in minutes, lack of oil flow over the rings,valve springs, or pistons will cause them to overheat rapidly, and valve control issues are far more likely than lower end problems
lack of cool pressurizer oil flowing over valve springs, will cause them to rapidly over heat and loose temper reducing their load carrying capacity, and rockers,lack of cool pressurizer oil flowing over lifters and cam lobes will cause very rapid over heating, lack of oil flow can cause rapid wear issues and when a cam or lifter starts to wear metallic crap can rapidly ruin bearing surfaces, and destroy clearances reducing oil control and reducing pressure.
your oil filter WON,t generally catch all the metallic crud before it can cause problems.
the cause of an engine failure can be tuning issues, lack of correct clearances, lack of lubrication or cooling , the result is frequently a piston trying to compress a non-compressible object like a bent valve that didn,t fully seat due to valve float or valve control issues or a chunk of busted ring land that fractured from over heating due to detonation , lack of a film of cooling oil, bad fuel/air ratio,or incorrect ignition advance curve.
One other factory to think about ,is that ARP main studs are stronger than factory main cap bolts, and and if a main cap starts moving under high stress at the upper rpm limits the oil pressure tends to drop as the clearances tend to increase.

Calculating Maximum Safe RPM

Max. Safe RPM = Mean Piston Speed (ft/min) x 6
Divided by Stroke in Inches

Example for a budget aftermarket forged crank in a 4-inch stroke small-block Chevy:
4,800 x 6 = 7,200 rpm

keep in mind valve train control problems can occur before piston speed is at max limits

Maximum Mean Piston Speeds for Above Formula:
Factory cast-iron cranks 3,750 ft/min
Aftermarket cast-steel cranks 4,500 ft/min
Factory forged cranks 4,600 ft/min
Budget aftermarket forged cranks 4,800 ft/min

Typical race aftermarket cranks 5,500 ft/min
High-dollar custom endurance race cranks 6,000 ft/min
ProStock/Mountain Motors 7,500 ft/min
Formula One 7,500+ ft/min [/size]

Re: redline

PostPosted: August 18th, 2011, 6:23 pm
by grumpyvette
Written by David Reher

Recently I've had several conversations with racers who wanted to build engines with long crankshaft strokes and small cylinder bores. When I questioned them about their preference for long-stroke/small-bore engines, the common answer was that this combination makes more torque. Unfortunately that assertion doesn’t match up with my experience in building drag racing engines.

My subject is racing engines, not street motors, so I’m not concerned with torque at 2,000 rpm. In my view, if you are building an engine for maximum output at a specific displacement, such as a Comp eliminator motor, then the bores should be as big as possible and the stroke as short as possible. If you’re building an engine that’s not restricted in size, such as a heads-up Super eliminator or Quick 16 motor, then big bores are an absolute performance bargain.

I know that there are drag racers who are successful with small-bore/long-stroke engines. And I know that countless magazine articles have been written about “torque monster” motors. But before readers fire off angry e-mails to National DRAGSTER about Reher’s rantings on the back page, allow me to explain my observations on the bore vs. stroke debate.

In mechanical terms, the definition of torque is the force acting on an object that causes that object to rotate. In an internal combustion engine, the pressure produced by expanding gases acts through the pistons and connecting rods to push against the crankshaft, producing torque. The mechanical leverage is greatest at the point when the connecting rod is perpendicular to its respective crank throw; depending on the geometry of the crank, piston and rod, this typically occurs when the piston is about 80 degrees after top dead center (ATDC).

So if torque is what accelerates a race car, why don’t we use engines with 2-inch diameter cylinder bores and 6-inch long crankshaft strokes? Obviously there are other factors involved.

The first consideration is that the cylinder pressure produced by the expanding gases reaches its peak shortly after combustion begins, when the volume above the piston is still relatively small and the lever arm created by the piston, rod and crank pin is an acute angle of less than 90 degrees. Peak cylinder pressure occurs at approximately 30 degrees ATDC, and drops dramatically by the time that the rod has its maximum leverage against the crank arm. Consequently the mechanical torque advantage of a long stroke is significantly diminished by the reduced force that’s pushing against the piston when the leverage of a long crankshaft stroke is greatest.

