connecting rod & rod length too stroke info



bore to stroke ratio

Postby grumpyvette » November 13th, 2010, 5:33 pm

in the constant quest for increased performance many guys install stroker cranks to gain displacement, many guys don,t consider the changes in stroke mandate other changes such as increased friction from the changed rod angles and clearance issues plus the reduction in peak rpms that the increased piston speeds the longer strokes mandate will require.
longer strokes tend to have increased rotational friction and lower rod length to stroke ratios
both tend to limit upper rpm cylinder filling efficiency, longer strokes are not necessarily, a bad thing as the increased displacement helps torque but your valve size needs to keep pace with the displacement, and in many cases the restriction in max valve size, and the related curtain or flow potential, that the engines bore limits you too, will cause major increases in stroke to only produce slightly more torque
the IDEAL bore to stroke ratio and rod length, (from seeing the results of many engine dynos) seem to fall close to a range of stroke should be between about .70-.75 of bore diam.and rod length should be between about 1.55-1.9 times the stroke, but theres been many successful combos that did not fall into that basic range
examples of very efficient engines that maximize the horsepower per cubic inch of their displacement
are the 302 Chevy with its 3" stroke and 4" bore and the 477 BIG BLOCK with its 4.5" bore and 3.76" stroke
neither is ideal but both engine combos tend to produce excellent power for their displacement when the correct components are used to build them, but Id also point out that the common strokes on the same bore (the 383 in the 302s case and the 540bbc in the 477s case)produce MORE TORQUE AND HORSEPOWER, just not as much horse power PER CUBIC INCH OF DISPLACEMENT

in an ideal world having a 1.9:1 rod/stroke is great, in theory its worth a few extra hp and reduces friction and thrust drag on the cylinder walls, in the real world as long as your crank counter weights clear the piston skirts at bdc by at least .060 and the pistons don,t get closer than about .037 to the heads at TDC the rod length is not super critical as long as the piston speed stays under about 4000 fpm (feet per minute)with components similar to stock and with forged balanced components 4500fpm may be tolerated for brief periods...youll seldom have problems unless the valve train or lube system fails

QUOTE
"Under-square engines

These produce strong torque at low to mid range rpm's because of the "leverage" advantage of a longer stroke. But, under-square can be a negative trait, since a longer stroke usually means greater friction, a weaker crankshaft and a smaller bore means smaller valves which restricts gaseous exchange; however, modern technology has lessened these problems (explanation?). An under-square engine usually has a lower redline, but should generate more low-end torque. In addition, a longer stroke engine can have a higher compression ratio with the same octane fuel compared to a similar displacement engine with a much shorter stroke ratio. This also equals better fuel economy and somewhat better emissions. Going undersquare can cause pistons to wear more quickly (greater side-loads on the cylinder walls) and can cause ring seal problems and lubrication problems; with increased loads on the crankshaft, pistons, the piston pins, connecting rods, and rod bearings (due to piston speed). In general, a longer stroke leads to higher thermal efficiency through faster burning and lower overall chamber heat loss. A longer stroke will have greater port velocity at a given RPM, more torque due to more leverage on the crank, will achieve it's greatest efficiency at a lower RPM. Smaller combustion chambers are also more efficient, with the flame front having a shorter distance to travel- this leads to being more detonation resistant, and having an advantage for emissions.


Over-square engines

These are generally more reliable, wears less, and can be run at a higher speed. In over-square engines power does not suffer, but low-end torque does - it being relative to crank throw (distance from the crank center to the crank-pin). An over-square engine cannot have as high a compression ratio as a similar engine with a much higher stroke ratio, and using the same octane fuel. This causes the over-square engine to have poorer fuel economy, and somewhat poorer exhaust emissions. Breathing is an important advantage for over-square engines, as the edges of the valves are less obstructed by the cylinder wall (called "un-shrouded"). The big bore can fit larger (or more) valves into the head and give them more breathing room.

With shorter crankshaft stroke (and therefore piston travel) parasitic losses are reduced. Ring drag is the major source of internal friction and the crankshaft assembly also rotates in a smaller arc, so the windage is reduced. Oil-pressure problems caused by windage and oil aeration are lessened."

you might want to read thru these links

http://www.strokerengine.com/RodStroke.html

http://www.wallaceracing.com/enginetheory.htm

viewtopic.php?f=53&t=343&p=6341&hilit=redline#p6341

http://www.purplesagetradingpost.com/su ... ngine.html

viewtopic.php?f=53&t=510

http://www.rbracing-rsr.com/runnertorquecalc.html

http://hemrickperformance.com/valve.aspx

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

http://www.wallaceracing.com/runnertorquecalc.php

http://www.swartzracingmanifolds.com/tech/index.htm

http://tomorrowstechnician.com/Article/ ... adder.aspx

http://victorylibrary.com/mopar/intake-tech-c.htm

viewtopic.php?f=53&t=343&p=6341&hilit=redline#p6341

http://victorylibrary.com/mopar/chamber-tech-c.htm

http://victorylibrary.com/mopar/cam-tech-c.htm

http://victorylibrary.com/mopar/rod-tech-c.htm

http://victorylibrary.com/mopar/piston_position-c.htm

viewtopic.php?f=52&t=1070
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Re: connecting rod info

Postby grumpyvette » December 3rd, 2010, 7:06 pm

Connecting Rods - Enginology
On The Rod Again...
From the November, 2010 issue of Circle Track
By Jim McFarland

http://www.circletrack.com/enginetech/c ... ength.html

Connecting Rods
Connecting rod geometry, particularly...

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read full caption
Connecting Rods
Connecting rod geometry, particularly center-to-center length, can have a material influence on a variety of engine conditions.
It is generally acknowledged that connecting rod geometry, particularly center-to-center length, can have a material influence on a variety of engine conditions. These include specific relationships to valve timing (camshaft design), cylinder pressure history, spark ignition timing requirements and torque output, the latter with respect to the actual shape of torque curves. We'll touch on the more important of these a bit later.

Depending upon specific applications, connecting rods are perhaps some of the most highly stressed parts in an engine, particularly those intended for racing. From the high loads experienced at and just beyond TDC piston position during combustion to the tensile and unsymmetrical loading caused by offset piston pin axis, loads that are actually opposite to combustion pressure loads and stresses set up by lateral inertia, connecting rods become virtual "whips" that mechanically join pistons to the crankshaft.

Further complicating the issue are vibratory loads caused by oscillatory motion of a crankshaft, rotating about its axis while spinning in a normal direction. Visualize this set of load conditions in very slow motion. Each firing impulse intended to accelerate crankshaft rotation is applied as a force delivered in a span of time. Because of its inertia, a crankshaft can't immediately increase its speed and, therefore, is momentarily deflected in the same direction as its rotation. This deflection is local to the crank pin to which the load-delivering connecting rod is attached. Then, because of its elasticity, the crankshaft (at that pin location) will spring back against its direction of rotation, continuing this back-and-forth oscillatory motion until the next firing pulse is delivered to that particular crank pin. The connecting rod is thereby required to absorb what amounts to a series of tensile and compressive loads caused by oscillations of the crank pin, during primary crankshaft rotation.

Keep in mind that we've just provided a very simplistic description of the load dynamics experienced by the connecting rod for only one operational cylinder. The complexity of this varying load environment is increased by orders of magnitude when you add another seven cylinders and turn up the wick on rpm. So, when you think about connecting rods as "shock absorbers," several issues come to mind.

