tumble and swirl, quench & squish


Staff member
I could fill chapters on these subjects, but its basically not something the average engine builder assembling off the shelf parts needs to be overly concerned with other than to understand the potential power may change due to the parts selected and the careful assembly and measurement during the engine assembly will effect his results. I think you can assume having minimal quench clearance to maximize the effect, is FAR less important on a low compression engine than on a race engine by the simple fact that most factory produced production engines have a quench or squish clearance thats so large that its most likely NOT producing its full potential advantage to the combustion process
that a tighter squish/quench clearance would produce, most likely because its far easier to build and assemble engines and have fewer interference or clearance and binding issues to do so!
your engines tendency to get into detonation is the result of higher effective compression that the fuel octane your using can permit, but your piston is only compressing the fuel air mix AFTER both valves seat.detonation is a condition that occurs as the result of both cylinder temperature compressions effect on your fuel/air ratio and several other factors that cause fuel to self ignite under some conditions, lowering the EFFECTIVE compression , lowering the heat in the combustion chamber , or designing the engines quench,squish,turbulence and other factors to greatly reduce the engines tendency to get into detonation will obviously help prevent it.
QUENCH DISTANCE BETWEEN the heads and pistons at TDC. to fall between,
as a general rule
shoot for a quench of
Minimum .045", max .060", more important is the squish area and chamber shape.
RANGE , DON'T GO INSANE IF it's ONLY POSSIBLE WITH YOUR COMPONENT CHOICES TO GET DOWN TO .050, QUENCH, its marginally less effective but still works)
on most engines,the idea is to force a compressed jet of that fuel air mix, trapped between the almost contacting surfaces,(PISTON ALMOST HITTING THE HEAD) into the combustion chamber at high speed this increased turbulence increases the burn speed in the combustion chamber, and the two surfaces so close together cool the trapped gases between them ,just enough to slow or prevent detonation, reducing the distance the flame front must travel ACROSS the cylinder allows less ignition advance to be used so, mechanical EFFICIENCY to be improved
BTW When a head gasket thickness is listed its supposed to be the compressed gasket thickness ,this can be important in determining quench distance.

It might help to keep in mind, the amount of TIME the piston, is in close proximity to the cylinder head, remember the 4 cycle ,
or 720 degrees the crank turns, the the very brief time the piston spends on the power stroke, there is only about 25-40 degrees of that 720 degree cycle here theres near peak pressure above the piston, and even at idle speeds (lets assume 800 rpm) the power stoke or peak pressure , is only lets say 35 degrees in a 720 degree cycle thats happening repeatedly 13.3 times per second.

now at 7000 rpm, thats happening 58.3 times a second, think it through, about 5% of the whole 720 degree repetitive cycle
has significant torque producing pressure above the piston, exerting thrust against the crank shaft journal,
the other 95% is required just to prepare the cylinder to fire and exert that pressure pulse, theres a power pulse at each 90 degree turn of the cranks rotation, on every other time it reaches TDC.
there's about a 600-700 PSI peak pressure above the piston, that rapidly drops off as the piston descends on the power stroke as the fuel burns, expending its energy, by the time the piston reaches mid stroke pressure is reduced to far less than it was during the first 30 degrees of rotation.
if you have a 4.25" bore 454 BBC theres about 14 sq inches of piston surface are, thus about 8500 lbs or more

(depending on the F/A ratio and compression )
of force very briefly above the piston, every 90 degrees of rotation.
that pressure drops of rapidly as the piston descends down the bore and after the exhaust valve opens it drops below atmospheric pressure briefly



If you take the time required to understand your options and having built an engine with close to ideal quench youll find its slightly less likely to get into detonation conditions that can quickly destroy a piston, resulting in massive engine failure.



Quench and Squish area explained.jpg








the first step is having the block decks accurately measured and if necessary machined so the block decks level and parallel with the crank center line.
in this example directly below the plugs GROUND electrode is ALMOST EXACTLY OPPOSITE, OF THE IDEAL INDEXED location , as its located wher the piston has the best chance of contacting the ground electrode,


example , now ideally youll want the ground electrode facing the roof of the combustion chamber for max clearance and flame propagation
similar too...

read the links







if the blocks 100% correctly machined, and you get a wide variation in piston to deck heights , remember theres some variation in both piston pin height and connecting rod length, if for example you find SOME PISTONS .003 AND SOME AT 009, that MIGHT BE A RATHER HIGH VARIATION
It would indicate the block deck needs to be machined or perhaps the pistons and rods could be interchanged (longer rods, re matched with shorter piston pin heights) to get a lower average variation
you should be trying to get a .038-.044 quench height, personally I try to be on the .042-.044 range.
if the pistons sticking above the block deck your calculation subtracts the distance from the compressed head gasket thickness, if its below deck you add that distance to the head gasket thickness to arrive at the quench distance.


KEEP in mind your engines effective or dynamic compression and your effective quench distance, the materials in the cylinder head, coolant and oil temps., ,ignition advance curve,plus the engines combustion chamber temperature and fuels octane level plus several other factors determine the engines detonation threshold or tolerance, each component choice effects the complete combo[/color]

"Determining seat timing: Since the early days of the internal combustion gasoline engine, engineers have known that the Otto four stroke engine is compression limited and that the quality of the fuel used determines the CR at which the engine could operate. However, it is not the Static CR but the actual running CR of the engine that is important. Compression of the air/fuel mixture cannot start while the intake valve is open. It may start slightly before the intake valve is fully seated. However, there is no easy way to determine this point so using the advertised duration number provided by the cam manufacture is the next best thing. Most cam grinders use .006" of tappet lift (hydraulic cam), although some use other values, with .004" being a common one. This duration is often referred to as the "seat timing". We will used advertised duration for calculating the DCR. "

quench is basically achieved by momentarily allowing the rotation of the piston past TDC so its placing the piston and heads flat matching surfaces as the piston rotates past tdc, so close too each other that and compressed gasses and air are forced out from between the surfaces in a near collision,between the surfaces , as the rapid exiting of the compressed fuel air mix has both a cooling effect and the turbulence increases the combustion, plus heat absorbing factors and lack of space between the surfaces prevents the fuel/air mix from burning, it also reduces the distance the flame front needs to travel once ignition starts, because the SQUISH has forced much of the previously trapped fuel/air mix into the combustion chamber and away from the cylinder wall, and the increased turbulence that results, speeds the burn in that smaller combustion chamber.

squish is the rapid forced displacement of the fuel/air mix from the quench area into the smaller main combustion chamber which rapidly increases turbulence and the burn speed ,by increasing the flame front speed,much as throwing a cup of gas into a camp fire would







Any decent, experienced machine shop can measure your cylinder heads combustion chamber,
and calculate the required clearances after measuring your heads combustion chamber, and then do the correct machine work on your piston domes,
machining the domes for adequate,spark plug,clearance
this is a very common issue and easily resolved,


and it seldom costs much to have done.











http://www.chevyhiperformance.com/tech/ ... nce_guide/

remember at peak hp your spinning 5000rpm-7000rpm
that means that at 6000rpm, theres a power stroke 50 times a second, so theres about .004 seconds 4 thousandths of a second between the time the piston reaches 30 degrees before tdc and swings thru to 30 degrees past tdc
you'll generally want to built any performance engine to the maximum compression levels the fuels octane and burn characteristics will allow you to and to maximize the port flow velocity in the engine to reach about 300 fps in in port flow rates, theres links to read at the bottom of the post with more info, and look over the pictures, but remember the proper cam timing, port flow rates, and your engines header design/exhaust scavenging will also effect the engines power and those factors tend to be even more important in the 4500rpm-7500rpm range in most performance engine build ups

as most guys know your engine needs less ignition advance as the engine idles than when it starts to increase in rpm levels,because although the cylinders compressed fuel/air mix ignites and burns the time it takes for the cylinders mix to burn and produce useful pressure over the piston varies with several factors and the engine rpms are a major factor in that burn speed. ignition timing tends to go from about 8-12 degrees before top dead center to about 34-39 degrees advanced by about 3100rpm-3300, but above,about 3100-3300rpm, the rapid piston movement, and increase in several other factors like the results of squish & quench tend to increase the combustion burn speed to the point further ignition advance becomes counter productive. ignition in the cylinders may start 36-40degrees before TDC but the max pressure builds in the 24-30 degrees after TDC
the quench area basically form a jet of burning f/a mix being shot into the chamber speeding the burn,quench and SQUISH are both different concepts, but when matched correctly speed the burn and reduce the tendency for detonation.

