50 degree valve seats in modern hot street builds: is it finally time?

[CGT]

Where the flow demand is highest 50 deg seats move more air.
Pretty simple.

Looking at average flow is BS, look at average flow between mid lift & .100" past peak lift to get a better idea.

Move more air at peak demand & you build more inertia in the charge, that improves fill at the end of the induction stroke.

Well said.

[hoffman900]

Low-lift flow has no good effect before TDC, only after BDC.
Before Top Dead Center the piston is moving UPWARD, pushing exhaust gases out of the cylinder; this is during the EXHAUST stroke. Before the intake valve opens, the volume and pressure of those gases in the cylinder determine the amount of reversion that will be present during overlap, and the amount of time needed for the system to recover and start flowing clean air and fuel.
When the intake valve opens, it is generally into a higher pressure area than exists in the intake port, burned exhaust gases clog the intake port, until cleaned by the downward intake stroke, after top dead center. The amount of reversion blocking intake flow determines when the port is opened up for clean air flow; this reversion is worse as RPM increases.
The less reversion, the sooner clean air/fuel flow begins and the higher the intake port velocity and inertia ram.
It is more important how the intake port REVERSE-flows at low lift, vs how it flows normally at low lift. The only time there is air flow at low valve lifts into the cylinder is ABDC, right before intake valve closings. The port velocities are generally very high at this point, due to inertia ram.
I am almost always more concerned with airflow above .250" valve lift than any below .250". The area .250" to .600"-.650" is the part that sets up inertia ram to fill the cylinder, and that's on the opening side of the cam.
All cams that I design have a little more duration on the closing side, to enjoy the inertia ram, and a seating rate that does not encourage valve bounce.

The words "Low-Lift Flow" are relative; what is low-lift in a .450" max valve lift cam is different from that of a .750" max valve lift cam.
However, whatever valve lift occurs BEFORE TDC is bad. The piston is moving upwards, pushing burned exhaust gases out the exhaust port, not moving downward, sucking intake charge into the cylinder. The only time intake charge enters the combustion chamber BTDC is in supercharged or turbocharged engines, when it starts entering with only .001" valve lift.
The intake portion of the overlap, the part BTDC, exposes the intake port to exhaust gas reversion, both pressure and volume of exhaust gas, until TDC, when the piston reverses direction and starts down on the intake stroke. The reversion that exists, and blocks the intake charge, must be cleaned out before clean air and gas start in. The less reversion present, the quicker the reversion is cleaned out, the sooner the good flow starts. The sooner good flow starts, the higher port velocity is at every degree of crank movement, and the better the cylinder filling.
The more high-lift duration a cam has, and the higher the inertia ram velocity is, the more charge gets into the cylinder.
Although low-lift flow helps fill the cylinder at the end of the cycle, I always prefer having bad reverse-flowing ports to hinder reversion vs having good normal-flowing ports, in terms of low-lift flow. Preventing reversion at the start is a lot more helpful than a few more CFM at intake valve closing,IMHO.

UDHarold

The UH profiles are my work, the H profiles are legacy Lunati grinds that they have made for years. I think they have eliminated the ones that over time do not work very well. My UH designs work as good as they ever have.
All my cams are unsymetrical, they have been since 1977.
There are a number of reasons why unsymetrical cams, done right, have wider power bands than symetrical designs.
1.---Longer power stroke on the exhaust, puts more torque into the crank. It must open fast so that it has the proper lift by BDC to minimize pumping losses on the exhaust stroke.
2.---Exhaust profile must have enough high-lift area to clean out old exhaust gases on the top-end. Residual exhaust gases are the reason for the turnaround in the horsepower/torque curve at peak power.
3.---Delayed opening of the intake allows both cylinder pressure and residual exhaust gases to drop vs symetrical intake openings. This reduces reversion, with reversion being our name for both the volume and pressure of residual exhaust gases when the intake valve opens.
4.---Reduced reversion allows earlier intake flow on the intake stroke, and an earlier increase in port velocity. The higher the port velocity, the higher the rate of cylinder filling on the intake stroke, and even on the compression stroke.
5.---Increased high-lift area gives more time for cylinder filling. The effective cam curve goes from about .250"-.300" on the opening side until the valve seats itself. The part of the curve that does the majority of work is from TDC to about 75* ATDC, wherever the piston is at max velocity. Even though the piston is slowing down after 75* ATDC, the inertia of the air/gas charge continues to increase the velocity, and ram-charge the cylinder.
6.---Because the cylinder should still be filling when the valve seats, it is important to have a closing velocity that prevents valve bounce. Otherwise, compressed charge is fired back into the port if the valve bounces off the seat, and terribly disrupts the next intake cycle.
The name of the game is putting more charge into the cylinder at all RPMs.
This works whether in hydraulic, solid, hydraulic roller, or solid roller versions.

At least, this is the way I see it......

UDHarold

Harold got this (and he wasn't the only one) 40 years ago, and I doubt he ever looked at a pressure graph.

[GARY C]

Thanks for the responses.

[plovett]

Regarding the UD Harold quotes: Do they imply that the last exiting exhaust does not ever pull on the opening intake valve? If so, why does valve overlap help in some cases?

paulie

[PRH]

Where the flow demand is highest 50 deg seats move more air.
Pretty simple.

Looking at average flow is BS, look at average flow between mid lift & .100" past peak lift to get a better idea.

Move more air at peak demand & you build more inertia in the charge, that improves fill at the end of the induction stroke.

Except when they don’t, as in the examples I posted on page 8 of this thread.

The steeper seat angle doesn’t guarantee greater high lift flow in all heads.

[In-Tech]

Correct, but as has been stated many times in this thread, it all depends.

If you look at the cutaways I did on the newer GM LT heads you can see those heads beg for a really good blended 55(or more, race environment). Plenty of other heads, not so much.

[hoffman900]

Regarding the UD Harold quotes: Do they imply that the last exiting exhaust does not ever pull on the opening intake valve? If so, why does valve overlap help in some cases?

paulie

Obviously, no one can speak for Harold, but he was a pretty prolific poster and left his thoughts scattered around across several posts and websites.

Harold has pretty much said that of course it does, but it's a byproduct of everything else.

Taking a step further, Harold designed his lobes with the above outlined. That flow doesn't happen until after TDC, but for dynamic reasons, of course the valve is going to open before TDC. Think of that post as the 32,000ft view on his design theory. You can't open the valve at TDC and expect it to be open far enough where you need it to. His point was whatever overlap there was after picking the correct duration / lift / centerline for a given combination is all that is needed and he never calculated overlap unless a customer asked for it. If you look at pressure diagrams (measured and simulated), you see exactly what he was getting at, and imo, he was pretty spot on in his thoughts. Of course valvetrain dynamics and everything else complicates this, but basic level, yes.

The point being, whatever overlap you have with proper events is likely enough (or more than you need, hence you need other ways to bandaid it).

I posted this a month ago in this thread, but I'll post it again. This pretty much has all the information anyone on here could want.

Here are a bunch of thoughts by many. I took time and typed up select sections of both Billy and Dr. Randolphs interview and lecture. It all interacts together. You should all read and re-read all of them.

Remember, think of everything in terms of pressure differentials and not static flow on a flow bench. Flow occurs due to deltas in pressure between the ports and the chamber/cylinder.