For example, consider cylinder pressure loads not as "hammer blows" to a piston but very rapid pressure rises that are influenced by combustion flame rate and net combustion pressure development. We also know that this pressure "history" is not constant or uniform as it is applied to a piston. Plus, whatever auxiliary forces are applied to a piston are also transferred in some way into the connecting rod. Rods can be designed too stiff, thereby transferring combustion pressure too aggressively to rod bearings and crank journal bearings. They can also be too flexible, and neither condition is acceptable. But in any case, rods need to absorb load spikes and minimize pressure transfer loss to prevent a waste of torque that's ultimately produced by the crank.

Perhaps one area of concern where connecting rod stiffness is important deals with vibratory loads produced by the torsional stiffness of a connecting rod's beam section, as piston weight is reduced. As you might expect, the reduction of rotating and reciprocating mass in an engine's crankshaft assembly can become a trade-off to the absorption of gas and mechanical loads by sheer mass alone. Visualize throwing a medicine ball to a 150-pound person and then to a 250-pounder and you may understand this more clearly.

Of course, to minimize the rotational resistance of a crankshaft assembly, reducing the weight of pistons and rods is a time-honored approach. However, compromising weight for strength and durability is the fulcrum about which this issue pivots. Perhaps one exception to this "rule" was in the early design of composite connecting rods (the so-called "poly motor" of years past), in which first-design rods were inordinately stiff and caused rod bearing failures for a lack of load absorption capability. On the other hand, lightweight materials that offer strength and low mass may be too costly to market, even in the average racing engine. So while other considerations must be included, the fundamental objectives should include strength, low weight, and durability.

In speaking with leading connecting rod manufacturers, you often hear that a high percentage of rod failures don't occur during the high pressure of combustion. Rather, it's during the exhaust stroke that a rod gets "yanked" away from TDC. This sudden movement of the piston causes abnormally high tensile loads in the rod's beam and leads to a fracture in this area, typically somewhere just below the piston pin end.

Also, failures can occur during either valve float or conditions of over-revving the engine. What happens is that the open valves (and lost combustion pressure) don't provide any sort of a cushion for pistons heading toward TDC. So when they pass through TDC, there's nothing to stop them from being "pitched" at the cylinder heads, often leading to another cause of tensile fracture in the beam section. In fact, the "effective" or dynamic weight of a piston passing through TDC under these conditions can be far in excess of its actual static weight. Multiple times, in fact.

Connecting Rods - Enginology

Connecting Rods
Connecting rods become virtual...

read full caption
Connecting Rods
Connecting rods become virtual "whips" that mechanically join pistons to the crankshaft and sometimes they fail. But situations like the one pictured above can be avoided by properly selecting and integrating various internal engine components.
Yet another common location for rod failure is a portion sometimes called the "hinge point," which is generally where a connecting rod's beam section changes in cross-section area (wide to narrow). Connecting rod designers frequently work in this area to determine the best compromises between rod strength and material selection. Of course, you should always include proper rod side-clearance, making certain not to provide excessive dimension that allows oil to create over-oiling of cylinder walls. Insufficient side-clearance can lead to over-heated and failed rod bearings, as well.

Finally, if we assume that a piston represents the "floor" of an engine's combustion space, then the rate of piston movement and time spent at each crankshaft angle will affect the rate of change in combustion space (volume). Of the reasons this is important, one is that piston movement can affect mixture density during the compression stroke (and subsequent flame rate and rise of combustion pressure). This, in turn, bears influence on spark ignition timing and the optimization of IMEP (minimizing "negative" torque). During an exhaust cycle, piston motion can also affect efficient cylinder evacuation and, therefore, is linked to proper exhaust valve timing.

Just considering these two peripherals of piston movement, we can immediately see that any changes to a piston's rate of travel may affect net cylinder pressure and power. Connecting rod length can, and does, influence cylinder pressure. Perhaps obscure is the fact that while longer connecting rods produce a larger included angle between rod axis and crank throw (stroke) at the same piston position and crank angle, it is piston motion approaching and leaving TDC and BDC that provides some interesting study.

Here's an example of that. As connecting rod length is increased, piston motion (both acceleration and velocity) away from TDC decreases. This results in a slower rate of pressure drop across the inlet path, therefore causing a reduction in intake flow rate (all else being equal). Unless compensation is made for this change in piston speed, some degree of volumetric efficiency may be lost.

In contrast to this effect upon volumetric efficiency (potential torque), piston "residence time" at and near TDC during combustion tends to hasten flame rate, correspondingly raise cylinder pressure per unit time, and enhance the tendency toward detonation. Reduced initial (or total) ignition spark timing, applied to reduce pre-TDC cylinder pressure, also increases IMEP by the reduction of negative torque. Or it can work against the piston as it approaches TDC during combustion.

Long rod combinations usually like intake manifold passages (actually heads and manifold) that help boost flow rates not provided by more rapidly descending pistons associated with shorter rods. So in addition to adjusting valve timing and lift patterns to match changes in piston speed needed to increase volumetric efficiency for increased rod length, port section areas and even carburetor sizing can be used to help restore reduced flow rates.

There is also the issue with reduced piston side-loading with long rod use. This reduction in friction horsepower has been attributed to power gains, especially when piston speed increases beyond about 2,500 feet/second. Improved ring life with long rods has also been claimed by some engine builders.

So while none of this month's Enginology was intended to advocate the use of short or long connecting rods, it emphasizes the importance of contemplating other engine functions that required consideration when making material changes to the rate of piston travel as a direct function of crankshaft angle. You will find that knowledgeable parts manufacturers, relative to the subject of connecting rod length, generally have a store of information linking how their components can affect an engine's ability to capitalize on rod length changes. If they don't, you may want to consider finding manufacturers who do. The concept of functional parts integration isn't without basis.
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Re: bore to stroke ratio

Postby enigma57 » February 17th, 2011, 4:38 am

:D Thanks, Grumpy! As it happens, I am right in the middle of a long stroke small block build and I will certainly read over all the links you posted before proceeding.

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 19th, 2011, 11:48 am

Figuring Compression Height of a piston

The compression height is the distance from the center of the wrist pin hole to the top deck of the piston. (see attached pix)

Custom pistons are available in any practical compression height to compensate for stroker or de-stroker cranks, long rods, or blocks which have been milled excessively, etc. Winston Cup & Busch motors usually have a very low or small compression height. (1.245 to 1.12 inches) The reason is that they have a shorter than stock stroke with a longer than stock rod length.
Ive always preferred to keep the piston pin out of the lower oil ring groove if I can, a compression height in the 1.5-1.375 range or larger will frequently allow that , but obviously each piston supplier will have a slightly different design, and obviously the applications will differ so you may be required to select something thats not always your ideal compromise
keep in mind theres some leeway, in that head gaskets can be used with different thicknesses to maintain a set quench distance, so if a piston sticks lets say as an example .010 out of the bore a .050 thick head gasket could be selected to maintain a .040 quench, if the pistons .015 down the bore, a thinner .025 head gasket could be used, so if your calculations show you need a 1.4" compression height a 1.390-1.410 could be used in most cases giving you a bit more choice selecting pistons
Image
1st thing you need to know is the block height. To find this you need to measure from the crankshaft center line to the deck (cylinder head mounting surface) of the block.
Image

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2nd Next thing is rod length. To determine exact rod length, you should have a good pair of calipers to measure with. Measure the size of the rod bearing opening (big hole) and the size of the wrist pin opening, and divide them in two (or one half) Finally, determine the distance between the two openings (center of the rod) and add the half you just calculated. That gives you the rod length. It comes out to be the distance between the center of the two holes.
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Finally you need the stroke length.