"Several prominent engine builders disagree with the need for swirl and/or tumble at engine speeds above 3-4k rpm.
mostly because the rapid flow of fuel/air mix and the rapid piston movement and heat in a running engine and the limited time available for fuel vaporization before ignition tends to lower the importance or partly reduce its effectiveness in homogenizing the mixture ratio throughout the combustion chambers, as the rapid squish and quench tend to speed up the mix and burn characteristics in the engine as the rpms increase

They (Reher-Morrison, etc.)indicate that at high rpms the fuel/air mixture becomes highly turbulent due to the squish areas of the chamber and heads with zero swirl/tumble can/do outperform heads with substantial swirl/tumble.
this has proven to be true in SOME engines while in a few the results were inconclusive at best but what seems to be consistent is that the higher the compression and the higher the engines speed the less important tumble & swirl become to the resulting PEAK POWER the engine may produce, but in almost all cases the introduction of tumble & swirl at low and mid rpms (under 3500rpm) tended to help lower emissions, and increase power and mileage)
valve seat and back face angles ,valve diameter and valve lift and duration effect the flow thru the curtain area

keep in mind that valve may be forced off its seat, too full lift and re-seating 50 plus TIMES A SECOND at near 5500 rpm, so theres very little TIME for gases to move through the very restrictive space between the valve seat and valve edge

Calculating the valve curtain area
The following equation mathematically defines the available flow area for any given valve diameter and lift value:
Area = valve diameter x 0.98 x 3.14 x valve lift
Where 3.14 = pi (π)
For a typical 2.02-inch intake valve at .500-inch lift, it calculates as follows:
Area = 2.02 x 0.98 x 3.14 x 0.500 = 3.107 square inches



They do, however, agree that swirl (on 2 valve engines) and tumble (on 4 valve engines) is very important at engine speeds in the 1.5-3k rpm range when piston velocity is slower.

Also, engines with good combustion chamber designs/dome shape can tolerate A/F ratios in the 13.5:1 range (BSFC numbers in the 0.36 range) without detonation concerns.

Larry Widmer (Energy Dynamics, aka Endyn) was one of the first engine builders to explore the concept of swirl/tumble in the early/mid 80s. Check his website for more information.

The "craze" lasted for almost two decades. It wasn't until recently when several serious engine builders conducted detailed studies and back to back dyno test were the theories of swirl/tumble in Hi-Perf. engines proven to be wrong."











these two graphs from



tricks like back cuts on valves increase low rpm and low lift flow, and factors like header primary length ,collector design for increased scavenging and matching intake port cross sectional area to the intended flow rate, and cam timing are critical

the chart pictured below is important but a bit mis-leading,
because it shows only about a 15% gain in power if you went from a 8:1 compression ratio to a 14:1 compression ratio,
in reality , you could more likely see an 15%-18% gain,
with some changes in cam timing and exhaust tuning alone,
and with other changes, 20% gains would be reasonable with that increase in compression.
so to put that in perspective, if you had a 9.5:1 compression 454 chevy,
that made lets assume 425hp, and swapped a cam and pistons,
and increased it to 12.5:1 compression .
it would be rather reasonable to expect, a gain in the range,
close to 10%, or a boost to near 470 hp once the correct octane fuel was used.




related info

btw what your trying to accomplish is maintaining the highest pressure after tdc that you can and the lowest pressure btdc without getting into detonation with the available octane fuel
thus maximizing the pressure curve where you can effectively use that pressure




http://www.chevyhiperformance.com/tech/ ... index.html


http://www.youtube.com/watch?v=W3qD6pYT ... re=related



read these and sub links







Volumetric Efficiency: Is calculated by dividing the mass of air inducted into the cylinder between IVO and IVC divided by the mass of air that would fill the cylinder at atmospheric pressure (with the piston at BDC). Typical values range from 0.6 to 1.2, or 60% to 120%. Peak torque always occurs at the engine speed that produced the highest volumetric efficiency.
keep in mind as rpms increase so do port speeds and volumetric efficiency UP TO A POINT, WHERE THE TIME LIMITATION TO FILL AND SCAVENGE the cylinder limits power








http://books.google.com/books?id=j8pgKX ... rs&f=false

http://www.scribd.com/doc/21485309/Comb ... ber-Design



http://www.eng-tips.com/viewthread.cfm? ... 46&page=47




http://www.circletrack.com/enginetech/c ... _new_fuel/

read thru these, threads and sub links,
keep in mind theres no exact compression ratio point where detonation becomes a sure problem, while there is a compression ratio range or cylinder PSI at a known altitude, at a known air and coolant temperature,where a known octane fuel at a certain known temperature in a known combustion chamber design with a known quench and with a set surface texture or roughness , and of a known diam. and with a known spark plug design and heat range, a set distance from the cylinder center-line where its far more likely to occur, but factors as minor as the valve edge and seat diam. and ignition timing curve,the engines rpm, the effective rate of cylinder exhaust scavenging and the fuel/ air ratio and a dozen or more factors can and do effect the results.
you can easily find that two similar engines will perform quite differently due to minor differences in the engine prep, things like a few spark plug threads exposed in the combustion chamber or a sharp edge on an exhaust valve can easily swing the results noticeably.



theres dozens of factors that will effect your engines tendency to get into detonation.



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heres a bit more useful info


"Quench and squish are not the same thing, and are not produced to the same degree by the same conditions.
Squish: gasses trapped between the piston dome and head are ejected across the chamber at high speed by the near-collision of the piston dome and head, causing turbulence and mixture homogenization. Squish occurs whenever 2 parallel surfaces approach each other closely at or near TDC. Too close = pumping loss. Too far: low squish velocity, less turbulence. The speed of the approach is a function of stroke length (which controls the absolute distance; longer stroke = higher F/S @ constant RPM) and rod ratio (which controls the relative speed change; low ratio = faster).
Quench: lowers the temperature of end gasses trapped between the piston dome and head by radiation and conduction to prevent a second flame front from igniting mix prematurely due to thermal shock, etc.
For motors with 3.5-4.5  bore, a quench distance of .040-.045 appears to work well (measured dry, cold and static with steel rods). This will result in almost .000 clearance (hot, wet and running) due to thermal expansion, rod stretch, piston rock-over etc.
Why are certain motors â safe  with higher quench distances?
The original intent was to have only relatively cool gas present in the quench band during the at risk € period, which begins (after ignition) @ TDC and beyond to 14° (the location of peak pressure), since the entire mix is not burning until at least 20° (?) ATDC. A motor with 4  stroke and 6.2  rods (n = 1.55) has moved about 4% of its stroke @ 20° ATDC, or .160  When the motor is running, the quench area is as wide as .160 inch during the critical flame propagation period, in which secondary ignition will cause knocking. This means that in a motor built to .040” clearance (cold), almost .000 (hot & running) STILL has .160 inch or more quench distance during its gas pressure rise period, but the trapped gas burns very slowly (if at all), and the flame does not spread to the main chamber to cause knocking.
Doesn't this mean that the stroke and rod ratio affect the width of the quench area during the high pressure period immediately after TDC up to 14° ATDC? A motor with a very long stroke and very short rod will have big movement (5” stroke, 1.5 ratio = .099 ) by 14° ATDC; how are these motors safe? Is there a point where the large piston movement makes the quench area so big during the high pressure period that quench doesn't work?

