On a board to which I suscribe a guy said his ideal cam was 220/230, .600/.600 with a 112 LSA with 1.5 rockers. How would one expect to use this cam? What kind of advertised duration would be good to keep this cam barely in the smog legal range? Could this cam work as a flat tappet grind?

Does it matter if a cam has the exact same specs if it is a roller or flat, hyrdraulic or solid?

Instead of more duration on the exhaust valve could you use more lift? Can you use more lift in general since the exhaust valve is typically smaller?

What controls picking the right size valves? Is there a minimum/maximum ratio difference between intake and exhaust valves?

Is it a safe assumption that larger valves require a larger chamber? Is there a way to quantify the intake runner to chamber ratio? Would it matter?

What dimensions are required to determine the maximum lift you can get out of a cylinder head? How would a head/cam work if the cam was set to give you maximum lift up to about 3500 rpm and tapered down to about 75% of that lift by 6000 rpm? Would you need a big or small intake runner? The lift couldn't be bigger at the end than the beginning could it?

Does a lower head angle (23, 20, 18, 15) work better towards a larger or smaller bore? Does the lower head angle allow for higher or lower lift?

If the intake velocity is as important as the flow, how does one compute it? How do you determine how much flow and at what velocity you need it to accomplish x horsepower or torque?

At 100% volumetric effiency are you getting all of the flow you can out of a head, getting as much flow as the velocity will allow, both or neither?

How do you match the intake manifold to the head to get the right flow/velocity?

Wow, you said a mouthfull.

There are some here who can probably answer each of your questions, though it would be book length to answer all of them in detail.

There are probably other books to read which are more in depth, Bogie may be able to suggest some, but my favorite book to suggest to people is "John Lingenfelter on modifying small-block chevy engines". This book is full of practical info for street/ strip vehicles, and gives you a good knowledge base on how things work together for their application. The most important thing is "your application". You have to be able to honestly admit what you want out of your motor, and what it will be doing most of the time. If you are looking for a full race motor, and have no buddies "in the know", then I'd suggest buying a crate race motor. If you are looking for the best street strip combo motor you can build yourself, then the book I mentioned above is a great (although becoming dated) source of knowledge to help you make your choices.

john lingenfelter's book is pretty basic for the kinda things that he wants to know, lingenfelter's book would be good for someone that has never built a SBC before and wants some introductory info, such as things like, which oil galley plugs are commonly left out and that drops oil pressure, this is useful info to know but his questions are extremely abstract and in depth, im a big fan of david vizard books, which go through the basics and in depth technical aspects, but like you said also, i already mentioned to him that his best bet for answers (which i would also like to listen to) would be from bogie, hope he is around sometime soon...

Jason, you keep asking questions like these and you'll find yourself in college majoring in mechanical engineering. I'm going to respond with your questions and an answer in bold type. I'm going to respond to your question concerning port velocities seperatly because it takes a lot of space and this forum will choke on all this at once.

On a board to which I subscribe a guy said his ideal cam was 220/230, .600/.600 with a 112 LSA with 1.5 rockers. How would one expect to use this cam? What kind of advertised duration would be good to keep this cam barely in the smog legal range? Could this cam work as a flat tappet grind?  (Most heads don’t show much flow improvement over a half inch, unless you’re a competitive racer such high lifts are not warranted, the extra machining is expensive and the wear is intolerable on a street machine. Rule of thumb is that lift should peak at 1/4^th the diameter of the valve, research shows that the rate of increase in flow abruptly reduces once the lift exceeds this point. For example a 2 inch valve will show significant flow increases for every .1 inch it’s opened till reaching .5 inch, after that point the rate of gain becomes substantially less for each additional .1 inch of lift. This is very obvious when you read flow rates for heads that are documented out to .6 or .7 inch of lift. )

Does it matter if a cam has the exact same specs if it is a roller or flat, hydraulic or solid? (If the measurement is at the valve it doesn’t matter what’s causing the valve to move. Roller cams can be, but don’t have to be, ground more aggressive that flat tappet cams. To some extent the same can be said for a solid follower cam versus a hydraulic. Typically for a street engine that needs to be emissions legal, you won’t be using a cam profile that comes anywhere close to optimizing a solid follower whether it’s a roller or flat design.)