Image

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Stroke length is;

twice the distance from the center line of the crankshaft main bearing journals to the cente rline of the connecting rod journals or ;

It is also the distance the piston moves up and down in the cylinder



Now that you have all the info, you can calculate the compression height of the piston;

Image

To calculate the compression height, use the following formula:

Block Height minus 1/2 the crank stroke, minus the rod length, minus the deck clearance (amount piston is "in the hole").

For example, a 350 Chevy engine with a stock 3.480 stroke, stock length 5.700 rod, standard .017 deck clearance and standard 9.025 block height would be:

3.480 stroke divided by 2 = 1.740

9.025 - 1.740 - 5.700 - .017 = a compression height of 1.568.

if you were building a 496 BBC the deck height on the standard blocks 9.8"
rods are typically 6.385"
stroke is 4.25", so half the stroke is 2.125" plus 6.385" rod length, subtracted from 9.8" deck height, =1.29" piston compression height

https://www.uempistons.com/index.php?ma ... iston_comp

related info

viewtopic.php?f=52&t=4081&p=12278&hilit=quench#p12278

viewtopic.php?f=52&t=727

http://www.kb-silvolite.com/calc.php?action=piston_comp

viewtopic.php?f=53&t=3061&p=8095&hilit=piston+suppliers#p8095

viewtopic.php?f=53&t=2208&p=5942&hilit=piston+suppliers#p5942

viewtopic.php?f=69&t=2645&p=6834#p6834

http://www.lunatipower.com/Tech/Pistons ... eight.aspx
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Re: bore to stroke ratio

Postby enigma57 » February 20th, 2011, 12:52 am

:D Great info, Grumpy! The diagrams make it much easier to visualize the relationship of the various component parts that make up an engine and how they work together.

Thanks,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 21st, 2011, 6:04 am

quote=enigma57

"If I replaced the entire reciprocating assembly with a crank having 1/8" less stroke and went to longer 6" rods and pistons having a comp. height of 1.0875", it would improve rod/stroke ratio somewhat.....

4.000" stroke / 5.850" rod = 1.462:1 (My present combo)
3.875" stroke / 6.000" rod = 1.548:1 (New combo)
3.75" stroke / 6.000" rod= 1.6:1


Or I could destroke to 3.300" and use 6.2" rods along with my present 4.125" bore pistons having 1.170" comp. height......

3.300" stroke / 6.200" rod = 1.878:1 (This combo should rev easily and nets 352.8 cu. in. displacement)
3.000" stroke / 5.700" rod = 1.900:1 (stock 265,283,302)"



ok lets say youve got a new dart block with a 4.125" bore and your selecting a crank, youve narrowed the choices to a 3.30", a 3.75" and a 4" stroke crank combo


I think your placing a bit more concern on rod to stroke ratio and rod angle than is warranted, under the conditions.
every choice has a compromise in some area, and as the rpms go up so does the stress on the rotating components and valve train.
Image
keep in mind that your likely to produce about 1.25-1.35 hp per cubic inch with a well tuned combo like you've described , lets assume the larger combo only makes 1.25 hp per cubic inch and a de-stroker with its better rod angle makes a full 1.35:1 cubic inch

427 x 1.25=534(4" stroke)
372 x 1.35=502 (3.75" stroke)
352 x 1.35=476 (3.30 stroke)
but keep in mind the larger engines going to have about a 30-60 ft lb or more torque advantage over the smaller displacement combos over most of the rpm range and that power will come in at about 400rpm-600rpm lower and its responsive low and mid rpm torque that tends to be far more important than peak hp on a street car.
and its going to be cheaper and easier to maintain valve train control,at 6300rpm and below a good hydraulic roller cam,possibly with a rev kit should control the valve train,
yes there's some advantage in the better rod stroke ratio, but it will quickly be compromised by the need to rev the engine to higher rpms and valve train control issues , in many combo, less say your limiting the engine to 4200fpm in piston speed?
a 4" stroke=6300rpm . or about the max rpm any hydraulic lifter valve train is likely to function reliably
a 3.75" stroke=6700rpm =surely solid lifter rpm with a stud girdle
a 3.5" stroke=7200rpm=surely solid lifter rpm =shaft rocker time
a 3.3" stroke=7600rpm=surely solid lifter rpm =shaft rocker time
I get asked why the 301/302-327-331 sbc engine, is just not seen as commonly any more, the basic reason is the 283-327 they were built from is no longer as common as the current 350 basic core engine the 383-396-401 gets built from, as the 350 sbc is far more common.
just some info guys....there's a GOOD REASON why the 302 is less than popular compared to a 383-401 stroker built from a 350 basic block.
THERE'S NO way a 302 with its 3" stroke and the higher stress on the valve train that rpms over 7000rpm that the 302 sees will match the results and dependability a 383-396-401 stroker combo with its 3.75"-3.875" stroke and under 6500rpm valve train stress will produce
lets say a 302 can produce 1.25-1.4 hp per cubic inch, you can do the same with a 396 sbc
your looking at say 410 hp for the 302 and a similar 396 sbc costing almost the same will produce 535 hp with the same hp per cubic inch WITH LOWER VALVE TRAIN STRESS, its a FACT your far more likely to have valve train problems at over 6500rpm than under that rpm
there's also a much faster ramp up on the torque curve with the larger displacement.
we USED to build 301-327-331 sbc when the cylinder heads flow limited the effective displacement that could effectively be fed, those days are long gone, with the current aftermarket heads.

example

http://airflowresearch.com/articles/article031/A-P1.htm

http://www.chevytalk.org/fusionbb/showt ... id/131229/

http://www.bracketmasters.com/small_blo ... 383_cu.htm

http://airflowresearch.com/articles/art ... A14-P1.htm

http://airflowresearch.com/articles/art ... A16-P1.htm

BTW USING A HYDRAULIC LIFTER VALVE TRAIN in a 302-331 SBC that's built for MAX HP,is about as useful as snow shoes on a snake

ID also point out the differences in bearing sizes and the difficulty in building good compression with flat top pistons with the shorter stroke combos



for those guys that think high rpms are the way too go....,think about this, in a correctly clearanced and balanced lower assembly,
piston speeds should almost always be UNDER 4000fpm with correctly reworked stock type parts , or 4500FPM with all forged aftermarket race quality parts, if you expect the lower assembly to live a decent life span, that's,
between 8000-9000rpm on a 302 and 6400-7200 rpm with a 383 so as you should see, its far more likely the valve train is the weak link that determines the RED LINE
since the cars engine speed is usually restricted to the rate of acceleration of the car due too the engine being locked into driving the drive train,the larger engine has a slight advantage in acceleration with equal rear gears but in the real world you'll run 3.73:1-4.33:1 rear gears with a 383 and 4.56:1-5.13:1 with a 302, making the crank acceleration rates similar, or higher with the 302.
personally Ive never seen any advantage to spinning a smaller stroke engine to higher rpms, to make power, the stress on the valve train and lower assembly tends to cause more parts failures, its a whole lot easier to control valves at 6000rpm than at 8000rpm, and it gets darn expensive when pistons kiss valves, keep the engine operating well within its safe/ low stress speed,limits and it will last far longer.
at 8000rpm the valves open and close in each cylinder 67 TIMES A SECOND, your approaching absurd inertial loads and control problems well under that at 6400 rpm where the valves open and close at 53 times a second




Image

[/color]
the discussion about whats the best connecting rod length to stroke ratio and what rod design should be selected has been hashed thru in a near endless debate, I,d suggest you pay a great deal of attention to the quality of the connecting rods, bearings used, quench,clearances and engine lubrication, and preventing detonation, and maximize oil and coolant cooling to keep both in a reasonable range (THERE THREADS ABOUT THIS) and use an internally balanced rotating assembly, and while longer rod ratios are in theory beneficial the proven benefits are usually minimal as long as you keep the piston speeds reasonable
before I even begin,to discuss this Id strongly suggest IF your planing an engine build, that you purchase a complete matched rotating assembly, thats internally balanced from a well known manufacturer, and an SFI certified DAMPER AND FLYWHEEL OR FLEX-PLATE because any attempt at matching, miss matched components will result in zero warranty on any problem, fitting, matching or breaking in,the components

anyone who understands physics and geometry and can use a calculator or do minimal research into those questions,understands that a change in rod length will also change several other of the parameters of the engine, not only do you change the rod length, you also change rod angles, ring drag, piston weight, piston pin height, ring stack height in some cases, crank counter weigh, piston dwell, exhaust scavenging timing etc.