To quote Billy Godbold on the exhaust side:

"On the exhaust side, some of the low intensity camshafts are very header sensitive. If you look at most of our catalog camshafts for for drag race applications, the intake lobes will be in that typical, you know, 30* to 36* major intensity, but you the exhaust lobes will be, you know, 36* to high 40*s major intensity. And People go why you use these soft exhaust lobes? The exhaust valve is lighter. Well Bernoulli starts playing a point. When you start going with really fast especially exhaust closing... when the piston is coming up it's pushing the last little bit of exhaust out of there, so exhaust velocity is really high, so if you put a long tail on the closing ramp on the exhaust, you'll get more time where you have high velocity of the exhaust leaving the chamber. If you think the back of it, the pressure in there would hurting you more than the velocity would help, but because the high velocity of the exhaust valve seat may be very close to the intake valve seat, you can actually get a low pressure right around that exhaust seat. This starts to pull some intake in early. If you can start moving intake in the chamber, say 30* - 40* BTDC instead of having to wait for the piston to start moving really fast 50,60,70* ATDC, it gives you a much longer intake stroke. Probably the reason why you guys all remember when people put the longest rod you can in there, then you look at a typical Pro Stock, it's a very very short rod, so looking at that type setup, the reason the short rods work so well.. you reach peak piston velocity sooner, so it gives you more time to pull on that port. So thinking about sometimes a softer exhaust can definitley give you more time to pull on the port, but generally on the intake side, I almost always see a faster camshaft , the camshaft is quicker off the seat gives you more area under the curve, just gives the port more time... I don't often see where I've seen the slower cams work better. if you think about it on the opening side, you may see something on the opening side where you can get more duration in on something that's limited on P-V by going to a softer opening..... on the closing side it's kind of interesting because if you look, one of the things that some of the NASCAR guys used to look at back in the '90s and 2000s, was you tried to get the earliest intake closing you could, but maximize the lift at BDC on the intake. That BDC lift, when you start to get to peak power and beyond, how much air is still coming into the chamber at BDC... for a given intake closing point, the more lift you can have at BDC, the more power you're going to make. Almost always.".

https://www.youtube.com/watch?v=whmOxK4XDYQ&t=1225s

Dr. Andrew Randolph on a 2010 NASCAR engine:

"You've got to start opening the exhaust valve before you get to BDC. So we're going to start opening them way back here, so you're going to be decreasing the pressure in the combustion chamber because now you're starting to exhaust what you've got, so you're not pushing on the piston as hard, so you have what I'll call and expansion, an exhaust expansion loss there, that's a time loss. But also on the other because you can't open it (the valve) fast enough, once you start the exhaust stroke, you're gonna' have a pumping loss because you don't have enough valve area, so you've got losses. You've got an exhaust expansion loss here and you've got a flow pumping loss there. So you get hit by time losses on both sides. Normally you're going to make your best power if those two loss mechanisms are balanced. They're each going to be about 20hp."

https://www.youtube.com/watch?v=rBZCnG1HwDM&t=1570s

The late Harold Brookshire:

Well, I may not be an expert on this, but for over 30 years I have designed only cams---not intake cams or exhaust cams, but just cams.
In most cases, I use the bigger ones as the exhaust and the smaller ones as the intake. In certain application, that may be reversed.
All my designs are always unsymetrical, with a fast opening side, and a gentler closing side.
I have not tried any other combination, other than in my thoughts.
My thoughts keep coming back to the way I have been doing it. The various Laws of Physics keep bringing me back to the way I have done it for 30 years.
Should I open exhaust valves any earlier and slower, and put less torque into the crank?
Should I shut intake valves any earlier and quicker, and cut off cylinder filling earlier, and possibly bounce the valves off the seat?
I keep coming back to the way I have done it for 30 years.
But I keep looking, and thinking.......

UDHarold

It depends on what you mean as "the best attributes for driving".
If we look at factory cars, that are made to have the "best attributes for driving", we see that they generally have wide, or very wide, lobe seperation angles, for good idle and vacuum, and a flat torque curve.
No matter what the duration or LSA, I rank the importance of valve timing in the following order:
1ST(& 2ND!)---Exhaust opening(which occurs first) and Intake opening.
3rd---Intake closing.
4th---Exhaust closing.

The exhaust opening point will govern how the intake fills, no matter where the intake opening occurs. It does this by cleaning out the cylinder of burned exhaust gasses, which determines the volume and pressure of whatever residual exhaust gasses are in the cylinder when the intake valve opens.
The amount of exhaust gasses in the cylinder at intake valve opening cause reversion---less exhaust gasses, less reversion, more exhaust gasses, more reversion.
The amount of reversion determines cylinder filling on the intake stroke, less reversion, earlier port flow, more reversion, longer port recovery time.
It goes on from there, but I have posted this a number of times, and I do not want to let the cat out of the bag too many times........

UDHarold

Digger,

The valve events needed for an engine with a 500 to 4500 power band are completely different in what they affect compared to an engine with a power band from 4500 to 9000, and to all variations in between.
Exhaust events are compromised by the need to extract all the torque possible before opening the exhaust valve, and the need to extract all the burned exhaust gases possible before the intake valve opens. After TDC the exhaust valve closing, if longer than needed for low-speed dynamic control, allows excess exhaust reversion to enter the intake from exhaust gas pulsations. This is why I claim the exhaust valve closing is the least important of the four events; It only affects low-speed power. at higher speeds the other events are much more important.
The exhaust cam gets in a race all its' own, a race between how much torque can be made by delaying exhaust valve opening, and how much exhaust can get out before the intake valve opens. How much exhaust gas is out of the cylinder affects how much intake charge can get in.
Peak Horsepower occurs when the exhaust can no longer effectively clean out the cylinder, and leaves residual gases in the cylinder, hindering intake filling.
Intake opening before TDC is to minimize exhaust gas reversion into the intake port, which aids in cleaning out the port entrance after the piston starts down ATDC. The sooner the intake can start flow into the cylinder, the higher the intake port velocity can be, the more inertia ram you can develop in the intake port, and the longer you can fill the cylinder ABDC.
As long as intake charge is still entering the cylinder, it doesn't matter too much when you shut the intake valve. I have had this proved to me for 30 years, with unsymmetrical cams still providing good bottom-end torque, even with delayed intake closings.
I'll talk more later, now it's Honey-Dew.....

UDHarold

Mark,

Look at any dyno engive curve; whether stock, hot street, street-n-strip, bracket, full-race, oval track, it makes no difference. They all show a peak torque point, then decreasing torque and increasing horsepower, then peak horsepower and a tumbling torque curve.
There are all sorts of theories to explain what is happening; Here is mine.
Peak torque is where peak cylinder pressure occurs, and after this point, torque starts decreasing, yet the horsepower curve is still going up. This is because horsepower is a measurement of torque and rpm, divided by a conversion factor ---5252---.
As long as the torque curve is dropping at a slow rate, the math of the equation keeps the horsepower number going up. But at some RPM point, the torque curve starts dropping more rapidly, and the equation shows this point as peak horsepower.
What happened at that RPM and afterwards, that made the torque curve drop? The torque is just a product of how much charge you got into the cylinder, etc.
You had to get less air/fuel charge in the cylinder.
I maintain that at the point of peak power, the exhaust cam became unable to get as much burned exhaust gases out of the cylinder as before, and some gases were left in when the intake valve opens, in excess of the normal left there. The cylinder had some residual exhaust gases in it, and could not accept the normal amount of clean air/gas mixture. Not as much got in, the engine produced less torque than normal, and the horsepower curve tipped over.
The next cycle, more residual exhaust gases were left in, and less clean air and gas got in.
The next cycle, it got worse, and the horsepower and torque curves are both dropping rapidly.
Wider LSAs and bigger exhaust cams promote better top-end torque and horsepower curves, a la ProStock cams.
At least, this has been my theory for about 30 years.......