Can a motor with very short stroke and long rods (where the quench area is still very small @ 14° ATDC) get away with looser quench?"

before you reach for your wallet, do some basic math and read a few dozen related links



USE THE CALCULATORS to match port size to intended rpm levels... but keep in mind valve lift and port flow limitations
http://www.circletrack.com/enginetech/1 ... ch_engine/






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Magnificent Quench

What is the most, exact precisely defined occurrence in all piston engines? It isn’t ignition timing, combustion, crank indexing, or valve events. It is Top Dead Center. You can’t build an engine with an error at Top Dead Center because TDC is what everything else is measured from. Spark scatter, crank flex and cam timing can move, but TDC is when the piston is closest to the cylinder head in any one cylinder. The combustion process gets serious at Top Dead Center and about 12 degrees after TDC, most engines want to have maximum cylinder pressure. If maximum cylinder pressure occurs 10 degrees earlier or later, power goes away. Normal ignition timing is adjusted to achieve max cylinder pressure at 12 degrees after TDC. If your timing was set at 36 degrees before TDC that is a 48 degree head start on our 12 degree ATDC target. A lot of things can happen in 48 degrees and since different cylinders burn at different rates and don’t even burn at the same rate cycle to cycle, each cylinder would likely benefit from custom timing for each cylinder and each cycle. Special tailored timing is possible but there is an easier way—“Magnificent Quench”. Take a coffee can ½ full of gasoline burning with slow flicking flame. Strike the can with a baseball bat and you have what I would call a “fast burn”, much like what we want in the combustion chamber. The fast burn idea helps our performance engine by shortening the overall burn time and the amount of spark lead (negative torque) dialed in with the distributor. If you go from 36 degrees total to 32 degrees total and power increases, you either shortened the burn time or just had too much timing dialed-in in the first place. If you have really shortened the burn time, you won’t need so much burning going on before Top Dead Center. Now you can retard timing and increase HP. Did you ever have an engine that didn’t seem to care what timing it had? This is not the usual case with a fast burn combustion but an old style engine with big differences in optimum timing cylinder to cylinder will need 40 degrees of timing on some and others only need 26 degrees. If you set the distributor at 34 degrees, it is likely that 4 cylinders will want more timing and 4 cylinders will want less ( V-8). Moving the timing just changes, which cylinders are doing most of the work. Go too far and some cylinders may take a vacation. Now what does quench really do? First, it kicks the burning flame front across and around the cylinder at exactly TDC in all cylinders. Even with spark scatter, the big fire happens as the tight quench blasts the 32 degree old flame around the chamber. Just as with the coffee can, big flame or small flame, hit it with a baseball bat and they are all big instantly. The need for custom cylinder-to-cylinder timing gets minimized with a good quench. The more air activity in a cylinder you have the less ignition timing you are likely to need. When you add extra head gaskets to lower compression you usually lose enough quench that it is like striking the burning coffee can with a pencil. No fire ball here and that .070-.090 quench distance acts like a shock absorber for flame travel by slowing down any naturally occurring chamber activity. A slow burn means you need more timing and you will have more burn variation cycle-to-cycle and cylinder-to-cylinder, result more ping. Our step and step dish pistons are designed not only to maximize quench but to allow the flame to travel to the opposite side of the cylinder at TDC. The further the flame is driven, the faster the burn rate and the less timing is required. The step design also reduces the piston surface area and helps the piston top stay below 600 degree f (necessary to keep out of detonation). All of our forged pistons that are lower compression than a flat-top are step or step dish design. A nice thing about the step design is that it allows us to make a lighter piston. Our hypereutectic AMC, Buick, Chrysler, Ford, Oldsmobile and Pontiac all offer step designs. We cannot design a 302 Chevy step dish piston at 12:1 compression ratio but a lot of engines can use it to generate good pump gas compression ratio. Supercharging with a quench has always been difficult. A step dish is generally friendly to supercharging because you can have increased dish volume while maintaining a quench and cool top land temperatures. You may want to read our new design article for more information. ".

By John Erb
Chief Engineer
KB Performance Pistons
The step design also reduces the piston surface area and helps the piston top stay below 600 degree f (necessary to keep out of detonation). All of our forged pistons that are lower compression than a flat-top are step or step dish design. A nice thing about the step design is that it allows us to make a lighter piston.

I don't see how a stepped or dished piston could have less surface area than a flat top piston. Also how would that make a lighter piston. These two go hand-in-hand, therefore whats true for one will be true for the other statement.
after having read thru that a few times I see where it might be confusing,
I think hes just assuming you understand what hes talking about in that the flat quench area if correctly clearance has about a .038-.042 thousandths, on a small block, , with a BBC the parts are heavier so you need to allow a bit more stretch at upper rpms so your quench or space between the cylinder heads flat matching quench area and the stepped area, that effectively slams into very close proximity to the head forcing any fuel/air mix to he squished out of the area between the flat surfaces, needs to be marginally larger, so theres effectively far less burn or heat in those areas,
on a BBC you can get away with up to about .060 because the burn is effectively (QUENCHED), so a much higher percentage of the actual burn process happens in the smaller area thats recessed in the piston, and combustion chamber, or put a different way, the quench areas not being heated to nearly the same extent as the combustion chamber and matched recessed area in the piston, and the rapidly (SQUISHED) volume that was formally over that area was forced in a jet of mixed burning material into that recessed area greatly speeding the burn time, that may effectively remove about 20%-to- 30% of the piston surface from exposure to the full heat of a combustion process as the quench/squish has effectively forced 20%-to-30%away from its surface area.
if youve got 20% or more of a piston not being exposed to the same heat levels and knowing aluminum transfers heat rapidly that would tend to reduce the heat over the whole surface some what, much like if you tried to boil water in a 12" diam. pot with it sitting so only 2/3rds were over the burner on a stove, so only about 8" of the base was over the burner, vs centering the pot on the burner.
a piston subjected to less heat should be able to be designed with a thinner and lighter upper structure as it doesn,t need the mass to act as a heat sink, and can rely to a greater extent on oil splash to carry heat from its underside. because the pin boss structure in the center all ready has the required structural mass and strength that can be machined away with no loss in structural rigidity, removing weight that is not required but is present in a flat top design
you can add a bit of chamber volume and reduce the potential hot spots that help cause detonation by opening and blending and smoothing the combustion chamber




detonation and rapid heat increases can ruin piston ring seal to the bore wall

theres a great deal more to the requirements to prevent detonation from causing problems than simply calculating the static compression.
combustion chamber shape,
quench, distance,
engine operating temps,
fuel octane levels,
oil getting past the rings and valve seals,
coolant temps,
ignition advance curve,
spark plug heat range,
connecting rod length,
piston surface shape,
inlet air temps ,
exhaust scavenging efficiency,
cam timing,
your fuel air ratio,
and several more factors will effect the results
In DAVID VISARDS book How to Port and Flow Test Cylinder Heads,
he says that he tries for a combustion chamber and piston dome combo that will allow 200 cranking psi
and a quench in the .038-.042 range, on a sbc, when he runs on 93 octane.
but for every octane number step lower, he suggests the cranking psi should be lowered by about 5 psi.
IE on 91 octane fuel being 2 points lower you reduce cranking pressure from 200 psi to 190 psi,
89 octane being two points lower yet you reduce cranking psi 10 psi lower yet, etc.











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By John Erb - Chief Engineer - KB Performance Pistons
The step design also reduces the piston surface area and helps the piston top stay below 600 degree f (necessary to keep out of detonation).
Ok I see what your saying about John Erb's statements.

So if your interpretation of what he is saying is right......then we have more heat (BTU's) focused over a small area. If we both agree that's a true statement, then John Erb's statement quoted above would NOT be true. If you take the same amount of heat over smaller area the temperature has to be higher.....NOT less.

To argue the other side. Time is also a factor for the surface temp of the piston. Less time for the transfer of heat means lower piston temp in the reduced area and more expansion of the gases. Time would be less with more turbulence and the mixing of fuel and air.

So it comes down to which is the bigger factor "Surface Area" or "Time".

Comments welcome from anyone reading, I encourage different opinions as I'm sure Grumpy does to !!!