If you take some time and actually calculate out , or do the math,what changes happen between lets say a 5.7” rod and a 6.0” rod on a 3.48” , or 3.75"stroke crank, in a 350 or 383 sbc.. You will soon see the actual amounts of the angle changes are very minimum. Then , your forced to ask your self how or even if these small changes in rod angle will affect the engines hp/torque, and to what extent, and will the changes be beneficial or hurt your results, and Id also point out that the compression and cam timing will also effectively change your results in some cases.

yes the internet is full of claims, claims that The motor will carry the power and torque curve father up into the rpm range with a longer connecting rod.
But you are forced to ask, by how much? and is the result , because of the rod length,change or because you used totally different lighter piston with a different ring package,and in many cases changed quench or compression etc. There are sure to be other changes that were required inside an engine with a rod length change, so you can't instantly conclude that the change in connecting rod length alone made the changes (IF ANY) you see on a dyno.

I doubt its even possible to build two virtually identical engines where only the rod length alone changed, not I personally have tried to install the longer rods and try to get a rod to stroke ratio as close to 2:1 AS OTHER FACTORS ALLOW.
this rod to stroke ratio,can be built with a 6" connecting rod and 3" stroke to build a 302 with a 283 crank and using a 327 or 350 4" bore block, or the same 283 crank in a 400 larger bearing size block with custom machined bearing spacers, to build a 4.155 bore and 3" stroke combo, correctly set up these combo are known to make very good horsepower per cubic inch, but theres no question that the added cubic inches of a 3.75" stroke crank far exceeds the results even with a less desirable rod to stroke ratio.
now a 350 with a 3.48" stroke with a 5.7' rod has a 1.64 rod to stroke ratio,
a 383 with a 5.7" rod has a 1.52:1 ratio
a 383 with a 6" rod has a 1.6"1 ratio and in every case Ive ever seen the increased displacement had a far greater effect on the 383 performance than the change in rod length seemed to induce.
as to connecting rods , from a mechanical limitations point of view, the rod bolts and the area they connect tend to be the connecting rod weak point, so Id strongly suggest ARP 7/16" rod bolts in cap screw rod designs, with the (I) beam in theory having a almost non-existent strength advantage IF THE ROD MATERIAL CROSS SECTION IS IDENTICAL, which it is usually not! I would simply suggest you shop carefully ,demand a 7/16" ARP rod bolt cap screw connecting rod of the length you prefer from a quality manufacturer, scat has been my connecting rod of choice for most engine builds


SBC
SCAT I BEAM
http://www.summitracing.com/parts/sca-25700/overview/

http://www.summitracing.com/parts/sca-2 ... /overview/


SCAT H BEAM
http://www.summitracing.com/parts/sca-6570020/overview/

http://www.summitracing.com/parts/sca-6 ... /overview/
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
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Re: bore to stroke ratio

Postby enigma57 » February 21st, 2011, 6:53 pm

:D Grumpy, your words are wise. I have a little time before I will be ready to assemble the short block and I have been giving plenty of thought to this aspect as well.

This being an engine for a road car...... With the 427 combo, I can give up some HP up top by re-gearing and keeping RPMs down, whilst still making gobs of torque in the low to mid range.

Back in 2003, a friend forwarded me a copy of a computer dyno program (Dyno 2003). Not sure how accurate these things are, but figured it would at least give some meaningful comparisons of changes in cam profile, so I ran calcs of various cam profiles having IVC that fell within a range of 8 degrees (ranging from 2 degrees earlier IVC than would result in 8.0:1 DCR to 6 degrees later IVC).

I ran cam profiles from one extreme to the other, from the very mild (and ancient) Isky 3/4 race E-4 solid flat tappet grind to the UltraDyne solid flat tappet short track cam I had Harold Brookshire grind me with wider 110 degree LSA some years back...... And several hydraulic and solid lifter cams in between.

By way of comparison, I plotted TQ and HP from 2,000 RPMs through 6,000 RPMs and also plotted average TQ and HP from 2,000 RPMs through 4,000 RPMs.

Geared as it is at present, my car will be turning 2,500 RPMs in 6th gear OD cruising at 70 MPH. HP and TQ averages for all cam grinds in the 2,000 - 4,000 RPM range were within 10 HP and 15 ft./lb. of one another...... Though peak TQ ranged from 3,500 RPMs to 4,500 RPMs and peak HP ranged from 5,000 RPMs to 6,000 RPMs, depending upon cam grind. (Given the use of the car, I have placed a self-imposed 6,000 RPM redline on this 4" stoke combo to limit piston speed to 4,000 fpm max at redline as you suggested some time ago.)

What changed most was where the engine made the most power. As expected, the milder cams made more power down low.

Interestingly, all were within 11 HP and 15ft./lb. of TQ at 3,500 RPMs. And as expected, the more aggressive grinds began pulling away from the milder cams which had made more power below 3,500. And from 3,500 RPMs to their respective redlines, each of the more aggressive grinds made up the difference between their lack of power at lower revs...... The averages from 2,000 - 4,000 RPMs being very close.

From 4,500 RPMs to their respective redlines, the more aggressive grinds pulled harder and the higher they spun, the more HP they made...... Whilst shifting TQ peak higher, but with ft./lbs. remaining very close to the numbers achieved at lower revs with the milder grinds. All in all, it was a very interesting comparison.

Here are some examples, beginning with the little Isky E-4 solid lifter grind from the '50s that makes the 427 into a low RPM stump puller......


Solid flat tappet......