UDHarold

Procision Auto,

For about 31 years I have had engine builders who have tried the same dual-pattern cam, on 2 different LSAs, and installed on the identical ICL. The intake timing events were identical, only the exhaust timing events changed.
The wider LSA, with an earlier opening and earlier closing, always made less bottom-end, and more top-end. The cam blew the bottom-end torque out the exhaust, and cleaned the cylinder out better on top-end. It would have a wider power band.
The tighter LSA, with a later opening and a later closing, made more bottom-end/mid-range torque, and dropped the power curve faster after peak horsepower. When it came into power, the torque curve would rise faster, and then drop faster after peak power.
It is my understanding that this is common to all cams.
One caveat: At very low-speeds, between idle and the main torque curve, wider LSA cams make a little better power that tight LSA cams. I believe this is caused by the amount of exhaust overlap ATDC. The wider LSA cams have less, the tighter LSA cams have more. Because of the time in milliseconds involved, the more time the exhaust valve is open ATDC, the more time there is for exhaust pulses to mess up engine airflow.
This is why cams with wide LSAs idle as good as they do.

UDAHarold

Overlap should never be figured for .050" lobe lift/duration.
Why?
Because the valve opens AT THE SEAT DURATION, not at .050".
When I was Comp Cams cam designer, I wrote a full-page ad for National Dragster that asked the question, What is the most important degree in your camshaft?
It is the degree BEFORE the intake valve opens!!!
If we could determine the amount of exhaust gas left in the cylinder, and its' pressure, we could predict how the intake lobe would fill the cylinder.
Playing a little 'numbers' game, if you have 2 cams, with the identical durations/lifts, either single or dual-pattern, but on 2 different LSAs, and both identical intake lobes are installed on the identical intake CL, guess what?
The cams have different power curves. They have the identical opening and closing points on the intake. Only the exhausts are different timing points.
Yet the cylinder is filled differently by the same intakes on the same ICL.
It is because of this difference of exhaust volume and pressure that exists before the intake valve starts coming off the seat.
The further before TDC that the intake valve opens, the higher this exhaust volume/pressure is, and the more reversion enters the intake port.
Air does not begin entering the cylinder until the piston has quit PUSHING the exhaust out the cylinder, the piston reaches TDC, and then begines the intake stroke. The more exhaust gas that has blocked the initake, the later clean air flow starts in the intake tract.
Using this theory, I beat Billy by 10 BHP at peak in the Dodge Hemi Truck engines, on my FIRST SHOT. The cam designs I sent in as a resume.
The engine builder sent me a check for $1000, he liked them so much.
My 10,500+ ProStock cam has 88 degrees of overlap at .020". It opens at 41 BTDC(Intake), closes at 47 ATDC(Exhaust). This is seat duration, not .050" duration.
It runs in the mid-6.60s, and has been nr 1 P/S qualifer a lot this past fall.
It was my first shot a a ProStock cam in over 20 years, and not a bad first try.
I think I'll stick around and keep doing cam designs the way I do.

Back in my old UltraDyne days, my most popular roller cam was the SB288/296R. It was 288/296 at .020, 255/263 at .050, 176/183 at .200, and .626"/.626" gross valve lift with 1.5s.
With 83 degrees of overlap, it dominated the NDRA for 2 years, driven by Jeff Purvis. IVO 43 BTDC IVC 65 ABDC EVO 76 BBDC EVC 40 ATDC
With 80 degrees of overlap, it won a LOT of races for 20 years.
IVO 41 BTDC IVC 67 ABDC EVO 77 BBDC EVC 39 ATDC
With 76 degrees of overlap, it won 2 of the 3 UltraDyne national wins in the Winston Amatuer National Championship. IVO 41 BTDC IVC 67 ABDC EVO 81 BBDC EVC 35 ATDC
These overlap numbers mean that this cam is either a Street/Strip cam, a Race cam, or a Pro Competition cam, although it is always the same cam, on different LSAs.
Of course, the overlaps are different. Did I decide on the overlap first, then figure out what the LSA should be, or did I decide on the LSA, then calculate the overlap for the customer?
If you really get into the study of overlap, then you will realize that the area under the overlap curves is more important than the overlap number itself. Obviously, rollers have more Area under the curve than flat tappets, who have more area under the curve than hydraulics, even though the overlap is the same in all the cases.
The area under the curve on the intake opening portion of the overlap has the most overall influence on cylinder filling.
The area under the curve on the exhaust closing portion of the overlap influences the engine mostly at very low engine speed, and it is mainly a function of how many milliseconds after TDC the exhaust valve is open, to allow reversion into the cylinder.

30-60-60-30 270/270 105 LSA 60 overlap
30-70-70-30 280/280 110 LSA 60 overlap
30-80-80-30 290/290 115 LSA 60 overlap

Normally, one would expect these 3 different cams to produce different results, but they all have identical overlap, so??????

Low-lift flow has no good effect before TDC, only after BDC.
Before Top Dead Center the piston is moving UPWARD, pushing exhaust gases out of the cylinder; this is during the EXHAUST stroke. Before the intake valve opens, the volume and pressure of those gases in the cylinder determine the amount of reversion that will be present during overlap, and the amount of time needed for the system to recover and start flowing clean air and fuel.
When the intake valve opens, it is generally into a higher pressure area than exists in the intake port, burned exhaust gases clog the intake port, until cleaned by the downward intake stroke, after top dead center. The amount of reversion blocking intake flow determines when the port is opened up for clean air flow; this reversion is worse as RPM increases.
The less reversion, the sooner clean air/fuel flow begins and the higher the intake port velocity and inertia ram.
It is more important how the intake port REVERSE-flows at low lift, vs how it flows normally at low lift. The only time there is air flow at low valve lifts into the cylinder is ABDC, right before intake valve closings. The port velocities are generally very high at this point, due to inertia ram.
I am almost always more concerned with airflow above .250" valve lift than any below .250". The area .250" to .600"-.650" is the part that sets up inertia ram to fill the cylinder, and that's on the opening side of the cam.
All cams that I design have a little more duration on the closing side, to enjoy the inertia ram, and a seating rate that does not encourage valve bounce.
An old UD example from 1980:
The 288/296R6 roller
The 288 intake cam opened like a 282* cam, with a 282*s reversion. At .050" it was a 252* cam on the opening side, at .200", 1* bigger than a Crane TR-260/4167, same aprox lobe lift.
On the closing side at .200", still the same. At .050", equal to a 258* cam, at .020", a 294* cam.
It was the most successful UD cam for 20 years, and is still a potent cam today, almost 27 years later.
I have let everybody know this data for about 20 years now......

UDHarold

The words "Low-Lift Flow" are relative; what is low-lift in a .450" max valve lift cam is different from that of a .750" max valve lift cam.
However, whatever valve lift occurs BEFORE TDC is bad. The piston is moving upwards, pushing burned exhaust gases out the exhaust port, not moving downward, sucking intake charge into the cylinder. The only time intake charge enters the combustion chamber BTDC is in supercharged or turbocharged engines, when it starts entering with only .001" valve lift.
The intake portion of the overlap, the part BTDC, exposes the intake port to exhaust gas reversion, both pressure and volume of exhaust gas, until TDC, when the piston reverses direction and starts down on the intake stroke. The reversion that exists, and blocks the intake charge, must be cleaned out before clean air and gas start in. The less reversion present, the quicker the reversion is cleaned out, the sooner the good flow starts. The sooner good flow starts, the higher port velocity is at every degree of crank movement, and the better the cylinder filling.
The more high-lift duration a cam has, and the higher the inertia ram velocity is, the more charge gets into the cylinder.
Although low-lift flow helps fill the cylinder at the end of the cycle, I always prefer having bad reverse-flowing ports to hinder reversion vs having good normal-flowing ports, in terms of low-lift flow. Preventing reversion at the start is a lot more helpful than a few more CFM at intake valve closing,IMHO.