  • DishPiston.gif
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just a thought, with a fast burn,combustion chamber and squish and quench making the burn more Effecient I see less piston area being subjected to the heat over a shorter period of time, and since the fuel/air mix holds the same potential BTUs I don,t see how the faster burn cycle will NOT produce slightly less heat transfer to the surface of the piston

read this link




http://www.kb-silvolite.com/article.php ... ad&A_id=39


look at this rather typical mild performance,cam timing card and its Timing:
TAPPET @ .004
Lift: Opens Closes ADV Duration
Intake 34 BTDC 80 ABDC 294 °
Exhaust 87 BBDC 37 ATDC 304 °
notice the intake valve opened 34 degrees BEFORE the piston reaches TOP DEAD CENTER, and doesn,t close until the pistons already moved, all the way to bottom dead center and back over an inch up the cylinder on the following compression stroke, as the piston drops away from top dead center the pressure in the cylinders rapidly reduced resulting in intake port flow to fill the cylinder but because the valves are frequently located off the cylinder center line and because of the rapid movement of the piston away from the combustion chamber the air flow swirls and tumbles to some degree inducing un-equally fuel air ratio distribution, as the piston sweeps upward on the compression stroke the compressing cylinder volume is compacted into the combustion chamber and a squeezing effect shoots the trapped volume trapped in the quench area into the combustion chamber, like a rubber mallet smacking a raw egg on a counter top,resulting in rapid mixing and burning of the trapped and compressed mix, of fuel/air mist.
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grumpyvette said:
just a thought, with a fast burn,combustion chamber and squish and quench making the burn
more Efficient I see less piston area being subjected to the heat over a shorter period of time, and since the fuel/air mix
holds the same potential BTUs I don,t see how the faster burn cycle will NOT produce slightly less heat transfer to the
surface of the piston

If there is 20% LESS AREA being subjected to the SAME HEAT (BTU) of combustion, but it takes 20% LESS TIME, then the surface
temperature of the piston is the SAME as...... when all 100% of the AREA is subjected to SAME HEAT over 100% of the TIME.

I think we are pretty much saying the same thing. It just comes down to how fast is a fast burn cycle.
its best done at a machine shop for consistency, its not designed to be very deep. its designed to tapper gradually getting deeper from cylinder wall to combustion chamber and enhance the (JET of squish and compressed FUEL/AIR MIX EFFECT, to speed combustion)
read thru the links

http://www.kb-silvolite.com/article.php ... ad&A_id=39

http://pesn.com/2005/10/13/9600187_Desi ... _Chambers/

The intriguing thing is Somender Grooves. Usually cut across the quench area and aimed at the spark plug. The theory is that in that last little bit of compression stroke, the gases in the quench are are ejected at high velocity toward the plug adding turbulence... Some guys point them towards the exhaust valves... But whatever direction they are pointed, I'm sure they add either swirl or turbulence.

In all cases I've read about, they are able to bump the timing a bit more w/o detonation

I realize this is all voodoo, but he does have a patent on this and some pretty extensive testing ...

Here's the link to his patent:
Somender Grooves2b.jpg



anyone used this yet?
A groove cut into the head to channel the quenched compressed fuel/air mix rapidly towards the spark plug.






the quench pads the area between the flat surface of the head and the matching flat surface on the piston
BOTH the piston and cylinder head must have MATCHED quench areas for quench to work

heres a cylinder head with a groove cut in the flat quench pad area to speed the quench gasses -trapped f/a mix jet forced into the cylinder to speed the cylinder burn time
read link

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"Interesting concept, Grumpy!

During combustion, the cylinder volume is very small. Heat losses into the piston head and cylinder head are unavoidable. In order to reduce the heat losses, the burn time needs to be as quick as possible. This can be achieved by high flame velocities, which are traditionally accomplished by increasing the laminar burning velocity, or by turbulence intensity, or both.

The highest laminar burning velocities are achieved by:
1) slightly richer mixtures of 13 : 1 and below or
2) squish promoting in-cylinder turbulence in the charge during combustion.

Written by Somender Singh

Many people want to know how they can cut their own groove. Here are a few steps to the process.

1. Chose a engine design that has some form of squish or quench. Consider both the piston top and combustion chamber when deciding.
(e.g. dish pistons reduce the squish percentage considerably.)

2. Run a base line test with regular pump gas and production compression ratio, normally near 8.5 to 9:1?

3. Remove the heads and raise the compression ratio up to 10:1, milling the heads is the preferred method.

4. Cut one groove in each combustion chamber like the one in the pictures. So far the straight channel seems to work the best.
(if using a SBC, the grooves will measure near 1 cc volume)

5. Ensure a 0.070" piston to head clearance is maintained, measured at engine assembly. Most hot rodders are tempted to raise the compression by reducing the piston to head clearance down to 0.040" or less, resist the temptation.

6. Bolt the heads back on, leave the tune up alone, measure and report the improvements. Check and correct the tune up as needed and put it in a car (dyno's aren't much fun). You won't be dissatisfied.

Unless you have a very small combustion chamber volume or a piston with a mini-dome, it would seem hard to raise static comp. to 10:1 whilst increasing quench to 0.070".

I have 65cc quench type chambers, but have dished top pistons to maintain static compression ratio.

Do you think he is talking about running very small chamber volume along with a quench type chamber and flat top pistons 0.030" down in the hole at TDC with head gaskets having 0.040" compressed thickness? Hmmm...... Interesting......

Also, (assuming this works as claimed)...... Regarding the groove(es)...... Which would be more efficient...... The single groove or the 2 parallel grooves?

Happy Motoring,

#2 Turbulance and Combustion Dynamics












port throats generally run 80%-85% of total valve diameter because you need to maintain sufficient valve seat contact area to allow sealing and cooling and some wear during operation

due to the limited time and inertia of the air/fuel mix theres usually minimal flow loss due to overlap



Good atomization is a big part of the picture – but it is far from all of the picture





David Vizard

In the first installment of this series we looked at the possible compromise of supposedly advanced combustion dynamics versus flow and found all was not as it was propertied to be. We also covered more homogenous mixture preparation (mixture quality) and found it was not a guaranteed route to HP. We also looked at the 1970 Chrysler UK Avenger engine and its dire need for a very well prepared mixture of air and fine droplets. The consequences of not having such was a loss of as much as 10/% of the engines potential output. To move on from here and stay on track lets look at the last two sentences of part 1:

‘I may not have understood but a fraction of what was needed about combustion dynamics at this stage but one thing was for sure. Just knowing a little more than the opposition not only allowed me to make my car go significantly faster but also to slow all the other Avengers in the class I was to race against on road course the following year.’

To see how this was done and a patent taken out on what might well qualify as the one of the best candidates yet for a 100 mpg carb let me refer you back to the Avengers induction system. From part 1 it was stated:

‘To get a good part throttle burn and clean exhaust the intake charge, delivered by the twin inch and a half Stromberg’s, was heated. This was achieved by having the intake manifold bolted to the exhaust manifold. Between the two was a 1/16 thick plate with a hole in it. Through this hole the exhaust flame physically played onto the under side of the intake forming a very hot spot. This, at part throttle, probably was sufficiently hot to vaporizing all of the fuel at any sane street or highway driving speeds. OK this might sound like stock boring stuff but now we come to the crux of the matter. My first discovery was that if the hot spot was semi eliminated by replacing the 1/16 thick plate with a hole in it by similar plates with no holes the power dropped from 78 RWHP to 74 even though the charge temperature dropped a whole bunch. With a quenchless chamber I thought that this might be the case and this test suggested to me that this type of chamber needed to have a fair amount of vaporized fuel and the rest delivered in really well atomized form’.

The crux of the matter is I used this piece of unlikely information to my advantage in more ways than one. First the rules for Production Sedans stated that the air element of the engine must remain stock but the fuel element could be changed as required to get the necessary fuel/air mixture characteristics. In other words the fuel side of the induction was free. This meant I could do whatever I wanted to the jets and needles used to calibrate this carb.