An engine produces peak torque at the rpm where it is most efficient. Efficiency is the result of many factors, including airflow, combustion, and parasitic losses such as friction and windage. Comparing two engines with the same displacement, a long-stroke/small-bore combination is simply less efficient than a short-stroke/big-bore combination on several counts.

Big bores promote better breathing. If you compare cylinder head airflow on a small-bore test fixture and on a large-bore fixture, the bigger bore will almost invariably improve airflow due to less valve shrouding. If the goal is maximum performance, the larger bore diameter allows the installation of larger valves, which further improve power.

A short crankshaft stroke reduces parasitic losses. Ring drag is the major source of internal friction. With a shorter stroke, the pistons don’t travel as far with every revolution. The crankshaft assembly also rotates in a smaller arc so the windage is reduced. In a wet-sump engine, a shorter stroke also cuts down on oil pressure problems caused by windage and oil aeration.

The big-block Chevrolet V-8 is an example of an engine that responds positively to increases in bore diameter. The GM engineers who designed the big-block knew that its splayed valves needed room to breath; that’s why the factory machined notches in the tops of the cylinder bores on high-performance blocks. When Chevy went Can-Am racing back in the ’60s, special blocks were produced with 4.440-inch bores instead of the standard 4.250-inch diameter cylinders. There’s been a steady progression in bore diameters ever since. We’re now using 4.700-inch bores in NHRA Pro Stock, and even bigger bores in unrestricted engines.

Racers are no longer limited to production castings and the relatively small cylinder bore diameters that they dictated. Today’s aftermarket blocks are manufactured with better materials and thicker cylinder walls that make big-bore engines affordable and reliable. A sportsman drag racer can enjoy the benefits of big cylinder bores at no extra cost: a set of pistons for 4.500-inch, 4.600-inch or 4.625-inch cylinders cost virtually the same. For my money, the bigger bore is a bargain. The customer not only gets more cubic inches for the same price, but also gets better performance because the larger bores improve airflow. A big-bore engine delivers more bang for the buck.

Big bores aren’t just for big-blocks. Many aftermarket Chevy small-block V-8s now have siamesed cylinder walls that will easily accommodate 4.185-inch cylinder bores. There’s simply no reason to build a 383-cubic-inch small-block with a 4-inch bore block when you can have a 406 or 412-cubic-inch small-block for about the same money.

There are much more cost-effective ways to tailor an engine’s torque curve than to use a long stroke crank and small bore block. The intake manifold, cylinder head runner volume, and camshaft timing all have a much more significant impact on the torque curve than the stroke – and are much easier and less expensive to change.

Re: redline

PostPosted: August 18th, 2011, 6:28 pm
by grumpyvette
Written by David Reher

Looking back at the 2004 season, I can attribute much of the performance improvement in Pro Stock to faster engine speeds. It’s difficult to believe that 500cid Pro Stock engines now routinely turn 10,000 rpm, but the truth is plain to see on the data recorders and on the time slips.

The trend toward higher and higher engine speeds was also evident in NASCAR stock car racing until the rulemakers applied the brakes with new restrictions on rearend and transmission gear ratios. Now the growing interest in fast bracket racing, Top Sportsman, and Top Comp eliminators is bringing this same high-rpm technology to sportsman drag racers.

Why does turning an engine higher make a race car run faster? This is my final column of the year, so I’ll offer my ideas and hope that they give racers something to think about over the winter break.

The simple explanation is that raising rpm effectively increases an engine’s displacement. This might seem nonsensical because the volume displaced by the pistons doesn’t change, but consider the effects of filling and emptying the cylinders faster in real time. An internal combustion engine is an air pump, and if we turn that pump faster, we can theoretically burn more fuel in a given amount of time and consequently produce more power. For example, an eight-cylinder engine running at 6,000 rpm fires its cylinders 24,000 times in one minute (assuming perfect combustion). Increase the engine’s speed to 8,000 rpm and it will fire 32,000 times per minute, a 33 percent increase. The volume of air and fuel that moves through the engine is now equivalent to an engine with a much larger displacement. There are also 8,000 additional power pulses per minute transmitted to the crankshaft that can be harnessed to turn the wheels and accelerate the car.

Raising engine speed is analogous to supercharging or turbocharging a motor; the goal is to increase the volume of air and fuel that moves through the engine. The airflow is increased with a forced induction system by pressurizing the intake system; in a naturally aspirated engine, the airflow is increased by raising rpm. If done correctly, both approaches will increase power.