UDHarold

Clint Grey:

No matter what the duration or LSA, I rank the importance of valve timing in the following order:
1ST(& 2ND!)---Exhaust opening(which occurs first) and Intake opening.
3rd---Intake closing.
4th---Exhaust closing.

UDHarold

I would have to agree with that order based on our data, at least for performance applications (we don't have too much data on lesser applications). Of course the intake and exhaust designs can always be changed to make bad things look better, but only to a point.

Clint Gray
TFX Engine Technology Inc.
www.tfxengine.com

Here is one from Isky's website:

Specifically, when the intake valve opens some 40 or more degrees before T.D.C. at the end of the exhaust stroke, very little (virtually no) exhaust gases remain in the cylinder. The piston is in the vicinity of T.D.C. (only .425" down the hole @40o BTDC - on a typical 350" Chevy with 5.700" rods) and no appreciable threat is posed to the forthcoming intake charge.

Well that's quite a statement to make. There is a very appreciable threat at IVO, more so at some engine speeds than others and very likely to occur if all the i's are not dotted and T's crossed.

At IVC there doesn't even have to be any reverse flow at the valve to cause reversion in the intake. The air is all stacked up at the valve at IVC and expands back up the intake. Happens even if flow is still entering the cylinder at IVC.

This is the first time I have seen this in writing. Excellent!

That's the way we always look at it. Exhaust first. Why start tuning on something that is messed up from the get go, and spend time chasing your tail. We use the pressure data to get a good exhaust situation over the important rpm range first, then everything else can be tuned properly within this rpm range knowing that the exhaust residual is a minimum, etc. When everything is pretty much as good as its going to get we then go back to the exhaust and look at further widening (if it matters) the rpm range where the exhaust works well, without messing up things over the critical part of the rpm range.

Spending a lot of time tuning an engine, then getting around to looking at the exhaust as a final tweak, is entirely backwards.

Clint Gray
TFX Engine Technology Inc.
Combustion/Intake/Exhaust Pressure Analysis Systems
www.tfxengine.com

Darin Morgan:

Thanks for all the great info. Which has led me to more questions.

Is the "best" disch coef the highest that you can get it? Can it be too high causing something similar to the sonic choke at the ssr? Is the average over the range .200"- max lift most important or are some valve lift ranges more important?

What differences might be expected in other engine components (cam, manifold etc) between two engines one with a good D.C. head and one with a great D.C. head? I would think a better DC would like a little more int. duration since the air at the valve is moving faster. If the DC is not as good the intake charge should reverse directions sooner at the end of the intake cycle.

Thanks,
Rick

(1) Is the "best" CD the highest that you can get it? YES

(2) Can it be too high causing something to sonic choke at the SSR? Yes it can choke but NO it cant Sonic choke. There is no such thing as Sonic choking in an intake tract. They operate at a maximum of about.5 to .55 mach 600+ft/sec, which is far below Mach. If the mean velocity in the intake tract where to exceed mach .55 the power would drop off dramatically. Sonic only occurs on the exhaust side and usually at lower lifts in super charged engines.

(3) The most important lift range is form start to finish. Attempting to place emphasis on one particular part of the over all lift curve is a mistake. Can you give up about little on the bottom of the curve to make it flow more up top? Yes but not much! You want the highest Discharge Coefficient you can get. Period. I think you are looking at this too much from a flow standpoint and not an over all efficiency standpoint. You want the window area exposed to the cylinder as efficient as possible. Speaking hypothetically, if your CD is higher, you could use a cam with less duration not more. Think about it.

As far as reverse flow and the reversion characteristics of a particular design during overlap or at the end of the intake cycle, well, that's a whole new thread within itself. I don't want to be vague about this but it's a pretty involved subject in its own right. I measure cross flow by pulling through the intake port and out the exhaust port on a chamber with a simulated dome and piston and see how much it can flow at each lift increment. I measure reverse flow (out the intake port) by pressurizing the exhaust port and reversing the procedure.

I would like the others to weigh in on this one.

I have a question, or a comment on reversion in a high speed engine. Is there really enough "time" for reversion to actually occur when running at speed?My impression was that intertia and wave/pulse tuning would all but eliminate almost anything but a "stall" at one point. I haven't investigated it enough, but what do you think?
Shawn

The compression and rarefaction waves present in the intake and exhaust tracts during overlap and other parts of the 4 stroke cycle are tuned in such a way as too accent the gas exchange process and increase VE ( trap mixture) BUT they only do so over a given amount of time and over a given range of rpm. That range of rpm in which the positive ( compression) and the negative ( rarefaction) waves work together is about 1500 to 2000 rpm. The reason for minimizing reverse flows is to not only broaden the range of time the waves are usable but to also manipulate the tuning factors of camming and manifolding to increase VE. Usually when these Antireversion steps are taken an instant increase in both low end torque ( below peak TQ ) and top end power ( above peak power) can be seen even though no other steps have yet been taken to optimize the situation. Lets say for instance you minimize the reverse flow of the intake port by some means and not effect the inlet flow. You now have an engine that will hang on a little longer past peak power when the waves are decaying in amplitude and are decreasing in there ability to feed the cylinder because you will revert less mixture back up the intake tract.

You asked the question,"Is there really enough "time" for reversion to actually occur when running at speed?

Yes there is but Its not as much a TIME factor more so than a factor of multiple dynamic events starting to work against you instead of for you! By trying to prevent reverse flows, your just trying to delay the inevitable

My personal opinion and from what I have learned about exhaust ports I have to say that small super fast exhaust ports make more power over larger high flowing exhaust ports except in the case of full exhaust systems such as Nextel cup engines. For some reason they like a slightly larger exhaust port but no where even close to what I would call large. Large and Small are ambiguous. In my book anything over about 110% of the valve area is large and anything under 105% of the valve area is very small but the exit velocity seems to play a role here as well. I try to adhere to the 105-108% in our pro Stock engines and it seems that I am not alone in my theory because many of the top notch heads I have seen are about the same or within about 2%. Another very big thing to consider in the tuning of exhaust ports is there sound or should I say the lack of sound. How smooth an exhaust port sounds and how quietly it can move the air are both very serious factors to consider. As the valve opens the sound of the ports should smooth up and get increasingly silent. The loudest portion of the exhaust flow on the bench is from .200 to .400 after that they should go increasingly silent with every lift increment. I have had exhaust ports that actually cracked and popped like fire crackers! With a little seat blending and chamber work I managed to smooth up the flow, gained a measly 2 cfm average and gained 26 horsepower and it still was not correct because the port was to big. The hardest thing I do is try and fix exhaust ports that are screwed up. Its much easier to fix intake ports!
Like an intake port, an exhaust port can be made to flow a great deal of air, Just make it big.

Some rules I live by.

(1) Exit area = 105-110 % of the valve.

(2) Exit air speed at a minimum of 300 and a max of 330 ft/sec mean.

(3) Smooth silent flow by at least .400 lift and absolutely by .500 lift.

I don't agree with two things because neither has ever shown to be beneficial.

(1) 30 degree seats
(2) Designing an exhaust port flow on the flow bench so as to make it flow the most air possible.

All high end engines that I deal with have 55 to 60 degree seats on the exhaust and or intake even Nextel cup heads. As you go steeper in seat angle the flow verses lift curve shows a drop up to about .400 lift and then it really takes off. If you look at it from the point of discharge coefficient you will see that the efficiency curve holds steady then actually increases at the higher lifts. Looking at it this way, would make me think that there would be no actual power loss with step angle seats and low lift cams but no clear advantage either.

http://webpages.charter.net/dsda/curtai ... tangle.xls


A 45 degree or 30 degree seat has yet to prove its worth to me even on engine with gross lifts low as.750. I have not messed with engines with lifts below that mark for over 18 years now so I could be wrong in cases where cam lift is limited to .500 and below or the chamber design is all messed up. I would love to here about anyone's R&D along these lines.

maybe a 30* on the intake? old poncho heads used this for good low lift flow......