------RPMs----- HP ---TQ

2,000 RPMs - 189 - 497 - Cam, Isky E-4 solid, flat tappet, 216 deg. int./216 deg. exh. @ 0.050", 108 deg. LSA, 0.425" lift at valve (both)
2,500 RPMs - 242 - 508
3,000 RPMs - 296 - 519 - Peak HP - 421 @ 5,000 RPMs
3,500 RPMs - 349 - 523 - Peak TQ - 523 ft./lb. @ 3,500 RPMs
4,000 RPMs - 391 - 513
4,500 RPMs - 421 - 492
5,000 RPMs - 421 - 442 - Avg. HP - 2,000 to 4,000 RPMs, 293.4
5,500 RPMs - 408 - 389 - Avg. TQ - 2,000 to 4,000 RPMs, 512 ft./lb.
6,000 RPMs - 376 - 329

------RPMs----- HP ---TQ

2,000 RPMs - 182 - 479 - Cam, Isky RPM-300 solid, flat tappet, 228 deg. int./228 deg. exh. @ 0.050", ground on 108 deg. (not 112 deg.) LSA, 0.448" lift at valve (both)
2,500 RPMs - 235 - 493
3,000 RPMs - 292 - 511 - Peak HP - 458 @ 5,000 RPMs
3,500 RPMs - 349 - 524 - Peak TQ - 527 ft./lb. @ 4,000 RPMs
4,000 RPMs - 402 - 527
4,500 RPMs - 439 - 513
5,000 RPMs - 458 - 481 - Avg. HP - 2,000 to 4,000 RPMs, 292
5,500 RPMs - 450 - 429 - Avg. TQ - 2,000 to 4,000 RPMs, 506.8 ft./lb.
6,000 RPMs - 427 - 374

------RPMs----- HP ---TQ

2,000 RPMs - 184 - 483 - Cam, Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve
2,500 RPMs - 239 - 501
3,000 RPMs - 297 - 520 - Peak HP - 517 @ 5,500 RPMs
3,500 RPMs - 359 - 538 - Peak TQ - 548 ft./lb. @ 4,500 RPMs
4,000 RPMs - 417 - 548
4,500 RPMs - 470 - 548
5,000 RPMs - 505 - 530 - Avg. HP - 2,000 to 4,000 RPMs, 299.2
5,500 RPMs - 517 - 493 - Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.
6,000 RPMs - 509 - 446

------RPMs----- HP ---TQ

2,000 RPMs - 170 - 446 - Cam, UltraDyne solid, flat tappet, 252 deg. int./252 deg. exh. @ 0.050", 110 deg. LSA, 0.525" lift at valve (both)
2,500 RPMs - 223 - 468
3,000 RPMs - 282 - 493 - Peak HP - 546 @ 6,000 RPMs
3,500 RPMs - 348 - 523 - Peak TQ - 555 ft./lb. @ 4,500 RPMs
4,000 RPMs - 413 - 543
4,500 RPMs - 475 - 555
5,000 RPMs - 519 - 545 - Avg. HP - 2,000 to 4,000 RPMs, 287.2
5,500 RPMs - 546 - 522 - Avg. TQ - 2,000 to 4,000 RPMs, 494.6 ft./lb.
6,000 RPMs - 546 - 478


Hydraulic flat tappet......

------RPMs----- HP ---TQ

2,000 RPMs - 176 - 463 - Cam, Comp Cams #CS NX268H-13 ground on 108 LSA, hydraulic, flat tappet, 224 deg. int./236 deg. exh. @ 0.050", 108 deg. LSA, 0.477" int./0.490" exh. lift at valve)
2,500 RPMs - 230 - 483
3,000 RPMs - 285 - 500 - Peak HP - 497 @ 5,500 RPMs
3,500 RPMs - 345 - 518 - Peak TQ - 522 ft./lb. @ 4,500 RPMs
4,000 RPMs - 402 - 527
4,500 RPMs - 447 - 522
5,000 RPMs - 484 - 508 - Avg. HP - 2,000 to 4,000 RPMs, 287.6
5,500 RPMs - 497 - 474 - Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.
6,000 RPMs - 486 - 425


------RPMs----- HP ---TQ

2,000 RPMs - 183 - 481 - Cam, Lunati 268 VooDoo hydraulic, flat tappet, 227 deg. int./233 deg. exh. @ 0.050", 110 deg. LSA, 0.489" int./0.504" exh. lift at valve)
2,500 RPMs - 237 - 497
3,000 RPMs - 294 - 515 - Peak HP - 479 @ 5,500 RPMs
3,500 RPMs - 355 - 532 - Peak TQ - 537 ft./lb. @ 4,000 RPMs
4,000 RPMs - 409 - 537
4,500 RPMs - 455 - 531
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 295.6
5,500 RPMs - 479 - 457 - Avg. TQ - 2,000 to 4,000 RPMs, 512.4 ft./lb.
6,000 RPMs - 455 - 399

------RPMs----- HP ---TQ

2,000 RPMs - 187 - 491 - Cam, Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)
2,500 RPMs - 243 - 510 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same
3,000 RPMs - 302 - 528 - Peak HP - 478 @ 5,000 RPMs
3,500 RPMs - 362 - 543 - Peak TQ - 547 ft./lb. @ 4,000 RPMs
4,000 RPMs - 417 - 547
4,500 RPMs - 458 - 535
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 302.2
5,500 RPMs - 468 - 447 - Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.
6,000 RPMs - 446 - 390

I believe you are right. Better to go ahead and build the larger displacement version. The key to keeping this long stroke engine together and keep it running for many, many miles will be to gear the car to keep RPMs down and build the engine for low and mid range torque.

Of the cams I ran through the desk top dyno, these three came out on top. I have ranked them below based upon how each did when comparing average TQ and HP produced between 2,000 and 4,000 RPMs (the engine speeds this engine will see the most out on the highway). Surprisingly, the little Isky 274 Mega short track hydraulic flat tappet grind came out on top with my engine combo....... Followed by the Isky 530-A short track solid flat tappet grind and the Lunati 268 VooDoo hydraulic flat tappet street cam......

Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)

Peak HP - 478 @ 5,000 RPMs
Peak TQ - 547 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 302.2
Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.


Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve

Peak HP - 517 @ 5,500 RPMs
Peak TQ - 548 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 299.2
Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.


Lunati 268 VooDoo hydraulic, flat tappet, 227 deg. int./233 deg. exh. @ 0.050", 110 deg. LSA, 0.489" int./0.504" exh. lift at valve

Peak HP - 479 @ 5,500 RPMs
Peak TQ - 537 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 295.6
Avg. TQ - 2,000 to 4,000 RPMs, 512.4 ft./lb.

* The Isky 530-A solid lifter grind made the most HP at peak and nearly the same TQ at peak as the Isky 274 Mega hydraulic.

* The Lunati 268 VooDoo hydraulic made nearly the same HP at peak as the Isky 274 Mega hydraulic.

* What I like about the 274 Mega is that it not only made more average power where I will need it most out on the road...... It made peak TQ and HP numbers very near to the others, but at 500 fewer RPMs.

What do you think?

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 22nd, 2011, 5:39 am

the isky looks good but Id be looking for a bit more duration and a bit tighter LSA,all your selections will work with obvious various power curve changes, but Id concentrate on maximizing the 3500rpm-5800rpm torque band
try these cams with 1.6 rockers, with that bore stroke and displacement and knowing what your intended use is Id be looking to find a cam with about a 105-107 lsa.236/242-to-239-245 duration and about a .490-550 lift to make a good compromise between power and durability while maintaining low valve train stress
keep in mind that longer stroke and large displacement will react well to a tight lsa and a good deal of overlap but youll want to select the duration to maximize the mid and upper rpm power as the low rpm power should not be much of an issue, yet limit the duration to maintain at least decent low rpm drive ability,remember there's a big difference between a N/A cam used in a naturally aspirated engine and a super charger or nitrous design cam,that would have a wider lsa.
as always why not talk to the cam manufacturer and get their input and talk to them about gearing, rpm power bands etc.and your better off going with a slightly milder cam than over camming the combo, but a 427 sbc needs to breath and it takes some duration and overlap to do that efficiently,keeping in mind your goals, Id suggest and air gap dual plane intake, 750-800cfm cfm carb and long tube 1 3/4 headers
your not building a 350 so you'll need a good deal more cam, than a similar 350 build would require, but don,t get crazy, seeking peak numbers, its maximizing BOTH average torque and peak hp that matters most, don,t forget its the headers,heads, and intake that will also have a noticeable effect on the combo results.
and it only takes about 60hp or less to cruise, at 70mph at part throttle, and you don,t need off idle tire smoking torque, but you do need impressive acceleration from about 3500rpm up to 6000rpm

http://www.crower.com/products/camshaft ... -3956.html

http://www.crower.com/products/camshaft ... -3967.html


BTW
http://www.crower.com/cam-card-finder/

click, enter part number


related info
viewtopic.php?f=52&t=1070

viewtopic.php?f=52&t=480

viewtopic.php?f=55&t=624&p=11125&hilit=port+sectional#p11125

BTW the DD DYNO software tends to favor the wider LSA selection more than real world experience suggests is valid in my experience and over state the low rpm tq, but the peak hp numbers are usually reasonably close
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: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 10:34 am