Instead of more duration on the exhaust valve could you use more lift? Can you use more lift in general since the exhaust valve is typically smaller? (Depends upon the port’s ability to flow a volume in a time period. If the port is at peak flow, adding valve lift won’t have an effect, however, spreading that flow over more time will help fill or empty the cylinder, therefore, more duration would be beneficial.

What controls picking the right size valves? Is there a minimum/maximum ratio difference between intake and exhaust valves? (RPMs where the engine’s power peaks and where it will operate controls valve, port and cam size. This is a case where one size does not fit all. The engine likes an intake velocity of about ½ the speed of sound. The design engineer attempts to hit that at the desired horsepower peak. A slow turning engine will have a mild cam, small ports and valves. This is relative to the size of the engine i.e. a 305 has smaller valves and ports compared to a 350 even though both may show peak power at the same RPM. Exhaust valves are close to 70 percent the size of intakes as tests have establihed this is sufficient. An engine that’s running a supercharger of some type or has nitrous injection may have a larger exhaust valve, a much longer duration cam lobe or both to compensate for more product being exhausted. A higher twisting engine needs bigger everything to keep the port velocities in the range of .5 Mach.)

Is it a safe assumption that larger valves require a larger chamber? Is there a way to quantify the intake runner to chamber ratio? Would it matter? (This is a sticky wicket; chamber shape, engine operating range, sparkplug location, camshaft timing and lift, number of valves, etc. all comes into play. Basically a large chamber with a large valve and port will perform at higher RPMs. SMOG heads of the 60s through early 80s tend to have large chambers, smallish valves and a chef’s stew of port sizes. They don’t do anything well. SBC 350 heads tend to use a 1.94 intake with a 1.5 exhaust on standard output engines, but you can find the same casting number with the same port size using a 2.02 and 1.6 combination on a high performance version. Given the cam difference between these engines, one is stuck to question whether the increased valve diameter really provides any benefit without also porting the head. You can use the valve size to control the gas velocity in the port throat where a larger valve will slow the velocity around the corner from the port into the throat and past the valve which will allow pressure recovery past the valve which may improve local flow. This is the stuff that makes flow bench testing interesting. Notice all the adjectives and adverbs, this is place where little is absolute one way or the other. Small nuances make for big differences much is controlled by atmospherics, the set up of test equipment and the person running the test. of one day by one person often bear scant relationship to readings taken on another day especially by another person or machine. If this was reducible to a couple absolutely dependable equations a 100 years later we wouldn’t sill be screwing around with flow-benches and dynamometers.)