A higher revving engine also permits the use of a numerically higher gear ratio to multiply the engine’s torque all the way down the drag strip. Let’s say an engine that produces 1,000 horsepower at 7,000 rpm is paired with a 4.56:1 rearend gear ratio. If this engine is then modified to produce 1,000 horsepower at 8,000 rpm, it can now pull a 4.88:1 or 5:14:1 rearend gear without running out of rpm before reaching the finish line. The numerically higher gear ratio gives the engine a mechanical advantage by multiplying its torque by a greater number to accelerate the car faster – in effect, it has a longer lever to move the mass.

I learned this lesson many years ago when I started drag racing. I raced my little 302cid Camaro against 426 Hemis and 440cid Max Wedge Mopars. The big-inch engines had thunderous low-end power, but my high-revving 302 with a 4.88:1 rear gear would just kill them because they were all done at 5,800 rpm. My small-block had much less torque and horsepower, but I could multiply the power it had with a steeper gear ratio. The same principle applies to racing a Pro Stock or a Top Sportsman dragster. By turning more rpm, we can use a greater gear ratio to produce more mechanical advantage to accelerate the car.

There are limits to engine speed, of course. Higher rpm increases parasitic losses from friction and windage. The stability of the valvetrain also restricts engine rpm. However, with the technology developed in NASCAR and in Pro Stock, racers are learning how to build engines that operate reliably at high rpm. Research and development on valve materials, springs, rocker arms, and pushrods are now being applied to serious sportsman drag racing engines. In fact, I wish that I had some of the parts that we now install in our high-horsepower sportsman engines for our Pro Stock program a few years ago!

While increasing rpm is generally a good thing for a racing engine, it also puts more responsibility on the owner. A high-rpm combination requires more vigilance and more maintenance than a low-rpm motor. It’s important to check the valve lash frequently and to look for early warning signs such as weak or broken valve springs. Neglecting these parts in a high-rpm racing engine can produce some very expensive problems.

Raising an engine’s operating range also requires complementary changes in the drivetrain and chassis. A high-rpm sportsman engine really needs a high-stall torque converter to realize its potential. With an automatic transmission, the engine speed should ideally drop 1,000 to 1,300 rpm after a gear change. For example, if the converter stalls at 6,700 rpm, the engine should be shifted at around 8,000 rpm. Shifting this engine at 7,000 rpm would simply put the engine back on the converter, causing the converter to operate inefficiently and wasting horsepower by heating the transmission fluid.

I’m excited about the emerging trend toward fast sportsman drag racing. I enjoy working with customers who want to go fast because it gives me an opportunity to deliver the benefits of our Pro Stock R&D to other racers. Not every racer wants or needs a high-rpm engine, but if the goal is to have a fast car, raising the redline is a proven approach.

Re: redline

PostPosted: October 4th, 2011, 5:06 pm
by grumpyvette
"HEY GRUMPY, I got a GREAT deal on some used imported small block connecting rods, from a local machine shop,Ive never heard of the company, but they say they are forged rods with 3/8" rod bolts" ... tions.html

Id recommend either of these,5.7" sbc connecting rods , yes they are more money, but Ive used them both in several builds,each, with both scat 9000 cast steel and scat 4340 forged steel cranks and they have quality rod bolts, and don,t have a wide variation in weight either, guys are using them even on nitrous engines that regularly spin 6600rpm with zero problems
the $170-$200 or so you save will be looked at as a really bad idea if the engine comes apart at high rpms and rods and rod bolts are highly stressed components, those cheaper import rods are NOT the same quality, and while they look like a bargain , they won,t continue to look like a bargain if one comes apart and the only thing you can salvage is the water pump and intake and valve covers like one guy I know had left when he tried spinning stock rebuilt chevy rods at 6700rpm
most chevy small block factory rods are VASTLY inferior in strength to many of the mid range and better aftermarket rods available.
a 7/16" cap screw type ARP rod bolt is EASILY 200%-300% stronger than a stock 3/8" factory rod bolt and frankly, the cost & TIME to correctly modify and prep stock rods is a total waste, its almost always cheaper to buy decent aftermarket rods.

theres an old saying in racing,