It was evident that two specific factors had to be accomplished here. The first was to convince other competitors using an Avenger GT (that was all the front runners) to replace the spacer between the intake and exhaust with one that did not have a hole in it. The second was to atomize the fuel so much better that it compensated for the hot spot in terms of mixture quality.

OK let’s start on the first deal here. How to slow down the opposition by convincing all and sundry to go to a no-hole spacer between the intake manifold and the exhaust. The first point here that made the execution of this plot easier is that I had not made it publicly known in any article that the Avenger lost 4 hp when the intake charge was cooled. The normal assumption would be that cooling the intake would increase power.

The race season in the UK starts late April for most clubs so about November - December the previous year I started calling a few influential Tech Inspectors all over the UK. The supposed reason was that I needed some info on this or that and it was their area of expertise. In reality the call was just a ploy to sow the seeds of a rumor that I would subsequently use to my advantage. During the conversation with each tech inspector I would ask if they had heard the rumor that these Avenger engines with there quenchless combustion chambers were prone to detonation when driven flat out for any length of time (as per a 25 mile road race). I told of rumored melted pistons and the like and made it sound pretty bad. After planting the seeds of this rumor I let a few weeks go by then called the chief tech inspector and during a conversation about some other matter asked if he had heard the Avenger’s rumored detonation and melted piston situation. Surprise - surprise – he had and from about half a dozen people at that. He mentioned that the situation seemed pretty bad and it would be a possible problem for those running Avengers. Since, on paper at least, it seemed like the most competitive car for the class he expected that there would be quite a few in the field of entries. So far so good – everything is going to plan. At this point I told him I had a simple fix for this detonation/melted piston problem (which of course did not actually exist) “The problem” – I told him, “stems from the excessive heat put into the intake charge by the hot spot. By using a blank spacer instead of the one with the hole in it the problem goes away.” His answer went something like this “and the cooler charge might just give you the edge as well”. My response – “Well I thought that if I do a drawing of the change you can send it out to all those running Avengers and tell them that this modification is acceptable in view of the susceptibility of these engines to melt pistons under race conditions. If all competitors are using it the playing field should once again be level”. This he agreed to and a drawing of the mod was in time sent to all those competing in an Avenger. Now ask yourself – the apparent promise of a little more power from a cooler charge and the avoidance of a melt down – all for under a ‘dollar forty nine’ – who wouldn’t use it? Part one of the plot is now complete.

On to Part Two of the Plot.

With the cooler intake the next deal was to make design changes specifically to the Stromberg’s jets to more finely atomize the fuel. What I came up with warranted a provisional patent (that was the way it was done back then). In a nutshell what I did was to redesign the base of the piston and the bridge that forms the venturi. The following drawing shows roughly what was done.

By moving the jet upward away from the bridge and knife edging the discharge hole the fuel was both atomized and dispersed far better. This resulted in an appreciable increase in output on the Chrysler Avengers quenchless chamber engine.

The key to the increased atomization and reduced wet flow was the sharp and microscopically ragged edge of the jet and the relocation of the discharge point within the carb body. By moving the jets discharge point away from of any nearby surface that the fuel might attach itself too prior to becoming dispersed in the air produced better down stream dispersion. So how was the atomization produced by this set up? The fuel left the jet like a fog. Where as flash photo’s had revealed small droplets in the stock Stromberg’s discharge from the jet the revised design showed none as the droplets were far too small to show as such. On the chassis dyno the results were very encouraging. With the hot spot in place the new carbs dropped about 3 - 4 hp. With the hotspot blocked an increase in output, by virtue of an increase in torque, amounting to some 14 hp was seen! So what we are seeing here is that mixture preparation in conjunction with temperature, is contributing to a better combustion process to the tune of some 20% increase in output. In addition to this drivability, throttle response and part throttle fuel economy were all improved.

At this point the value of not only a cool intake charge carrying the correct mixture but also one having (on average) appropriate fuel droplet sizes and dispersion of such for the engine concerned is showing to be a distinct advantage. So how did all this work out on the track? The carb changes along with a whole host of selected and/or blue printed parts netted a totally legal engine that would leave even highly illegal cars for dead in the water. The first time out with this engine we put the champ from two years previous who was supposedly the 'King of Mallory" , with, what we later found to be a highly illegal engine, down by about 200 yards per lap.

Here I am racing with the reigning Champ, Bill Sydneham at a wet Silverstone. First place was decided by the width of a headlamp bezel. I had a number of door handling races with Bill but, because of his clean sportsman like driving, never got a single ding in the bodywork.

At this point we could ask if this quenchless chamber engine was something of an enigma. Were we fixing some inherent shortcomings that showed very positive results on this engine but would less likely show as much on more conventional engines? Well that could be but if this engine was, so to speak, acting as a magnifying glass on combustion dynamics it is still a good tool with which to work. However later down the road it was found that this seemingly odd-ball engine was not so far from a mainstream case as might at first be believed.

British Touring Car Championship Year.

After our dazzling show of speed during the last few races of the previous year Chrysler’s race boss, Des O’Dell, gave my three man team a car and all the factory parts we may need to build a BTCC car. For the US readers this championship is for a manufactures title and is contested on an international level. It’s a bit like having Cup Car racing with every major world manufacture competing.

Built by myself and my crew ( Colin Ashdown-Pogmore and Hugh Murray) this Group 1 Avenger, with it’s all iron pushrod engine, proved to have the speed to be more than a match even for the twin cam sporty cars from Italy, Germany and Japan.

We were up against twin cam engines of Alfa Romeo, Lancia, Renault, Toyota etc as well as the big bucks of Ford motor company, GM and the like. How did we do – we came out of the gate fast and by about the forth race our two buck, all iron, pushrod powered shopping car was a better rocket by far then the competition’s cost no object twin cam sporty specials. Did we win any race’s – hell no. Our competition’s engines barely made 8000 rpm. Our first engine of the year had a shift point of 8800! By the middle of the season we were turning this pushrod engine to 10,500 and, between two corners at Brands Hatch, to 11,000 rpm. What that meant was during our test sessions (i.e. the race) we broke about one each of everything that could break, sure we would have the fix by the following race but that did not exactly help our cause on the day. Also we were running these races as part timers. From Monday morning to Thursday, 8 am to 6 pm, we all had full time jobs.

Here’s Thruxton, a well know UK track some 120 miles west of London. Not only was I fast here but was also given an unofficial title of wheel lifter of the event. On one 110 mph plus turn I balanced the car like this for a couple of hundred yards - every lap!

From 7 pm till 1 or 2 am the following morning and Thursday to Sunday evening we either worked on the car or raced. Our budget for the year was less than what most teams spent per race. By the end of the year our team had managed a second place plus a couple of thirds, a class pole and half a dozen fastest laps. During six races the car had broken in a new engine in practice which put it on the grid on either the last row or last but one row. Sounds bad at this point but we were breaking in. The good part is that before the end of the first lap the number 66 was the class leader! I said the car was fast – and fast is exactly what I meant. So where did all this speed come from? No one thing in general but I can say that cylinder head flow, especially low lift flow, was significant along with cam design, exhaust and, very important, mixture characteristics and combustion dynamics. Let’s start with mixture characteristics.

Starting Point - Weber Revamp.

The homologated (that means the ones the car is supposed to have stock) carbs are a pair of side draft Weber DCOE 40’s. These came equipped with 30 mm main venturis. The rules allowed us to change main ventures for any design we wanted but the hole had to be no-bigger than 30 mm. Also the auxiliary (booster) venturi was free. This gave me scope to make new auxiliary venturi’s based on what I had learned from the Dellorto design mentioned in part 1. The result was a 6 – 7 hp increase throughout the rpm range over anything that could be built using off the shelf Weber parts. At this point I concluded that I had achieved about as good a fuel atomization as the engine needed so attention was turned to the cylinder head.