What dimensions are required to determine the maximum lift you can get out of a cylinder head? (The point of contact between the top of the valve guide and the bottom of the spring retainer is the final arbiter. In a way this and the length of the valve stem, assuming the dimensions of the head casting can’t be changed, establishes the available range of lift. The cam bearing diameter also sets a limit on how tall a lobe can get. Loads imposed back down the pushrod into the lifter and reacted between the lifter and the cam establish how much rocker ratio you can use which also places a limit on how much lift you can force with a bigger ratio rocker. Component loading goes up to the square of lift and RPM so small increases in lift net large increases in forces on the parts.) How would a head/cam work if the cam was set to give you maximum lift up to about 3500 rpm and tapered down to about 75% of that lift by 6000 rpm? (This is what is happening to cylinder filling without needing to do anything mechanically. The fight is to keep this from happening. The steady loss of cylinder filling above the torque peak is why power curves for both torque and horsepower peak over and decline at some point in the RPM range. Therefore, finding a mechanical means of reducing fill efficiency would worsen the problem of maintaining high RPM power.) Would you need a big or small intake runner? (If you reduced cam timing with RPM, you’d have to increase port and valve flow capability to maintain power. However, as speed of the flow is diminished in the port, as would happen with larger valves and ports, the power drops in proportion to loss of gas inertia. The best flow is achieved between .2 and .6 Mach.) The lift couldn't be bigger at the end than the beginning could it? (Conventional mechanical cam designs would not allow the lift or duration to increase. One can conceive of mechanical solutions that would add duration and lift as RPMs increased. This could be achieved by using a cam with a split lobe on concentric cores that could be individually controlled to provide a broader and perhaps taller lobe as rising RPM allowed use of increasing gas dynamics in the ports. But such a system would be complicated and thus expensive. Additional lift can be achieved with a variable ratio rocker, but again the complication, weight and cost make such arrangements impractical. The future of the 4 stroke piston engine probably includes electric valve actuation. This will allow tailoring duration and lift to speed and load at an affordable price with acceptable electro-mechanical complication.)

Does a lower head angle (23, 20, 18, and 15) work better towards a larger or smaller bore? Does the lower head angle allow for higher or lower lift? (This is somewhat but not completely independent of bore diameter and does not have a direct effect on lift. The intention is to both tighten the chamber and provide better entry and exit angles between the ports and their valve pockets. The turn to and from the valves is very costly to flow efficiencies. The poster child for bad exhaust departure angles between the valve and port is the Nail Head Buick. The severe obtuse angle forced the engine to use considerable power to pump the exhaust around this corner. This reached the end of “Blow Down” theory where it was believed that the hot expanding exhaust gases would always find a way out of the cylinder. Ford’s 351 shares a similar trait with 50 horses being lost to the tight turn behind the exhaust valve. Unlike the Buick, it’s solvable with a milling machine and a hunk a aluminum where you hog off the exhaust ports and install a chunk of metal with a more graceful exit angle exhaust ports.)

Bogie

oldBogie,

Thank you for answering my questions! On a different board these same questions were treated as baseless and not well thought out. Thank you for respecting my interest in the subject matter.

In the interim I have come to understand that the size of the chamber is not as significant as the shape. I have aslo come across the MACH number stuff, but I am looking to understand more on its application. There's a book by C.F. Taylor that is on my list to order to help me understand this stuff.

Asyou probably can tell the goal is to pick parts that work well together for a specific application. I have decided that the application would be a street driven car where the functional operating rpm range is 2500-6500 rpm. As I just purchased and refurbished a 2-bolt 400 block I have almost a clean sheet of paper on which to write.

I am trying to pick parts, but it is difficult to do from catalogs and internet sites where you can't physically see the parts, let alone take measurements to verify that they will work for you. And their functionality is even more challenged by the fact that it is hard to determine what works together.

I am trying to determine what is the difference in performance at one set of specs vs. another. I have come down to heads, valvetrain and exhaust. Since there are numerous sizes on all of these things available I've tried start at the most limiting factor and work to optimize the whole thing based on the limits that it creates, but it seems to move around the system. At first I thought it was the cam (pcik a cam and a lift, get heads that flow in proportion to the lift, get pistons that support the compressed air/fuel mass etc.) My most recent hunch is that the system limitation for my application is going to be the long tube headers that are readily commercially produced. For a third gen camaro the most popular size is 1 5/8". If I could get some flowed, then I could determine the  conditions under which they at 6500 rpm, and build from there.

I have purchased a rotating assembly. With 64cc heads the compressions will be at 9.06:1. I may replace the pistons with something less than a 30cc dish to bring up the compression. I've been trying to determine an appropriate limit to spend on these parts so I could sharpen my focus on the ones in that price range specifically. The breakpoints seem to be $600, $1000, and $1400.

$600 gets me some brand new Vortec heads.
$1000 gets some better iron heads or some low level aluminum heads that don't flow any better than the Vortecs.
$1400 gets me some pretty good CNC'd aluminum heads more or less how I want them.