Effects of a longer Rod
* Less rod angularity reduces wear.
* Lower piston velocity and acceleration reduces tensile loading of the rods.
* Less ignition timing is required which resist detonation.
* Compression can be increased slightly before detonation is a problem.
* Less intake runner volume is required and high rpm breathing is improved.
* Reduces scavenging at low rpm (weaker low RPM power).
* Longer TDC dwell time. (high RPM efficency).


Effects of a shorter Rod
* Increased rod angularity increases wear.
* Increased piston velocity and acceleration increases tensile loading of the rods.
* Increases scavenging at low rpm (increased low RPM power).
* Reduced TDC dwell time. (Reduced high RPM efficiency).


What they forgot to mention about the long rod is that the positive gain at TDC, is partially offset by wasted dwell time at BDC. The other problem is that on some piston designs, is the close proximity of the wrist pin hole and the bottom oil ring rail, on the piston, can cause flexing in the ring and less effective oil control.
below They used a 3.5" stroke for both rods, which is very close to a 350's 3.48" stroke.
As the graph shows, even 2" longer rod does not perform miracles, the better rod ratio has some effect but nothing thats going to totally make or break a combos effective power curve provided you keep the piston speeds well within the component strength limitations.
as stated before with basically stock components 4000fpm in piston speeds a good compromise, and with all forged and balanced components 4200fpm is doable, even 4500fpm for a second or two can be tolerated in some combos but stress goes up rapidly as rpms increase and stress is cumulative.
Image
Image
Image
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: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 11:03 am

http://www.iskycams.com/techinfo_index.html

Rod Lengths/Ratios: Much ado about almost nothing.

Why do people change connecting rod lengths or alter their rod length to stroke ratios? I know why, they think they are changing them. They expect to gain (usually based upon the hype of some magazine article or the sales pitch of someone in the parts business) Torque or Horsepower here or there in rather significant "chunks". Well, they will experience some gains and losses here or there in torque and or H.P., but unfortunately these "chunks" everyone talks about are more like "chips".

To hear the hype about running a longer Rod and making more Torque @ low to mid RPM or mid to high RPM (yes, it is, believe it or not actually pitched both ways) you'd think that there must be a tremendous potential for gain, otherwise, why would anyone even bother? Good question. Let's begin with the basics. The manufacture's (Chevy, Ford, Chrysler etc.) employ automotive engineers and designers to do their best (especially today) in creating engine packages that are both powerful and efficient. They of course, must also consider longevity, for what good would come form designing an engine with say 5% more power at a price of one half the life factor? Obviously none. You usually don't get something for nothing - everything usually has its price. For example: I can design a cam with tremendous high RPM/H.P. potential, but it would be silly of me (not to mention the height of arrogance) to criticize the engineer who designed the stock camshaft. For this engine when I know how poorly this cam would perform at the lower operating RPM range in which this engineer was concerned with as his design objective!

Yet, I read of and hear about people who do this all the time with Rod lengths. They actually speak of the automotive engine designer responsible for running "such a short Rod" as a "stupid SOB." Well, folks I am here to tell you that those who spew such garbage should be ashamed of themselves - and not just because the original designer had different design criteria and objectives. I may shock some of you, but in your wildest dreams you are never going to achieve the level of power increase by changing your connecting rod lengths that you would, say in increasing compression ratio, cam duration or cylinder head flow capacity. To illustrate my point, take a look at the chart below. I have illustrated the crank angles and relative piston positions of today's most popular racing engine, the 3.48" stroke small block 350 V8 Chevy in standard 5.7", 6.00", 6.125" and 6.250" long rod lengths in 5 degree increments. Notice the infinitesimal (look it up in the dictionary) change in piston position for a given crank angle with the 4 different length rods. Not much here folks, but "oh, there must be a big difference in piston velocity, right?" Wrong! Again it's a marginal difference (check the source yourself - its performance calculator).

To hear all this hype about rod lengths I'm sure you were prepared for a nice 30, 40, or 50 HP increase, weren't you? Well its more like a 5-7 HP increase at best, and guess what? It comes at a price. The longer the rod, the closer your wrist pin boss will be to your ring lands. In extreme situations, 6.125" & 6.250" lengths for example, both ring and piston life are affected. The rings get a double whammy affect. First, with the pin boss crowding the rings, the normally designed space between the lands must be reduced to accommodate the higher wrist pin boss. Second, the rings wobble more and lose the seal of their fine edge as the piston rocks. A longer Rod influences the piston to dwell a bit longer at TDC than a shorter rod would and conversely, to dwell somewhat less at BDC. This is another area where people often get the information backwards.

In fact, this may surprise you, but I know of a gentleman who runs a 5.5" Rod in a 350 Small Block Chevy who makes more horsepower (we're talking top end here) than he would with a longer rod. Why? Because with a longer dwell time at BDC the short rod will actually allow you a slightly later intake closing point (about 1 or 2 degrees) in terms of crank angle, with the same piston rise in the cylinder. So in terms of the engines sensitivity to "reversion" with the shorter rod lengths you can run about 2-4 degrees more duration (1-2 degrees on both the opening & closing sides) without suffering this adverse affect! So much for the belief that longer rod's always enhance top end power!

Now to the subject of rod to stroke ratios. People are always looking for the "magic number" here - as if like Pythagoras they could possibly discover a mathematical relationship which would secure them a place in history. Rod to stroke ratios are for the most part the naturally occurring result of other engine design criteria. In other-words, much like with ignition timing (spark advance) they are what they are. In regards to the later, the actual number is not as important as finding the right point for a given engine. Why worry for example that a Chrysler "hemi" needs less spark advance that a Chevrolet "wedge" combustion chamber? The number in and of itself is not important and it is much the same with rod to stroke ratios. Unless you want to completely redesign the engine (including your block deck height etc.) leave your rod lengths alone. Let's not forget after all, most of us are not racing at the Indy 500 but rather are hot rodding stock blocks.

Only professional engine builders who have exhausted every other possible avenue of performance should ever consider a rod length change and even they should exercise care so as not to get caught up in the hype.

if you do a bunch of research, or connecting rod length to stroke ratios and piston speeds etc., youll find a general consensus that the ideal rod ratio is near 1.9:1 on the stroke length,so something like a 3' or 3.25' stroke with a 6" connecting rods near the ideal , i also have built enough engines like a 406 chevy with a 5.7' rod or a 540 big block CHEVY with 6.385" rods that have closer to a 1.5:1 ratio that made exceptionally good power,to know that I read thru this thread and many other similar threads about rod to stroke ratios, that the phrase's

"TEMPEST IN A TEA POT"
"MOUNTAIN OUT OF A MOLE HILL"

BOTH COME TO MIND!"
yes theres measurable advantages to be gained in well matched component selection,and careful assembly, but I think the degree of concern here seems a bit excessive compared to the potential gains you might expect from the difference in parts being selected, compared to several other areas that potentially effect engine power that you might be concerned with.
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
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Re: bore to stroke ratio

Postby enigma57 » February 23rd, 2011, 12:31 pm

:D Thanks for the cam recommendations and timing card location info, Grumpy! I ran the Crower cams through Dyno 2003 with my 427 engine combo specs. Used valve lift with 1.6 rockers as suggested.