One of the factors we were stuck with was a 0.390 valve lift. This meant intake valve acceleration and flow, especially from low lift became important. Now I guarantee you will hear arguments countering the value of low lift flow but before the year is out I will have shown both theoretically and in practice that this is totally wrong. This Avenger engine is the first part of proving low lift flow is important. I won’t go into too much detail here because combustion dynamics is the subject but suffice to say that the low lift flow on my head was about 40% more at 0.050 than the Cosworth head while the flow at full lift was identical. Although there is more to it than just low lift flow it’s worth noting that my Avenger head made 11 hp more than the Cossy one and that was what the factory used the following two years!

The intake port was critical – here I used a much smaller port than the competition and it was rough finished with an out-of-round 80 grit ball wheel. The rough surface produced cut the tendency of the fuel to coagulate and form rivulets prior to entering the cylinder. Notice I say it cut the tendency – it did not cure it by any means – just made it a lot better.

With the mixture and intake port situations addressed and reasonably fixed it was time to look at the combustion chamber. I felt we just had to be able to do a better job in terms of power than the stock chamber. As it happens the rules specified such things as valve sizes, compression ratio etc but did not specify combustion chamber shape. This being the case we started finding the heaviest pistons (there was a lower weight limit and factory original pistons had to be used) and bringing them down to weight by machining the piston crown. What this did is allow the top ring to be nearer the piston crown thereby cutting the ring land volume. That little space is, as we shall see in part 3 of our combustion dynamics, way more influential than you may suspect. This becomes apparent as I go through the scenario I intend to use to explain such.

I addition to the piston mod the chamber form was also investigated. On the flow bench it was found that better flow could be had by forming a shallow chamber around the intake and exhaust valves. This necessitated machining the top of the block to get back to the 9.9/1 (as I remember) CR called for. This move was done a step at a time from one build to another. Essentially we were building, for race and R&D combined, about 1-½ engines per race. Each time a build or rebuild was done the chamber in the head was increased and the chamber volume residing in the block reduce by machining the block deck. Each time this was done the package more closely approach a conventional chamber with squish. At each new spec some 0.020 more material had to come off the top of the block to bring the CR back up to 9.9/1 and each time more power was seen. When the situation got to where the piston was 0.080 down the bore, which produced the best results to date, I decided that it looked worthwhile to go the whole hog here and put the entire combustion chamber into the head and deck the block for a tight quench rather than possibly do 4 more builds to get there. The results on the dyno were just shy of startling. If all had followed previous form I would have expected about 6 hp more from this combo – instead it was 8 less!

So why am I highlighting these negative results? Simple – I want to emphasis that the subject we are dealing with here is far from simple. I had no idea why power went down then and here we are 30 years later and I am still shy of an answer. In this instance the results were about 180 degrees apposed to all the other tests I have been involved with. This Avenger engine liked to have the piston stop 0.120 (120 thousandths) short of the head face (0.080 down the hole and a 0.040 head gasket) for best results. For just about every other engine I have done tests on like this that piston to quench source gap is about the worst in terms of low detonation resistance and poor combustion. It really begs the question as to whether or not we can give an engine too much quench action. It is this factor, and crevice volumes such as the ring land volume that we will look at in the next installment.
#3 Turbulance and Combustion Dynamics

#3 Turbulence and Combustion Dynamics

OK, lets be frank here. It is OK to start an article with OK especially if a) you have little idea what the rules concerning English grammar are and b) there is no editor looking over your shoulder. But to get going here let’s take a look at where we left off in part two of this subject. This, in yellow, is as follows.

So why am I highlighting these negative results? Simple – I want to emphasis that the subject we are dealing with here is far from simple. I had no idea why power went down then and here we are 30 years later and I am still shy of an answer. In this instance the results were about 180 degrees apposed to all the other tests I have been involved with. This Avenger engine liked to have the piston stop 0.120 (120 thousandths) short of the head face (0.080 down the hole and a 0.040 head gasket) for best results. For just about every other engine I have done tests on like this that piston to quench source gap is about the worst in terms of low detonation resistance and poor combustion. It really begs the question as to whether or not we can give an engine too much quench action. It is this factor, and crevice volumes such as the ring land volume that we will look at in the next installment.

We can, from the forgoing, draw, with some degree of certainty, a tentative conclusion. Namely that applying a universal rule can, just once in a while, turn around and bite your rear end. Just to keep my feet on the ground I have a little policy that all lessons learned – no matter how certain or obvious they may seem, should be regularly reappraised for anomalies that may actually reveal something that was overlooked or misinterpreted.

If the big fuel droplet deal with the Mini Cooper engines intrigued you here’s another somewhat mystifying case concerning – once again – a Mini engine. After a successful season with a 1293 cc Mini Cooper S hill-climber (finished second in championship and only narrowly missed first spot due to going of on honeymoon and missing a round) a customer asked if I would build him a blown, bored and stroked version of this engine. I did just that – capacity was stretched to 1442 cc and a large Shorrock supercharger installed. Just so that we did not have to use an intercooler as there was no room for such, the boost was limited to about 12 psi. Dyno testing was to be on a chassis dyno. The successful engine from the previous year made just on 100 hp at the front wheels. After a break in period and a change of oil and plugs to a race grade type the engine was given its first wrestling match with the dyno. This resulted in only 85 hp in spite of the 12 pounds of boost going into this big ‘A’ Series engine. This was scary – the customer was standing right there watching and here we were with a car that sounded for all the world like it could break the unlimited land speed record yet it was way down on even a conservative estimate of what it should make.

As it happened Mike Lane, GFN’s F1 correspondent, had helped out a lot on this build and was on hand during the dyno testing. Mike had built one of his slick close ratio ‘knife-through-butter’ shifting gearboxes for this car. Between us we waded into a search for all the likely causes of great noise – no power. Ignition timing, valve lash, ignition box etc. (had an early two spark transistor ignition box designed by the guy who was scientific advisor to the war ministry – the guy had an IQ that was about off the scale and our box was about like a lightening bolt generator) were all diligently checked for function and ruled OK. At several points along the way we made dyno runs but with the same results – about 85 hp. Finally we got all the way down to pulling the front end off the engine to check cam timing. It was right where we thought it should be. During reassembly some inadvertent throttle pumping flooded the motor. With the engine virtually cold it may not have fired up to well on these super cold grade plugs. If they were also wet it certainly wasn’t going to be that happy from a cold start. So the Champion race plugs were pulled and an equivalent heat range of Autolite race plugs installed.

When fired up the motor sounded no less and no more wicked than before but the dyno numbers were – at 142 hp - nothing short of a techno shock. Although a pleasant surprise it was very much a case of ‘what is the world is going on here?’ This was such a surprise that the Champions were re-installed and re-tested. Same killer sound – just 85 hp! Now I have to tell you that in just about every other Mini application Champion plugs were as good as or better than anything we could find but here was an anomaly. This engine apparently did not like anything with Champion written on it. Anyway with no further ado we went on to finely tune the big carb on the engine and get the timing right on the money. After being tuned on for about 3 hours this engine ended up pumping out just over 170 hp at the wheels. I had hoped for about 185 but that’s dyno testing for you. It’s like having an eye glass to clearly focus on reality. This may not always be as gratifying as fantasy but regardless of positive or negative results you do get to learn a lot more of what it takes to win races.

Atomization Optimization.

I realize that I had mentioned looking at crevice volumes at the start of this part of our investigation but before going there let’s take another look at atomization. First it is easy to jump to the conclusion that the better the atomization is the better the power output will be. If only it were this easy! In reality it is far more accurate (but still not 100% true) to say that as the fuel is better atomized (and/or vaporized) the better the Brake Specific Fuel Consumption (BSFC) will be. This number should not, as is so often the case even with pro engine builders, be confused with the mixture. It is only roughly connected to the mixture. It is in no way a measurement of mixture ratio, only a consequence of such. To get a better understanding of this read Dusty’s upcoming story on the in’s and out’s of BSFC.