The difference in flow $600 to $1400 at .400 seems to be 24 cfm with the difference at .500 around 39. That's changing intake runner size of course, but that doesn't seem to be a lot. I understand speed costs, but there's no reason to just throw money at either. This tells me I am not buying flow as much as something else. Could that be velocity? How much difference in rpm range or power output is there in 39 cfm? To add the the fuel I have read that the engine doesn't always perform at the 28 inches of water, sometimes more than double, and that can have a more telling tale with which to compare given the similarities at this level. Again those numbers aren't published either.  

The range in intake runner was about 70cc but doesn't tell me anything about velocity because is is no description of shape.

So how can one make intelligent purchasing decisions with the information given? All of the heads work if given the right compression, cam, and displacement otherwise the companies would go out of business. How does one tell what to surround the ultimate flow limits with to maximize the area under the curve?

Thanks,
Jason

Ok Jason here's the rest of my response. Somewhere in here I again mention that you should be thinking of a university degree in mechanical engineering. The reason I say this is that your thinking about the fundamental questions rather than just bolting on parts 'cause everybody else bolts on these parts. I read your response to my first reply and will work on further comment to you.

If the intake velocity is as important as the flow, how does one compute it? How do you determine how much flow and at what velocity you need it to accomplish x horsepower or torque? (On the surface this looks easy to compute, but because of the compressibility of gases, thermal effects inside the engine, etc. it isn’t. The math gets real ugly real fast, if you’re a glutton for punishment start reading a few texts on Computational Gas Dynamics. Generally speaking, tests have concluded that .5 to .6 Mach within the inlet port is as fast as you want. Faster velocities lead to local sonic speeds which lead to shock waves that reduce flow in the port. The text book answer to this question looks something like Z = ((b/D)^2)(s/Ci a)) where Z= the speed of sound in Mach number: b = the bore in inches: D = the valve head OD: s = average piston speed: Ci = a flow coefficient that in my opinion is a really huge fudge factor intended to emulate volumetric efficiency: a = the speed of sound at the inlet air temperature and pressure. I’ve never been happy with Ci computations and I’m always wary of fudge factors that are so big they steer the outcome of the equation. Boy, will this statement tick off the Rocket Engineers that probably peer in here from time to time.

However, the back yard hot rodder can make some reasonably good approximations without having to get an engineering degree by assuming that at given RPMs the cylinder fills to 100 percent (or less, or more if you prefer) in some period of time determined by the duration of valve timing. This is my big fudge factor equivalent of Ci in the previous equation. The other big assumption possible with a compressible fluid is that the speed of the total flow is governed by the smallest area in the zone of total flow. This, for the intake will be the actual flow area of the valve opening. Typically a wedge chamber flows through about 1/3 the available open valve area. This is more true at high RPMs simply because inertia effects push the flow hard against the far wall of the valve pocket.   If you then calculate the average or smallest cross section you can now derive the length of the column of gas that must pass through a “pipe” to fill the cylinder. That will give you the speed the column has to travel at to fill the cylinder in the allotted time. Of course in the real world that speed may not happen, this is just an episode of #### and Jane playing with numbers.  Now keep in mind this is really a crude calculation, but it gives you an idea or a place to start. In the real world flow will mirror both the motions of the piston and valve and the resultant velocities of where the piston is in relation to degrees of crankshaft and movement in the bore. This ain't symetrical. So a cylinder of a 350 has to induct about 44 cubic inches of mixture. At 6000 rpm with a 220 degree cam, that’s 100 revs per second which is .01 second per revolution. There are 360 degrees in a revolution and the cam of 220 degrees would be 61% of that .01 second period which is .0061 second to induct 44 cubic inches. If the available area of the open valve is 1.04 square inches. 44 cubic inches divided by 1.04 square inches nets a column of mixture 42.3 inches long. To convert that into feet per second would be 42.3 inches divided by 12 is 3.52 feet which divided by .0061 sec nets a velocity of 577 feet per second. The speed of sound in the atmosphere at STP is about 1100 ft per second. So we have a Mach of 577fps/1100fps or .52 Mach at the carb inlet.  Of course in the real world Mach is affected by temperature and pressure both of which are not at STP inside an engine. Plus the speed of sound in a mixture of air and fuel will be different from air alone. Add to that, without a supercharger, it’s highly unlikely that an SBC is getting anywhere near 100 percent volumetric efficiency at 6000 RPM.  But without taking you into a class on computational fluid dynamics (most of which is as full of voodoo as thermodynamics), this is about as good a way to hack at the problem as any. In Charles F. Taylor’s book “The Internal Combustion Engine in Theory and Practice” he states that a 283 Corvette engine with a 1.72 inch intake at .4 inch lift has a Z of .521 Mach. This book is published by the MIT Press, . There are two volumes, Volume 1 is rather harder to find than Volume 2. Costs vary widely from 40 to 150 dollars per volume. If you decide to get an engineering degree and build motors for a NASCAR team in some obscure mountain village, I’d suggest you get a copy of both volumes.