The only reason a 30 degree seats works better than a 45 or steeper seat is the fact that the Poncho chambers are machined FLAT making for VERY poor pressure recovery. You cant expect to turn air any more than 15 degrees without consequences. With a flat chamber and a 30 degree seat the air has to turn more than that and a 45 degree seat makes the situation even worse. If the Poncho had a chamber that was designed properly, steeper seats would work.

There are many proponents of the " flow curve must match the camshaft lift curve" theory but I am not one of them. Some people still believe that if the camshaft has a maximum of .700 lift that the area under the flow curve must be maximized in this area as well and anything that happens to the flow curve after .700 lift is of no consequence. Nothing could be more incorrect I assure you! Its like that old theory about 30 degree seats. They flow more down low ( .050 to .350 lift ) so they should make more power for cam profiles at or slightly above .400 lift because they maximize the area under the curve in that area,right? Wrong. You can put a properly designed 55 degree seat and chamber, decrease the flow at .050 to .400 lift and make more power with cams with only .400 lift. You have to design the thing correctly and its tricky. You cant just throw steep angle seats in any head and have this work. You must have convex chambers and good pressure recovery in the chamber or its disastrous. The steeper the seat angles and the larger the throat area, the more important the chamber design becomes.

You turn the air less, use less energy doing so.
You maximize the potential flow in an area more conducive to flow from a piston speed stand point.
You have proper pressure recovery in the chamber ( Equal exit velocity around the entire circumference of the valve head. A controlled deceleration of the air like a venturi divergent angle.)
You get more air fuel mixture in the cylinder.
It makes more power.

That's my theory and I am sticking with it until someone can come up with a better one.LOL
_________________
Darin Morgan
R&D-Cylinder Head Dept.
Reher-Morrison Racing Engines

Calvin Elston:

have been building exhaust headers for professionals for over 15 years. Cup, PS, Stock, SuperStock, Comp, IMSA, SCCA, ARCA etc. I have been busy making venturis in the exhaust systems at the head to header interface.
In more than 60-70% of applications, I can make more power and torque using a tube size with an area as much as 10-12% smaller than the given port area. I would say without hesitation that this is because head porters and engine designers have always been developing their engines and heads based on data from engines that did not have a header or system that was helping them or working effectively. Therefore, without a exhaust header that would help protect the low piston speed power and also reduce system pressure at maximum rpm's; cam duations, intake flow to exhaust flow ratios, maximum lift flow numbers, (when low lift flow is critical on exhaust, which I think is what you folks are discussing in this thread),etc, have all gone astray.
I am often able to allow engine builders to revisit previous conclusions about many aspects of their engines after they have seen what the exhaust system can do for them.

Yes, the port area at least at the header flange was larger than needed. I am not saying the whole port was too large. Bowl volumes are important! the flow bench is what makes exhaust ports bigger in the castings. as the port is increased in area you can flow more air. Usually you can make more horsepower because the headers are too big too and the engine is just pushing everything out, so bigger is always better untill the same pressure that is pushing the exhaust out is also keeping the inlet air from coming in. If you neglect the power of the finite pressure waves in your system and how they are effecting things, the wall comes very soon. Smaller systems usually will carry power further past power peak

If you keep dropping the pressure in the header, and that is not hard, you will eventually get to the condition which you are seeing which is overscavaging. Changing the camshaft around is effectual only to the degree that it was too large or too tight in the beginning. Lets try stepping back further from the situation. There is a basic amount of duration you need to get to a certain rpm and power level with a given engine combo. There is a wall there that you hit when your just dropping header pressure and correspondingly reducing exh duration and overlap, (and all the other tricks there are). It does not allow you to really make too much more power with the same size intake valve. It is a diminishing return. The adjustment is to reduce the size of the exhaust valve. This allows you to keep the duration or cam timing where it really wants to be and still exhaust well enough. The smaller exhaust valve allows most engines to use a larger intake valve and this is the ultimate purpose in using a properly tuned header. The well tuned exhaust side allows a larger intake valve which is the easier way to move more air into the cylinders than sucking from the back door. If your header is not functioning well, you will not be able to get there. Very few engine people get this. I believe if you look at PS type 2valve engine development over the last 10 years, you will see a very large increase in power and rpm, but exh valves are the same or smaller in diameter and the intakes are larger.The merged collector and corresponding header and valve sizing has been a important factor.

It helps to not think in terms of "port" verses "tubing". The tube becomes the port. You do not introduce tight radius for the same reason you do not do it in the ports. The only way to overcome the loss, especially in the transition area and the first 6-8" of the header, is by using larger tubing diameters.
And the problem keeps going down the toilet. "Well, if we made a bigger header, it would make a little more power",(because we overcome the flow loss up front). But the larger tube now needs an even tighter radius to clear the obstruction... And you are now so far away from what the engine wants for a header, it really does not matter what you do from there, the engine won't respond to changes. The smallest diameter tube allows for the largest radius. It can even be smaller in area than the port, especially if it allows a larger radius. (the application can trump this though) The first 6-8" of the header is just as important as the first 3 or 4" in the cylinder head. The same rules and thinking that a head porter uses apply to the header. Most headers I see break all the rules. The last thought is this, the reason the first 10" of the exhaust track is so important as opposed to just the port in the casting? The gas particles, the mass...only gets that far before the ex valve closes behind it, and it no longer is connected to its pressure differential, if you will. At that point in the exhaust track, everything changes. But anything you do to the "particulate-mass" flow in the port while the valve is open that violates flow, will cost you power and you can not get it back. You want to get the "mass-flow" as far away from the ex valve and the cylinder as you can. Larger "anything" in this area does not usually help. I do understand that many applications force the header to break all these rules and larger is the only answer if you need maximum power.

Larry Meaux:

For the past 25 years I have based my cam design on one theory.
What happens before TDC in the intake cam is bad, and minimising the bad helps the engine breathe its maximum.
All the work done by the piston in starting airflow is done in the 1st 75* ATDC, up to the point of maximum piston velocity. From the point of maximum piston velocity on, the piston is progressively slowing down and pulling LESS hard on the intake port with every degree of rotation. Yet because of inertia, the velocity in the intake port is increasing, up to a max at BDC. If everything is done right, the cylinder is still filling when the initake valve shuts after BDC.
By minimizing Reversion before TDC, the piston starts airflow earlier, vs earlier intake valve openings which let in more and higher pressure exhaust gases. The less reversion, the earlier airflow starts after TDC. By having a cam with lots of mid-lift and high-lift area, the valve has more time(duration) to fill the cylinder with harder-flowing air/fuel(inertia ram).
The exhaust cam has its own part to play, I'll cover that later.
This has been my intake theory for 26 years.

UDHarold

i totally agree with your Post.

---------------------------------------i have a lot of Dyno and DragStrip Data
from "before, under old NHRA SS Rules",
where you couldn't backcut valves
-to-
"after new NHRA SS Rules",
when you could backcut valves.

These SuperStock Engines were
Chevy 283's with 1.720/1.500 valves
Chevy 327's with 1.940/1.500 valves
Chevy 350's with 1.940/1.500 valves

Chrysler 318,340,360 with 1.880/1.600 valves
Chrysler 340 with 2.020/1.600 valves
very high velocity Ports, relatively small Port Volumes,
relatively small valve Heads.