At this point the question is ‘can the fuel be atomized and vaporized too much?’. Let me set the scene. It’s about 1977 and I am just starting testing some of the trick carbs built by Tucson’s premier carb builder/designer Dave Braswell. The year before at the 76 SEMA show I got to talk with Holley’s then chief engineer Mike Urich. In our conversation I was amazed to find that, as far as Mike knew, Holley had done no official research on the effects of booster design on fuel atomization. I mentioned this somewhat surprising tidbit of info to Dave Braswell and he immediately volunteered assistance and carbs to do some testing on what we perceived as a typical street tuned small block Chevy. Here is how things unfolded.

The tests involved two carbs, each about 750-800 cfm. One had high gain fine fuel spray dog leg style boosters and the other typical low gain courser spraying straight leg boosters.

Shown here are the most common style of boosters found in a Holley 4150 series four barrel carb (the BG boosters follow a similar range of patterns). They are shown from #1 to 5 in order of signal strength per cfm of flow. Usually the greater the gain the better the boosters ability to atomize fuel. A point here is that the stepped dog leg (#3) does a significantly better than #2 even though the signal strength is very similar. This is because the step on the underside has an enhanced fuel shearing capability. Atomization however is not dependant solely on the booster design. A high gain booster works more effectively with a larger air corrector bleeding into the emulsion well of the carb (shown right) The more the fuel can be emulsified prior to the booster the better the atomization is likely to be.

The engine was run with three intake manifolds types. The first was the stock exhaust heated and consequently hot running intake. The second was an aftermarket two plane aluminum intake with the heat crossover blocked off. The last was a Victor Jnr intake which of course, being an air gap style intake, was significantly cooler running yet.

On the stock intake, which was also the hottest by far, the tricked up Braswell carb lagged the nearer stock carb with it’s bigger fuel droplets, by some 8 hp (nominally a 360 horse engine) but the fine fuel delivered by the trick booster carb produced, by a small margin, the best BSFC both at WOT and part throttle cruise. On the heat blocked after market two-plane the carbs were very close in terms of output but we are still considering a relatively hot running manifold here. The BSFC with the fine fuel droplet booster carb was as much as 8% better especially at part throttle. On the cool running Victor Jnr the finely atomizing booster equipped carb was unbeatable anywhere in the rpm range. It made about 12 hp more and the brake specifics were all better (lower) numbers by a substantial margin.

The left three boosters show the typical form that most boosters follow for a typical down draft US production style carb regardless of brand. The center pair of boosters are ‘dog-leg’ Holley boosters as viewed from the underside. The one on the right has a step just before the emulsion exit hole. Although it does little to boost the signal this step does atomize the fuel far better. On the left is a BG booster. Many of the BG carbs have this style of remove and replace booster making this type of carb very useful when doing development work.

So what does this tell us? The results indicate that there is an optimum fuel droplet size that balances the need for some vaporization against the need not to evaporate too much fuel and spoil the engines volumetric efficiency. Hot running engines can offset the negatives of big fuel droplets from the boosters but cold ones cannot. Cold intakes need the ratio of the fuels surface area to volume increased (which is just what happens as the fuel droplets get smaller) so that the loss of the heat as a vaporization source is compensated for by increased in the fuel’s evaporative surface area.

Let’s skip along here a few years. In the early 90’s I got involved with booster development with the Carb Shop in California. The plan was to develop a Super Booster that not only gave a big signal but also did not obstruct the flow of the main venturi to any greater extent than a regular high performance booster. If a high gain booster can be used it means that for any given application a bigger carb can be used for more top end before drivability and low speed output suffer. Well the program produced some trick looking high gain boosters which just before Christmas (and unknown to me) found their way into the carb(s) of a front running Cup Car team. On the dyno in the crisp December days just before Christmas in Mooresville NC (For the benefit of those non-NASCAR folk this is the ancestral home of all Cup Car teams) these boosters paid off in the teams Daytona 500 engines to the tune of about 10 hp. So it was with great expectations that the team headed off in early January for the Daytona 500 in Florida.

It was a hotter than usual January that year in an otherwise normally hot Florida. With the new carb the car was well off the pace. In frustration the trick booster carb was replaced with the old one and the car immediately ran on the money. The lesson here is that you can absolutely guarantee that too much of a seemingly good thing is – well not so good. The percentage of fuel atomizing prior to entering the cylinder was such that any gains in better combustion were overridden by the drop in VE caused by the added fuel vaporization taking place within the intake manifold.

So when is a high gain booster any good to a race car engine – maybe rarely if ever as things stood for a typical Cup Car engine of the early 90’s but let’s move on a little.

About this time (early 90’s) I am heavily involved in thermal barriers. It’ something I have looked at on and off since Formula Ford days in the late 60’s. There we found a temporary 2 hp (it barley lasted a race) by using high temp exhaust paint on the pistons (Sperex I believe). In this case a relatively extensive study was made of the effect of thermal barriers in race intake manifolds.

Using a single 4 barrel carb on a single plane intake the effects of various boosters with the intake were explored both with the intake runners ‘raw’ and with them thermal barrier coated. The booster that worked best with the raw runners was of the stepped ‘dog leg’ variety shown in the earlier drawings. When the manifold was coated and used in conjunction with this booster the power figures were within about a horsepower or so – unchanged.

So what’s the deal here? With the raw ports and mixture temps measured on #2 runner it was seen, from a carb intake temp of 84 degrees F, a drop to 55 degrees F due to the evaporation of a portion of the fuel. By dividing the plenum front to back and using one end of the engine to drive the other (and no fuel to the front float bowl) it was found that the air at #2 without any fuel picked up (allowing for a few corrections) about 10 degrees of manifold heat (more at low rpm and less at high). With the coated manifold the motored #2 intake runner temp was between 5-8 degrees less so the thermal barrier was doing what it was supposed to – keeping heat out of the intake charge. What was not happening here was the realization of any increased power due to the cooler charge. When the charge temperature was measured on a functioning #2 cylinder the drop in temperature from the carb to head/manifold interface of #2 runner was only barley changed and if everything had been working as before it should have been at least 5 degrees cooler. What this indicated was that the cooler running intake was not allowing as much fuel to vaporize and therefore the added wet fuel arriving at the cylinder was compromising the combustion process. At this point high gain annular discharge boosters along with appropriate (bigger) air correctors were installed. With the same air to fuel ratio the better atomization restored the % of vaporized fuel entering the cylinder. The temperature at the #2 runner with an air fuel mixture passing through dropped to 49 degrees F. The dyno showed some meaningful gains at this point. Essentially the cooler running more finely atomized charge had the effect of jacking the entire torque curve in an upward direction. On a nominally 450 horse engine the torque at 3000 rpm rose by 11 lbs-ft (6.3 hp) and by 7 lbs-ft at 6200 rpm (8.3 hp).

The RS range of Barry Grant carbs, have replaceable sleeves (main venturis) as well as replaceable boosters. This allows the end user to fine tune the carb combination to a degree that should leave little on the table. On the right we have a Holley 950 carb with stepped dog-leg boosters. A carb equipped with such boosters is a good all-round choice as this style of booster is application versatile.

Conclusions to this point.

In this section of our look at ‘In Cylinder Combustion and Turbulence’ several things have become more evident. First we have not actually arrived at discussing much of anything about what happens to the charge after it is actually in the cylinder. The fact of the matter is that what happens within the cylinder can be greatly influenced by how the intake charge is ‘prepared’ prior to it’s arrival at the cylinder. Also the cases put forward here are one or two from maybe a dozen thermal barrier/booster/mixture preparation tests. All similar in their intent and all showing that the dyno testing rule ‘make only one change a time’ so often touted by do-it-yourself performance magazines is seriously flawed.

What is creeping in here is the worth of thermal barrier coatings. I have worked with many of the leading companies such as Swain Tech and Polydyne Coatings over the years but, since about 2002, spent a lot of time working with Calico Coatings in Denver NC. They have a great facility and being close I can go and visit to discuss whatever experiments I am into at that time. Without exception, they have been ready to help in such tests and that has allowed me to move along on coating tests at a rate that would have otherwise not been possible.