Of the 2 Crower grinds, my engine combo favours the larger of the 2 cams (00273S). Here are the figures I got.......


------RPMs----- HP ---TQ

2,000 RPMs - 178 - 466 - Cam, Crower 00258S hydraulic, flat tappet, 236 deg. int./242 deg. exh. @ 0.050", 106 deg. LSA, 0.479" int. / 0.475" exh. lift at valve
2,500 RPMs - 228 - 479 - Note: lift shown is with 1.6 rockers
3,000 RPMs - 282 - 493 - Peak HP - 425 @ 5,000 RPMs
3,500 RPMs - 338 - 508 - Peak TQ - 508 ft./lb. @ 3,5000 RPMs
4,000 RPMs - 387 - 508
4,500 RPMs - 416 - 486
5,000 RPMs - 425 - 446 - Avg. HP - 2,000 to 4,000 RPMs, 282.6
5,500 RPMs - 419 - 400 - Avg. TQ - 2,000 to 4,000 RPMs, 490.8 ft./lb.
6,000 RPMs - 395 - 346

------RPMs----- HP ---TQ

2,000 RPMs - 170 - 447 - Cam, Crower 00273S hydraulic, flat tappet, 239 deg. int./248 deg. exh. @ 0.050", 107 deg. LSA, 0.539" int. / 0.547" exh. lift at valve
2,500 RPMs - 225 - 473 - Note: lift shown is with 1.6 rockers
3,000 RPMs - 285 - 499 - Peak HP - 510 @ 5,500 RPMs
3,500 RPMs - 352 - 528 - Peak TQ - 546 ft./lb. @ 4,500 RPMs
4,000 RPMs - 415 - 544
4,500 RPMs - 468 - 546
5,000 RPMs - 496 - 521 - Avg. HP - 2,000 to 4,000 RPMs, 289.4
5,500 RPMs - 510 - 487 - Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.
6,000 RPMs - 502 - 439


Also ran the 2 Crane solid lifter cams mentioned on your link here......

viewtopic.php?f=52&t=1070


------RPMs----- HP ---TQ

2,000 RPMs - 163 - 427 - Cam, Crane 110921 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 106 deg. LSA, 0.518" int. / 0.536" exh. lift at valve
2,500 RPMs - 217 - 456 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same (used 1.5 rockers for simulation)
3,000 RPMs - 275 - 482 - Peak HP - 509 @ 5,500 RPMs
3,500 RPMs - 343 - 514 - Peak TQ - 539 ft./lb. @ 4,500 RPMs
4,000 RPMs - 407 - 535
4,500 RPMs - 462 - 539
5,000 RPMs - 492 - 517 - Avg. HP - 2,000 to 4,000 RPMs, 281
5,500 RPMs - 509 - 486 - Avg. TQ - 2,000 to 4,000 RPMs, 482.8 ft./lb.
6,000 RPMs - 504 - 441


------RPMs----- HP ---TQ

2,000 RPMs - 166 - 435 - Crane 114681 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 112 deg. LSA (per DD2003), 0.518" int. / 0.536" exh. lift at valve
2,500 RPMs - 220 - 463 - Note: 1.5 rockers used for simulation
3,000 RPMs - 279 - 488 - Peak HP - 511 @ 5,500 RPMs
3,500 RPMs - 345 - 518 - Peak TQ - 546 ft./lb. @ 4,500 RPMs
4,000 RPMs - 413 - 542
4,500 RPMs - 468 - 546
5,000 RPMs - 500 - 525 - Avg. HP - 2,000 to 4,000 RPMs, 284.6
5,500 RPMs - 511 - 528 - Avg. TQ - 2,000 to 4,000 RPMs, 489.2 ft./lb.
6,000 RPMs - 504 - 441


And here are the figures I got for the 2 Isky cams that came out best in prior calculations.....


------RPMs----- HP ---TQ

2,000 RPMs - 187 - 491 - Cam, Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)
2,500 RPMs - 243 - 510 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same (used 1.5 rockers for simulation)
3,000 RPMs - 302 - 528 - Peak HP - 478 @ 5,000 RPMs
3,500 RPMs - 362 - 543 - Peak TQ - 547 ft./lb. @ 4,000 RPMs
4,000 RPMs - 417 - 547
4,500 RPMs - 458 - 535
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 302.2
5,500 RPMs - 468 - 447 - Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.
6,000 RPMs - 446 - 390

------RPMs----- HP ---TQ

2,000 RPMs - 184 - 483 - Cam, Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve
2,500 RPMs - 239 - 501 - Note: 1.5 rockers used for simulation
3,000 RPMs - 297 - 520 - Peak HP - 517 @ 5,500 RPMs
3,500 RPMs - 359 - 538 - Peak TQ - 548 ft./lb. @ 4,500 RPMs
4,000 RPMs - 417 - 548
4,500 RPMs - 470 - 548
5,000 RPMs - 505 - 530 - Avg. HP - 2,000 to 4,000 RPMs, 299.2
5,500 RPMs - 517 - 493 - Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.
6,000 RPMs - 509 - 446


The Isky cams were run with 1.5 rockers, as I tried 1.6 on intake only, then 1.6 on exhaust only and then 1.6 for both. Power remained the same over the lower RPMs and showed only a small increase (4 - 5 HP & ft./lb. TQ) at peak TQ and HP with 1.6 rockers on both.

The Crower cams were run with valve lift calculated for 1.6 rockers as suggested.

The Crane cams were run with 1.5 rockers, as I tried 1.6 on intake only, then 1.6 on exhaust only and then 1.6 for both. Power remained the same over the lower RPMs and showed only a small increase (4 - 5 HP & ft./lb. TQ) at peak TQ and HP.

I have ranked the 2 Isky, the 2 Crower and the 2 Crane cams below by how they came out based upon average HP and TQ between 2,000 and 4,000 RPMs (the engine speeds my engine will see most out on the road)......


Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)

Peak HP - 478 @ 5,000 RPMs
Peak TQ - 547 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 302.2
Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.


Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve

Peak HP - 517 @ 5,500 RPMs
Peak TQ - 548 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 299.2
Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.


Crower 00273S hydraulic, flat tappet, 239 deg. int./248 deg. exh. @ 0.050", 107 deg. LSA, 0.539" int./0.547" exh. lift at valve

Peak HP - 510 @ 5,500 RPMs
Peak TQ - 546 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 289.4
Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.


Crower 00258S hydraulic, flat tappet, 236 deg. int./242 deg. exh. @ 0.050", 106 deg. LSA, 0.479" int./0.475" exh. lift at valve

Peak HP - 425 @ 5,000 RPMs
Peak TQ - 508 ft./lb. @ 3,5000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 282.6
Avg. TQ - 2,000 to 4,000 RPMs, 490.8 ft./lb.