100.0 % PerCent of all my Dyno testing so far
with the above Engines have ALL shown
HP and Torque Losses looking at Power Curves
"BELOW" the RPM point of Peak Torque
as Low-Lift Flow CFM is "increased"
by BackCutting or Porting Bowls differently.
These are Solid Roller Cams ranging from
272 deg to 286 deg @ .050" intake durations

around 4500 or so RPM, these Engines were about
dead-even in HP and Torque numbers,
but below 4500 RPM , and as you looked at
Power Curves towards 4000 or 3800 or 3500 RPM's,
you saw Torque Losses with "better Low-Lift Flow"
on these various SS Engines.

So far, 100.0% PerCent of Dyno testing has shown=>

1-Increase Low-Lift Flow = Loss of lower RPM Torque,
below the RPM point of peak Torque occurrence.

2-Increase Low-Lift Flow = sometimes more Hi-RPM HP,
and sometimes the Power Curve was widened a few 100 RPMs,
but never did i see more Torque at the very lower part of
the Torque Curve....increased Low-Lift Flow always
has "hurt" low-end Torque numbers ....with the Cams those Engines run.

Conclusion is, better Low-Lift Flow seems to
sometimes increase Hi-RPM HP,
and decrease or "hurt" very low to lower RPM Torque + HP
with those Engines/Cams.

Even on an off the showroom floor OEM Engine
that might have its RPM point of Peak TQ
ocurring at 3500 RPM...if you go and make
a substantial increase in better Low-Lift Flow,
i'm certain you will definetly see a LOSS in TQ
below 3500 RPMs !

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

Another relationship or another way at looking at Low-Lift Flow
or even some of Mid-Lift Flow=>

i can very easily make just about any
BigBlock Chevy Head w/2.300-2.350/1.880,
have much better Head CFM Flow numbers from .050",
.100",.200",300,.400",.450" Lift increments,
than a 500 cid NHRA ProStock Head.

this is my example point=>
an Edelbrock Victor Race head 2.350 Int valve 45 deg seats,
can be dead-even or better than a ProStock Head
with 2.525" 55+ deg seat angles between .050" to .450" Lift,
but,
the Edelbrock Head will never make 1400 HP on 500 cid N.A.
the Edelbrock Head might make 1000 HP or so at best on 500 CID
with the "same exact or better " Low-Lift Flow Numbers,
...so the Low-Lift and some of Mid-Lift Flow is not
overwhelmingly determining the amount of HP and Torque.

Where you do see a great deal of differences in Flow CFM Numbers
between the Conventional Style-Head -VS- the ProStock Heads
are in the "upper" Mid-Lift -to- High Lift Areas,
where a ProStock Head's Flow Numbers are a ton better,
and 300-400 HP better .

so "equal" Low-Lift Flow does not equal same HP or TQ
in same size CIDs...so Low-Lift Flow is obviously
not the "Controlling Factor".- in the reasons why
there is such a huge HP difference.
---------------------------------------------------------------

Mixture Inertia ram-effect combined with a positive pressure
acoustical wave, at around the time of .400" towards .050" Lift
or so , can be on average of 3 PSI or higher.

So if your Heads Flow Tested at .200" Lift were = 140.0 CFM @28"
...that same 140.0 CFM @.200" Lift may now be capable
of moving approx. 241.0 CFM @28"...about a 100+ CFM gain
from the above effects

so Low-Lift Flow will help you more at the Intake Valve
closing point at higher RPMs, and hurt you more at the
very Lower RPM part of Curve by increasing reverse Flow
or reversion at Intake Closing point,
and filling the cylinder too easily too quickly at the
start of the intake stroke at the lower RPM Curve.

you must need a certain amount of pressure drop
to create the great ram-effect at the intake valve closing point.
if you have too much low-lift flow too soon ,
you place a "kink" or hurt the formation of the depression
curve early in the stroke and hurt ram effect later at IVC

Where your Engine makes its Peak Torque,
is basically where your Engine is breathing its best,
after the RPM point of Peak TQ,
Cylinder filling is less and less.

Below the RPM point of Peak Torque occurrence,
there is not enough Int and Exh System velocities + Length Tuning +
due to CSA's and cylinders being fed too easily too quickly !
you are loosing your ram-effect at the IVC


Too much Low-Lift Flow can cause all kinds of problems
like reversion and overscavenging.
Increasing Low-Lift Flow is sort of like
increasing Cam's Duration at .050" and increasing
the OverLap Period....so that may force you to
use less Duration @.050" + spread Centers.

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

if you have a relatively lower RPM Range Engine
with good to great Low-Lift Flow,
you can or most probably need to shorten up
Cam Duration to offest the great Low-Lift Flow,
sort of like the Duration Numbers the 4 or 5 valve Head
engines have to use.

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

When i purchased my first FlowBench, a SF-110 in late 70's,
i immediately started trying to correlate what
Valve Lifts were important to gain great Flow Numbers at.
Going back and forth from SF-110 to DragStrip,
it started to show using .85% PerCent of the theoretical
Cam's Max Valve Lift.
Example=> .700" Lift Cam times .85% = .595"
or rounded-off to .600" Lift, so i made sure i had
great Flow in a .200" Lift range,
from .400" to .600" Lift Flow on a Bench for a .700" Lift Cam
or
from .600" to .800" Lift Flow if you had a .900" Lift Cam

later on with newer SF-600 Bench + Dyno tests + DragStrip test,
it looked more like .87 % was better to use..not a lot of difference.

By using at least a .37 Lift/Diameter Ratio Camshaft on Intake Side
and using .87% PerCent of that Lift to develope Heads on a FlowBench
has worked pretty well as a starting baseline.
The all-out max-effort 2-valve Head Engines prefer
at least a .39 to .42+ Lift/Diameter Ratios
with great Mid to High Lift Flow Numbers

Also i might add...with current technology
of "Lofting" the Valves...the .87% PerCent Factor i used
may be revised upwards to develope Heads on FlowBench
up or above the Cam's Max Lift point.

---Sorry for long Post

The more I learn the less I know - I should be taking notes.

Is there a way to reduce reversion at lower rpms, and improve scavenging at higher rpms to get the best of boths worlds?

If so - wouldn't that improve average power or "area under the curve"?

I know that's a slightly different question than referencing cfm at lift variations - but when the engine is working - it is at lower rpms you want no reversion and good velocity, and under high rpms you want great scavenging and CFM.

Are they mutually exclusive requirements?

Is there a way to reduce reversion at lower rpms, and improve scavenging at higher rpms to get the best of boths worlds?
you could go towards ProStock technology
by 55 or so degree seat angles,
spreading Lobe Centers,
creating hi-flowing, high velocity, highly efficient ports

as an example of Piston CFM Demand
on a 500 cid Engine turning 9500 rpm
the piston is trying to move this amount of CFM air,
if the Engine cannot get this amount of air at each degree,
it will cause the CFM Demand to grow larger
and "shift" towards BDC

so it shows you need around 579 CFM @28"
to fully satisfy demand, but that can climb
to more or less another 20+ CFM
and shift downwards from Lag times due to Mixture Inertia

Great post!

What are the design differences that result in good low lift low vs good high lift flow? Are you talking mostly about seats and chambers determiniung the difference or some characteristics of the ports too?

for Low-Lift Flow gains,
it will be mostly from your choice of valve seat angle combinations
like if you wanted to use 30 degree seats
-vs-
45 deg seats or 55 deg seats
..that would be your first choice,
then you would work the Chamber to unshroud the Valves
to further increase Low-Lift Flow ,
and use one or multiple valve backcut angles.
and possibly increasing valve's margins to .080 to .120" thick margins

to hurt Low-Lift Flow
go towards 55 deg seats
hurt even more ,
reduce valve's margin thickness,
one or no backcuts
lay back SSR to Apex Peak

The Port doesn't have a major influence on Low-Lift Flow ,
its all in the Valve Job + Chamber + Bowls + a portion of SSR
+ Valve Head Shape

[NewbVetteGuy]

Thanks for the curated comments and tying together most of the nuggets of wisdom in one place, hoffmann900/Bob!