In reality the gas velocity is variable both to valve lift and related piston speed both of which are close to parabolic functions. This statement holds truer on the start of the intake cycle than the end, but in a running engine is influenced by gas inertia to a very large extent. If we could start an engine from standstill to 6000 RPM in the first revolution of the intake cycle you’d see the gas flow start slowly as the valve cracked open and the piston began to move from TDC. The piston would gain speed and the valve would open further with maximum mixture speed occurring around 90 to 110 degrees from TDC. At this point the piston begins to slow to BDC and reversal for the compression stroke. However, the column of incoming mixture has a lot of inertia at this point and a greater ratio of flow per amount of valve lift is occurring than when the engine first started this cycle. So if you plotted flow against crank degrees, there would be more flow as the crank proceeded in spite of the decreasing valve opening. A late closing valve takes advantage of this inertia which at high revs is so significant that you can hold the valve open while the piston is 40 to 50 degrees into the compression cycle. Now our magic engine has gone from standstill to 6000 RPMs in less than 2 revolutions so we can see what happens to the intake on the next open cycle.

If i may take more of your time, bogie, these questions have also gotten me thinking, first off, in something a little more concrete, how is this carb inlet velocity measured on a dyno? and second off if ports flow best at 20%-60% Mach then how do these numbers effect the velocity numbers aimed for throughout the lift curve when testing heads on a flow bench? also if a engine is tested and the port velocity at peak HP is far from optimal how is this adjusted, and how can it effect how the engine functions? is it adjusted by simply valve size and intake tract volume?

P.S. if you have any more information on the subject of 351 Ford exhaust port modifications and you were talking of earlier it would be greatly appreciated, and that how general is that 50 HP quote?

One more thing, please tell me about your past history, where did you go to school and what did you major in (dumb question), why did you choose to help people on something like the CHP board? is there any other sources that i can find of your work to read? and why dont you answer questions at the Hot Rod board?

                                                                               Thanks for your time

If I didn't have a job that requires 24/7 availability I would finish the Mechanical Engineering Technology degree I started. But alas $1000 heads aren't bought on drive-thru money at Burger King...

I have Lingenfelter's book. It helped me determine what I wanted to do. If left me hangin' on how to do it. I have some of Vizard's books also. In each section a statement or two more would have helped a lot.

I think the authors are cagey about really lettin you in on the deal because it is so easy to be proven wrong 5-10 years later. All of these books could use an update. Guys are doing more with different parts, and different companies.

Jason

Bump, id like to bring my questions back up if you missed them bogie

Message: 1041.8

Velocity on a dyno is measured with what essentially is a pitot tube which is similar to how airspeed is read for an airplane. It makes use of ram pressure of the passing airstream. There are also several computer programs that compute velocity and volumetric effiency through Mean Effective Pressure (MEP) and other equations.