Crane 114681 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 112 deg. LSA (per DD2003), 0.518" int. / 0.536" exh. lift at valve

Peak HP - 511 @ 5,500 RPMs
Peak TQ - 546 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 284.6
Avg. TQ - 2,000 to 4,000 RPMs, 489.2 ft./lb.


Crane 110921 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 106 deg. LSA, 0.518" int. / 0.536" exh. lift at valve

Peak HP - 509 @ 5,500 RPMs
Peak TQ - 539 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 281
Avg. TQ - 2,000 to 4,000 RPMs, 482.8 ft./lb.


As you predicted, of the 2 Crane cams, the Dyno 2003 program favoured the wider LSA version of the same cam.

Not sure why the Isky cam grinds made more comparative power than did the Crower and Crane grinds (in the RPM range my engine will be run at most), but my engine combo seems to respond better to the Isky cams, at least in computer simulation. And as you say, real world performance may vary.

Really appreciate your assistance, Grumpy!

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 12:51 pm

I,d doubt you'll be disappointed with any of the better selections, but if you do get the car dyno tested after its tuned it will be interesting to compare the real world results with the dd dyno predicted wild guess the software shows
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: bore to stroke ratio

Postby enigma57 » February 23rd, 2011, 3:34 pm

:D I agree, Grumpy! It will be very interesting to see how this works out in terms of real world TQ and HP at the planned operating range. FWIW...... I plan on using this engine as a test bed to see how several intake systems work, including a home brewed IR intake. It will be a while before the car is ready to go, but I will keep in touch and let you know how things are progressing.

There may be some compromises required when fabbing the exhaust system because I do not want the hassle of header leaks on a road car with full length, muffled exhaust. Nor do I want to fry starters nor have engine oil temps increase due to oil filter and header clearance issues.

On the other hand...... As I have made the decision to build this engine for torque and keep RPMs down...... I may be able to get the sizing I need in heavy wall tubing and fit it as I have in mind after all. At present, I am thinking in terms of welded thickwall upswept headers similar in function to the old MOPAR Max Wedge cast headers of the early '60s. These were actually a 4 - 2 - 1 Tri-Y design and I believe something similar would enhance torque all the more.

Best regards,

Harry

BTW...... I am thinking about going to a standard volume big block oil pump. I understand from speaking to several engine builders that Melling and others they make pumps for to sell under their brand name have gone through several production changes in recent years. First, they began utilizing thinner pump body castings sourced from the ChiComs with the result that a high percentage of pumps used in performance and heavy duty applications developed cracks in the pump bodies.

Then they tried to get by with powdered metal pump gears...... Which also didn't work out well, either. I understand that they are working their way back to building pumps as they once did...... But that there are still pumps in warehouses that have the thin body castings and / or the powdered metal gears. My dilemma is where to get a really good and reliable oil pump and shaft that will last the life of this engine. Any ideas?
BOHICA! - How's all that 'change' workin' for ya? Image
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 4:11 pm

http://www.moroso.com/catalog/categoryd ... code=17006

http://www.melling.com/Aftermarket/High ... Pumps.aspx
watch this video
http://www.summitracing.com/parts/MEL-10778/?rtype=4

the melling 10990 is generally a well built pump but the price has gone up significantly since they started building them correctly again.
its basically a big block 5 bolt pump set up to run in a small block application
Ive run them in most of my sbc engines built recently and its in my corvettes 383

Image

http://www.summitracing.com/parts/MEL-10990/

btw

10778 used for big block engines

http://www.summitracing.com/parts/MEL-10778/
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: bore to stroke ratio

Postby enigma57 » February 27th, 2011, 12:48 am

:D Thanks, Grumpy! That's good info. With the 383 small block applications you build, do you keep the spring that comes in the 10990 pump or replace it with the supplied standard pressure spring?

And have you used the 10990C (version with the anti-cavitation grooves)? On a mostly street driven application, I am a bit leery of the stated 'reduces pressure at idle'.

http://www.melling.com/Aftermarket/High ... Pumps.aspx

10990

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The 10990 is a Big Block style oil pump made to fit the Small Block applications.
* The drive and idler shafts have been extended to allow for additional support in the cover. Additional support eliminates dynamic shaft deflection at increased RPM levels.
* The cover is doweled to the pump housing to assure alignment of the shaft bores.
* The relief valve has a screw-in plug instead of a pin.
* The housing and cover are CNC machined and phosphate coated.
* An additional spring, the original stock replacement is supplied which will reduce bypass pressure if needed.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Patent No. 5,810,571.

10990C (Anti-Cavitation)

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The same as the 10990 except with the addition of grooves machined in the housing and cover. The grooves reduce cavitation effects in high RPM applications.
* Using this oil pump will reduce pressure at idle.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Racing applications only.
* Patent No. 5,810,571.


I have a 7 quart Moroso #20190 pan with kickouts on both sides. It is 8-1/4" deep......

http://www.moroso.com/catalog/categoryd ... code=11002

In the notes, Moroso oil pickup #24320 is spec'd for big block oil pumps, so am I right in assuming that this would be the correct pickup to use with the Melling 10990 pump and this pan?

I see that Moroso now has pans notched for strokes over 3.80" as well. I will hold off until I have clearanced my block, trial fitting the crank and pistons and oil pan to see if my #20190 pan will work. But if not, I will likely go to the #20195 pan that has the same dimensions, but is notched for longer stroke crank throws.

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 27th, 2011, 8:33 am

enigma57 wrote::D Thanks, Grumpy! That's good info. With the 383 small block applications you build, do you keep the spring that comes in the 10990 pump or replace it with the supplied standard pressure spring?
I use the lower pressure spring

And have you used the 10990C (version with the anti-cavitation grooves)? On a mostly street driven application, I am a bit leery of the stated 'reduces pressure at idle'.

http://www.melling.com/Aftermarket/High ... Pumps.aspx

10990

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The 10990 is a Big Block style oil pump made to fit the Small Block applications.
* The drive and idler shafts have been extended to allow for additional support in the cover. Additional support eliminates dynamic shaft deflection at increased RPM levels.
* The cover is doweled to the pump housing to assure alignment of the shaft bores.
* The relief valve has a screw-in plug instead of a pin.
* The housing and cover are CNC machined and phosphate coated.
* An additional spring, the original stock replacement is supplied which will reduce bypass pressure if needed.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Patent No. 5,810,571.

10990C (Anti-Cavitation)

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The same as the 10990 except with the addition of grooves machined in the housing and cover. The grooves reduce cavitation effects in high RPM applications.
* Using this oil pump will reduce pressure at idle.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Racing applications only.
* Patent No. 5,810,571.


I have a 7 quart Moroso #20190 pan with kickouts on both sides. It is 8-1/4" deep......

http://www.moroso.com/catalog/categoryd ... code=11002

In the notes, Moroso oil pickup #24320 is spec'd for big block oil pumps, so am I right in assuming that this would be the correct pickup to use with the Melling 10990 pump and this pan?

ALWAYS VERIFY WHAT COMPONENTS ARE TO BE MATCHED IN ANY APPLICATION ,WITH THE MANUFACTURERS ,and ALWAYS ASK IF THEY HAVE OTHER RELATED TIPS,INFO, or RELATED WARNINGS but yes your more than likely correct but as Ive stated ..ALWAYS ASK[/b]

I see that Moroso now has pans notched for strokes over 3.80" as well. I will hold off until I have clearanced my block, trial fitting the crank and pistons and oil pan to see if my #20190 pan will work. But if not, I will likely go to the #20195 pan that has the same dimensions, but is notched for longer stroke crank throws.

Best regards,

Harry


hope that helps
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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