A bit of devil's advocate thinking:
Some of the Calvin Elston comments and previous comments in the thread, definitely make me think that many of the benefits of a steeper valve seat angle are shared by the benefits of a high-velocity Elston-style header / exhaust system (although obviously not MORE CFM @ high lift); like the ability to keep up the low and mid-range torque while being able to go with a slighly bigger duration cam (and the anti-reversion effects), but when you factor in cost and longevity, maybe you're better off sticking with a 45 degree valve seat profile and putting your money in a well-designed, high-velocity style exhaust for most street builds...


On a completely different line of thinking, without more before and after 45 degree and 50 degree air flow chart comparisons between the same set of heads, I have to wonder if the 50 degree profiles, even with pretty aggressive, high lift street hydraulic roller cams are spending enough time in the good air- I get that giving up the low lift flowz is a good thing, but giving up good mid lift flow in exchange for slightly better high lift flow I don't think would be a good trade-off. If you're just starting to get gains back at close to your max lift, I could see that being a problem to power production. The HotRod article used a 1.94" intake valve, I'd expect the CFM gains to require more lift with a 2.02" valve. The EM's motor used a huge 2.065" intake valve, which should gain back some of the low lift flow and curtain area that would've been lost if the stock 2.02" 45 degree valve was just replaced with a 2.02" 50 degree valve, right? (Not very applicable for a typical street SBC; we've got data for Vortec heads with 1.94" valves, we've seen the diff between a 2.02" 45 and a beastly 2.065" 50 degree valve, but precious little else for hard data...

The question for me is really at what lift do you start seeing the gains with a pretty typical 2.02" valve? -For a given valve job does this always occur at a particular Lift to Diameter ratio or is it more head and/or bore dependent?


Somebody's got to have some more before and after 45 vs. 50 degree bench flow #'s that they're sitting on and could share, right? Seems a pretty critically important missing piece to the puzzle...


I'll admit, I'm pretty scared to be the guinnea pig, but I probably won't need my 195cc Profiler heads until end September/Early October if someone wanted to do an A/B flow or dyno test and would be willing to post the results back here, I'd probably pay to ship my heads and cam to you and obviously pay for the new 50 degree valves "for science"... Unless the consensus is that this is a really bad idea and I'll likely regret it, but I'm not getting that vibe from this thread.


Adam

[PRH]

Good reading.

This seems pretty logical to me:

Darin Morgan wrote:
There are many proponents of the " flow curve must match the camshaft lift curve" theory but I am not one of them. Some people still believe that if the camshaft has a maximum of .700 lift that the area under the flow curve must be maximized in this area as well and anything that happens to the flow curve after .700 lift is of no consequence. Nothing could be more incorrect I assure you! Its like that old theory about 30 degree seats. They flow more down low ( .050 to .350 lift ) so they should make more power for cam profiles at or slightly above .400 lift because they maximize the area under the curve in that area,right? Wrong. You can put a properly designed 55 degree seat and chamber, decrease the flow at .050 to .400 lift and make more power with cams with only .400 lift. You have to design the thing correctly and its tricky. You cant just throw steep angle seats in any head and have this work. You must have convex chambers and good pressure recovery in the chamber or its disastrous. The steeper the seat angles and the larger the throat area, the more important the chamber design becomes.

[Stan Weiss]

Good reading.

This seems pretty logical to me:

Darin Morgan wrote:
There are many proponents of the " flow curve must match the camshaft lift curve" theory but I am not one of them. Some people still believe that if the camshaft has a maximum of .700 lift that the area under the flow curve must be maximized in this area as well and anything that happens to the flow curve after .700 lift is of no consequence. Nothing could be more incorrect I assure you! Its like that old theory about 30 degree seats. They flow more down low ( .050 to .350 lift ) so they should make more power for cam profiles at or slightly above .400 lift because they maximize the area under the curve in that area,right? Wrong. You can put a properly designed 55 degree seat and chamber, decrease the flow at .050 to .400 lift and make more power with cams with only .400 lift. You have to design the thing correctly and its tricky. You cant just throw steep angle seats in any head and have this work. You must have convex chambers and good pressure recovery in the chamber or its disastrous. The steeper the seat angles and the larger the throat area, the more important the chamber design becomes.

What some have been asking about is the need to change cam events when changing valve seat angles.

I don't know if Darin still read ST or not, if so hopefully he will step in here.

I see Darin talking about a max lift staying the same I don't see him saying that the other valve events will also remain constant.

Stan

[Warp Speed]

Many have commented on that very question, but if your not Darin, David ect., I guess it doesn't matter.

[RevTheory]

I've been hesitant to post this since about 5:30 AM because it involves you-know-who but I may as well. Plus, it involves TFX pressure data (although not compiled as of the writing) so I find it interesting.

"To start things off, when comparing ‘bad low lift flow heads’ versus ‘good low lift flow heads’ there seems to be a trend of just comparing two different heads on the same long block set up and basing the results off of that. In simple terms a basic A versus B test. This however, is an unfair comparison because either way the engine is built, be it with ‘good low lift flow’ heads or ‘bad lift low flow’ heads, the camshaft should have been optimally spec’d for which ever head combination is being used. So, if you take the cylinder heads off and put the opposing cylinder heads on, the cam is now incorrectly spec’d for this new combination that the cylinder will see, and it cannot produce to its full advantage. As DV was quick to point out an engine does not know how far the valve is lifted: it only recognizes how much airflow exists at the crank angle prevailing at that moment. For instance if our engine has 80 degrees of overlap, and intake flow at TDC is X, then the engine sees X amount of airflow not X amount of lift. In other words the flow could be from either an efficient valve at a lower lift or a less efficient valve at higher lift. The engine does not know nor care which of these two situations exists, it cares only that pressure differentials from the exhaust, across the chamber and to the intake are such so as to optimally scavenge the combustion chamber. Now if we make the flow X more or less at any given point during the overlap the cylinder will either under scavenge or over scavenge because it still only sees air flow for any given crank angle during the overlap period. From this it follows that while the overlap used is application dependant (a lot for a race engine and minimal for a tractor engine) the LCA is dictated by the flow per degree of crank rotation during the overlap period not the valve lift. I am hoping here that my explanation is clear enough to make the point but if I am not quite making it (and Lord knows this is not a simple subject) David’s “How to Build Horsepower” book explains this both clearly and in detail but does take a lot more pages than I have room for here to do so.

During both his seminar and during my interview with DV one misconceived aspect of the induction cycle was repeatedly stressed. This misconception in many cases proves the hinge point on which the ‘too much flow at low lift is bad’ and the ‘high lift flow matters most because that is when the piston demand is highest theory is based. As DV pointed out and as was demonstrated by the TFX port and in-cylinder pressure demo at DV’s seminar we saw that for the whole induction event , the biggest draw on the intake port was produced by the exhaust pulse scavenging the combustion chamber not on piston travelling down the bore at peak speed. For example, on the demo 525 inch big block Chevy my partner and I built, the TFX pressure instrumentation showed, at TDC, a differential of 11 PSI from the intake runner through the combustion chamber to the exhaust! This broke down to about a 3.5 psi pressure pulse in the intake and a near astonishing 7.5 psi suction from the exhaust. The depression caused by the piston traveling down the bore was a mere 1.5 psi!