Velocity throughtout the lift curve is variable to a lot of things but what you're searching for is probably best answered by amount of lift against piston postion in the bore and RPM.

One needs to be careful of flow bench data. There's a lot of pitfalls. Basically flow bench data is comparative and has little or nothing to do with the real world. I say that based on two primary considerations. 1) no intake manifold. The intake has a lot of effect upon how the mixture enters the valve pocket. If the intake is forcing the mixture flow to the runner wall favoring the center of the cylinder, this as much as anything is an inertia effect, the mixture has little chance to correct itself to the port wall adjacent to the cylinder wall prior to entering the valve pocket. This significantly reduces swirl inside the cylinder and has bad effects on cylinder fill as well as distribution of the fuel with the incoming air. Like how many times do you operate your engine with a clay or plastic airfoil that provides symmetric air entry into the port. 2) Your engine is not a fan, it's difficult to emulate flow effects as related to valve opening, piston position and speed related to posiiton and RPMs with a fan. There is a powerful and variable relationship of port velocity to valve opening, piston location and piston velocity (RPM). Flow at a given lift less than maximum lift will vary not only by the amount of opening but by whether the piston  is early, mid, or late in its stroke because specific piston speed is variable by location as the crankshaft rotates and the mixture also develops inertia which tends to keep it flowing in fairly high numbers at during late cycle lowering lift against a rising piston if the engine is turing sufficient RPM.

If the designer misses on port or valve size, hopefully this will be discovered at test and rectified against the intended use of the engine. Basically valves and ports that are too large make an engine lazy in the lower RPM ranges but are necessary for high RPM feeding, Ford again, only because more than Chevy they tend to go over the top with "good ideas". The tunnel port and Boss 302 really suffered from insufficient port velocity. You had to beat the daylights out of the engine to get any power out of it. If you missed a shift you were in big trouble as the engine had no grunt with which to pull its revs up. You could actually miss a gear and stall at at 5000 RPM.  The reverse is true for small ports and valves. Small ports and valves will emphisis low and mid range torque at the expense of top end power. The TPI engines are a recent example of this, they have lots of midrange grunt but crap out by 5000 revs as at those speeds they just can't get a breath anymore. Now for a street engine hooked to an automatic this is pretty great as an automatic really likes down low torque, but the same engine with a T-56 will drive you nuts because there's no top end power that you can use with your ability to control the upper rev range with the T-56.

The optimum solution would be a design that has variable cam timing and port sizing such that low speeds had a short durtation valve event and small ports. As RPMs went up the valve timing and port size would increase with the engine's demand for more mixture. This would provide a nearly uniform high velocity in the ports which improves most all operating characteristics. But the solution is obviously complex, expensive and failure prone, so it isn't done. So we all live with a compromise solution at one end of the operating envelope or the other.

The 351 Cleveland solution is to mill the exhaust ports off back to and putting a machine surface on the exhaust side of the head casting. A block of aluminum 1.0 to 1.25 inches thick is machined square as it will have to seal both to the header gaskets and provide a gas tight fit to the head, this is why the head casting was machined. New ports are machined into the block of aluminum which pickup the factory location at the head and raise the outlet at least 3/4s  to 1 inch or higher if you have chassis space for the headers. An 4 open sided cones are milled along side the port to provide access for the sparkplugs. With extremely careful measurement, the floor of each exhaust port is machined to accept a locating pin of 3/8s to 7/16s inch diameter. A matching hole is carefully machined into the aluminum adapter. The aluminum adapter is drilled and tapped to accept the headers. It is located on the head, clamped in place and a machine hole large enough to pass the threaded portion of a 5/16th to 3/8s shouldered bolt is made. This should be done in at least 5 posiitons along the length, but sometimes 3 to 4 is all you can get.    The hole extends into the head within the rocker box portion of the casting. You can tap the casting, but you really need a nut as the casting is none to thick in this area and is not intended as a place to secure a bolt. Spot facing the inside is a good idea to help insure clamping loads are evenly distributed. Grinding a reasonably thick hardened washer helps spread the clamping loads which reduces the possibility of cracking in this area. One slick alternative way of doing this if you mount the header a the right heigth to pull it off is to bring a stud that passes from inside the rocker box with a nut and washer, as previously described, thru the adapter plate in the location of the upper exhaust header flange attachment. Then provide a countersink for a nut that clamps the adapter plate and is located under where upper  headers' nut clamps it to the plate. Of course the adapter plate needs to have 5 head bolt holes drilled into it the exhaust side head bolts. A hardened washer should be used under the bolt or nut if you use a stud. Put it together with a high temp sealer between the head and the adapter as well as a sealant on the bolts to keep oil from migrating along them from the rocker box.