Back on the subject of piston induced depression. Here David pointed that if the piston depression is more than 20 inches of water, which is around - .7 PSI, then the head(s) need to flow more air for the cubes/rpm combination being used. So if one were to run a head with ‘bad low lift flow’ the opportunity of taking full advantage of one of the most critical aspects of induction and exhaust tuning would be substantially lost. If the maximum advantage of pressure pulse and inertial tuning is actively sought then it will be seen that low lift flow is vital. As DV so aptly demonstrated air is considerably heavier than 99.99% of the performance community thinks it is. From this it also stems that port velocity is also more important than 99.99% of the performance community think it is.

Also playing a role here is the fact that any lift less than peak valve lift happens twice in a cycle where as the flow at max valve lift happens only once. So now we see that the start to our induction cycle and the end ramming effect for more trapped mass is dependent on a cylinder head’s ‘low lift flow’. The late Smokey Yunick reinforced the fact that lower valve lift flows are to be focused on because they are used twice in a cycle.

Another misleading aspect(and one that is touted by some very respected University book reading) is that the intake valve closure point is the most important and influential valve event in the event cycle. Using the pressure traces from the TFX gear and our dyno tests DV quickly shot that assumption to ribbons. Using the pressure data he demonstrated that it is vital, for an optimal induction cycle, to get it right in the first (opening) half. He also showed from this that if the first part of the induction cycle was not optimal there was no possible way to make a full redemption in the second half. This again was a pointer to the fact that strong low lift flow is a prime requirement if the best volumetric efficiencies are to be seen. Although relegated from the number one spot the intake valve closure point is still important if the most of inertial ramming of the cylinder is to be made. However DV was adamant that no matter when the intake was closed it could never compensate for incorrect event timing during the overlap period. For what it is worth DV also pointed out that the 5 hp per cube outputs of current F1 engines are largely due to low lift flow per cube figures as much as three times that of a current cup car engine.

Another issue though more minor in nature seems also to be entwined with the low lift flow arguments is that a cylinder head with really good ‘low lift flow’ sacrifices ‘high lift flow’. Due to our knowledge of ports and valve seats gleaned from any one of a number of big Superflow flow benches at the university even my own testing has found that to be not entirely true. Achieving above average low lift flow may hurt a fraction of high lift flow but not necessarily kill it to the point of offsetting any possible gains from the better low lift flow. As we discussed in the above, the advantages of flowing better at lower valve lifts far outweigh the odd CFM or so that might be lost at high lift.

Here David outlined one of his research projects done many years ago while working on the 2 liter Pinto engine. With the induction (twin side draft 45 DCOE Webbers) and exhaust dimensions held constant he used two different heads on the test engines long block. The first of these was one of DV’s own super ported two valve factory iron heads which we will call head A. The intake flow for this head was a little over 150 cfm at 0.250 lift and 218 at 0.650 lift.
Head B was a near stock Cosworth 4 valve head which delivered almost 190 cfm at 0.250 lift and 214 cfm at 0.500 lift After testing the two valve head the 4 valve head was cammed such that it delivered the same torque at 3000 rpm. The tests showed the greater ‘low lift flow head’ (B) even with slightly less peak flow achieved about 30 HP gain over Head A. In case you might think this an isolated instance DV also has some extensive tests with domestic small block V8’s. but it’s not my place here to go publishing all the results of tests done with the intent to use it in articles and books he is going to write.

Tests done with the TFX gear during DV’s seminar at UNOH and the dyno results put forward along with DV’s explanations has left no doubt in my mind that good low lift flow is the performance engine builders friend not, as seems so often to be believed his enemy. I would also like to point out that the arguments for such are based on hard core testing not personal opinions. As such they are not intentional attacks on those with contrary views but in the light of current test data it might be a good idea to re-think any possible counter arguments. Which, if you feel you want to discuss anything I have said here brings me to my final point. At this moment in time Goodson and Myron Cottrell of TPI Specialties are negotiating with DV to do a streamlined two day version of the 3 day ‘How to Build Horsepower’ UNOH seminar he did back in March. Details will be posted on Davidvizardseminars.com."

It's been years since that was posted so maybe it's time for a review as things always change.

[user-30257]

Many have commented on that very question, but if your not Darin, David ect., I guess it doesn't matter.

Bingo.

[Stan Weiss]

Many have commented on that very question, but if your not Darin, David ect., I guess it doesn't matter.

Jay,
Someone posted a quote from Darin I replied.

When I get time I will have to reread what has been posted. Or it could be that some things are just too obtuse for me to understand.

Stan

[hoffman900]

Rev,

Don't be nervous, take out the silliness and I think everyone is here to learn.

David's TFX data comes up every so often, and why I have no doubt it is accurate, I don't believe it shows what he thinks it does.
Taken from another thread...

Pressure graph 2.png

Pressure graph 1.png

Only one thing worse than no data...............bad data.

Honest question: why is it bad data? Isn't he just showing what the sensors showed?

Perhaps Clint can chime in as I believe it was his equipment.

I want to make it perfectly clear that the graphs posted on the previous page are not graphs generated by our equipment.

I could be mistaken but I don't think David has ever published any intake/exhaust pressure graphs generated by our TFX equipment, as far as I recall he has only used our TFX combustion pressure traces.

In any event the traces posted in this thread so far, are not ours.

I want to make it perfectly clear that the graphs posted on the previous page are not graphs generated by our equipment.

I could be mistaken but I don't think David has ever published any intake/exhaust pressure graphs generated by our TFX equipment, as far as I recall he has only used our TFX combustion pressure traces.

In any event the traces posted in this thread so far, are not ours.

I'm guessing but I'd say that the "Cylinder Pressure" listed in the DV traces from the previous page may be some sensor arrangement related to pressure across the valve or who knows what the arrangement was. One thing is clear though the "Cylinder Pressure" trace is clearly NOT cylinder pressure relative to atmospheric pressure.

My apologies, Clint. I'm not sure where I got that (I may have imagined it ), but apparently, I was wrong.

Now, what needs to be done regarding DV's tests to make them useful? Some conversion factor, perhaps?

No problem, maybe it was David's black background graph. All of our TFX graphs use black background too, as opposed to the more usual white background.

I wouldn't say that David's graphs are bad data, but a bit more clarification as to the sensor locations and what the trace actually is would help.


Usually the greatest depression (relative to atmospheric pressure) will be part way down the intake stroke and it can be several psi at high rpm. Flowbench tests are typically in the vicinity of 1 psi, which is not nearly enough.

It is possible with certain builds and under certain rpm conditions and intake/exhaust combinations to have the greatest depression in the cylinder occur at overlap.


At TDC the pressure in the cylinder (ignoring all else) does not tell the flow story very well. For example, lets say at some rpm the exhaust port pressure wave was -10 psi and the intake port pressure wave was +5 psi (due to pressure reflections generated from the previous cycle). At TDC the exposed intake area is greater than the exposed exhaust area. The cylinder pressure is likely to be somewhere between +5 and -10 psi, and considering the previous movement of the piston, and the ratio of exposed intake/exhaust areas etc. etc. the cylinder pressure at TDC will tend to be more than midway between +5 and -10 (i.e. -2.5 psi). For arguments sake lets say the cylinder pressure at TDC is 0 (i.e. atmospheric pressure). One could think that since the pressure in the cylinder was atmospheric at TDC then there would be no flow at TDC, but that would be very wrong. The cylinder is just a pocket between +5 psi and -10 psi, a wave pressure differential of 15 psi. The flow through the cylinder at TDC would actually be fairly substantial, despite the cylinder pressure being atmospheric pressure at TDC.

The piston creates almost twice the negative pressure he is showing at overlap.
Overlap can create a pretty big pull but..............

These posts were from four years ago (to be fair, your quote is from 2011, 7 years ago). I had hoped David corrected this information or at least got some insight from Clint (and others) himself so doesn't continue to lead down the wrong path.