Me, I graduated in different levels of mechanical and electrical engineering from some well known schools in southern California, Houston, Texas and Cambridge, Mass. I answer questions on the CHP board because my first love is Chevy, I've also done a lot of Fords and a few MoPars over the years. I occasionally show up on the Sallee Chevy board. These can take a lot of time (I did take auto shop but not typing in high school) so I just haven't got to Hot Rod because of time. I got into this rather accidently after realizing that a lot of guys didn't have the engineering backgrounds to make the best of their situation, so I jumped in to try and provide some "drive by wisdom" based both on the practical aspects of having done this stuff for a lot of years and being able to provide some engineering background of why we're doing what we're doing.

I hope this stuff helps somebody at least some of the time.

Bogie

Even though I don't post very often, I don't think I'm the first person to learn an absolute ton from all the things you've posted. Thank you!

Brian

Bogie,

Your wisdom and knowledge have given me a belittling (sp) sense of how much I don't know. I once had aspirations of being a nascar engine builder. Such asprirations have passed, but I still love to learn as much as possible about the hows and why's of high performance and or efficient engine design/building.  The relatively basic knowledge of small port = good torque down low sacrificing high end power vs. large port = high end hp, sacrificing low end torque... I knew. Thanks to you and your engineering degrees (not to mention your ability actually write well) I now know more about "why". 

Now I'll throw another question at you. My last motor was an 11:1 street/strip 406 with AFR 195's and 5.7" rods. The new motor is a 13:1 (ish) 406 with ported bow-ties and 6" rods. I realize shorter rods have a lower rod/stroke ratio, and create higher piston speed, which gives you more torque down low, and the longer rods lose some low end torque due to lower piston speed, but make up for it on the top end, and put less lateral force on the cylinder walls. The Question is: Is there an ideal rod/stroke ratio for an acceleration (drag race) engine, or is it all about sacrifices?

Dave

Regarding the pros/cons of rod length and ratios, here's an interesting web site...

http://www.iskycams.com/techtips.php

Give me a few days and I'll give you guys a formula to compute piston speed by crank degree. This lets you see how rod length and stroke affect piston speed.

A P.S. one of the best natural engineers I ever met dropped out of school in the 9th grade. But he's one of those guys that looks and just knows, it's a talent or a gift from God however you look at these things every much as Elvis' muisc talent.

Bogie

Interesting site, I fully concur with the rod length evaluation. At best this is pushing about a 1% difference and spending a lot of money to do it.

Now don't get me completely wrong, if someone is scratch building a whomper stomper race motor and needs a set of custom rods for strength reasons, there's no harm in spending the money on a set of 6 inchers versus 5.7 inchers. But if it's a stock through well hopped up street 350 with a set of sufficiently strong 5.7 rods in it, the money spent on a set of 6 inch rods with the anticipation that there's some major power to be gained against the 5.7s, I'm afraid this rather substantial amount of money would be better spent elsewhere in or on the engine.

Bogie