We have hit the subject of coatings and it is, as we have seen, relevant to our present topic but here it seems appropriate to branch the subject of ‘Coatings’ off into a separate category (thread). That is where we will go for an in depth look at what coatings can do for us.

The coating features though are still some weeks away - meanwhile if you need Calicoe’s phone # it is 704-483-2202

As for the ‘In Cylinder’ subject we can conclude that for a given fuel, ambient weather conditions and a host of other factors there is a certain ratio of wet to vaporized fuel that will be optimum for best output. Based on everything we have discussed from part 1 of this series to date we can say that keeping the fuel in suitably small droplets, allowing only a given % to vaporize and avoiding, as far as possible, wet flow streams, is a major factor toward increased power from an engine (other than a Mini that is!)

But before we wind up here there’s one more point I want to make. All the temperature measurements of the intake charge temperatures were done at the intake manifold to cylinder head interface. But heated or not the fact of the matter is the intake manifold is not the greatest source of heat input into the charge. A case comes to mind here during the testing of a 2000 cc Cosworth BDA engine about ’90. I had reason to turn off the dyno cell lights for a photo of the near white hot exhausts seen during a run on this 280 -281 horse injected engine. Being in the dyno cell with an engine turning 8500 plus rpm can be a little unnerving but as I walked past the deafening intake I realized that I could, down the intake stacks, see the intake valves glowing very dull red. At this point it had not occurred to me just how hot the intake valves could get but consider this. Since the intake valves are of greater area than the exhaust they would, during the combustion cycle at least, pick up more BTU’s of heat that the exhaust. And where do they dump a whole load of this heat? You guessed – they put it right into the intake charge. This results in one positive and one negative. First the positive – that is the fuel is further vaporized before entering the cylinder (by how much I have little idea) and second the intake charge is heated - and that is not so good. So I asked myself what would the trade off be between these possibly competing factors. The head was pulled and the intake valves alone were coated. A week later the results were in. First a look down the intake stacks in a darkened cell revealed that the valves were no longer visibly glowing. I fully realize that this is hardly a really scientific way to measure the valve temperature but that was all I had to do the job. The fact they were no longer visibly glowing meant, at a good estimate, they were probably a hundred degrees or so cooler. Secondly power figures at 283-284 hp, were slightly up over those with the uncoated valves.

This is an example of the head I did for Ryan Garcia’s giant killing 79 Mitsubishi Lancer. The Calico Coatings applied thermal barriered valves are clearly seen here. What is not so obvious is that this head is my patented Poly Quad design that, with no additional moving parts, emulates a Honda V-Tech variable valve timing.

So what we see here is that with high pressure fuel injection (60 psi in this case) and line-of-sight ports exiting into an open area of a combustion chamber with no chamber wall directly in the path of the entering charge we see an increase in power. The question is are the wet flow dynamics of a typical two valve V8 head such as to deliver similar results or do we need to see greater fuel shredding at the seat and/or greater heat input to vaporize more of the charge by this or other means to get the benefits of cooler intake valves? That, along with crevice volumes (this time for real) is what we will get to next.

David Vizard
I recently had a conversation with a guy who said his machine shop suggested he run an exceptionally tight quench distance, the "machinist" suggested that the piston should just barely miss impacting the heads and the tighter you get the quench distance the better off you would be,in fact suggested a .010-.015 piston to head clearance..using a .017 shim head gasket and the pistons sticking a few thousands above the block deck, as the ideal combo... this guys obviously a BULL SHIT ARTIST in my opinion, with little or no real world experience in building engines
the fact is that the previously suggested .038-.044 quench is about the minimum doable clearance due to both piston rock in the bore and rod stretch at high rpms, I don,t know a single factory produced engine that even has that tight of quench.

when you hear stuff you might suspect is "BULL SHIT" concerning an engine clearances or valve train geometry...
STOP and get several EXPERIENCED and KNOWLEDGEABLE opinions,
don,t take some guys opinion until you verify the infos correct,
do some research and consult several good knowledgeable sources,
not a couple of your drinking buddy's at the local hang out!
get out a feeler gauge and look at how thick a .042 feeler gauge is and...
think a bit,
what happens when metals heated several hundred degrees..it expands!
what happens when you put thousands of pounds of tension on rods and rod bolts/..they STRETCH!
what happens to rod bearing clearances, and piston pin clearances under high loads?
cranks flex, its not going to take a genius to realize that a combination of factors can significantly reduce piston to head clearances reducing the quench distance at higher rpms in a running engine!






do you really want a 70lb-90lb rotating assembly spinning at 6000 or more rpm to have the pistons just avoiding contact with the cylinder heads , by even less than .038-.044 inches, when you start out, checking those clearances? because it darn sure will be tighter during the actual operating conditions once the engines up to temperature and spinning at 6000 rpm or higher??
common sense will tell you that theres both machining errors, metal stretching under high stress and other factors, can significantly reduce piston to head clearances reducing the quench distance at higher rpms in a running engine!
I have seen plenty of cylinder heads run on engines with .040 or less quench that on disassemble showed witness marks of the piston just kissing the heads even with that clearance.
Last edited by a moderator:
Same kind of advice was going on in the Pontiac V8 crowd a few years back Grumpy.
Run .005 or .000. Quench.
Many respected Pontiac engine builders recommended.
I found advice insane.
Fel pro 1016 Pontiac race headvgssket.039 compressed.
Machine block for true zero deck height
Runs great and no pistons kissing the heads.

hi grumpy , will like to say thanks for the information
, it has helped me build a good fast running engine, this form is very informative, thanks again
always glad to help out!....why not get your buddies to register




read these threads and pay attention to the pictures










https://www.memoparts.com/img/cms/Documents/Piston Failue.pdf

info you can use
Engine Building
There isn't a universal set of rules that govern all engine building. The following is information that has worked successfully and should be considered when building a performance engine.

A high performance race engine, by its definition, indicates that limits are going to be pushed. The limit that is of most concern, as far as pistons are concerned, is peak operating cylinder pressure. Maximizing cylinder pressure benefits horsepower and fuel economy. Considering the potential benefit, owners of non-race engines, from motorhomes to street rods, also look to increasing cylinder pressure. Increasing the compression ratio is one sure way of increasing cylinder pressure but its not the only way. Camshaft selection, carburetion, nitrous and supercharging can all alter cylinder pressures dramatically.

Excessive cylinder pressure will encourage engine destroying detonation with no piston immune to its effects. The goal of performance engine builders should be to build their products with as much detonation resistance as possible. An important first step is to set the assembled quench distance to .035". The quench distance is the compressed thickness of the head gasket plus the deck height, (the distance your piston is down in the bore). If your piston height, (not dome height), is above the block deck, subtract the overage from the gasket thickness to get a true assembled quench distance. The quench area is the flat part of the piston that would contact a similar flat area on the cylinder head if you had .000" assembled quench height. In a running engine, the .035" quench decreases to a close collision between the piston and cylinder head. The shock wave from the close collision drives air at high velocity through the combustion chamber. This movement tends to cool hot spots, average the chamber temperature, reduce detonation and increase power. Take note, on the exhaust cycle, some cooling of the piston occurs due to the closeness to the water cooled head.

If you are building an engine with steel rods, tight bearings, tight pistons, modest RPM and automatic transmission, a .035" quench is the minimum practical to run without engine damage. The closer the piston comes to the cylinder head at operating speed, the more turbulence is generated. Turbulence is the main means of reducing detonation. Unfortunately, the operating quench height varies in an engine as RPM and temperature change. If aluminum rods, loose pistons, (they rock and hit the head), and over 6000 RPM operation is anticipated, a static clearance of .055" could be required. A running quench height in excess of .060" will forfeit the benefits of the quench head design and can cause severe detonation. The suggested .035" static quench height is recommended as a good usable dimension for stock rod engines up to 6500 RPM. Above 6500 RPM rod selection becomes important. Since it is the close collision between the piston and the cylinder head that reduces the prospect of detonation, never add a shim or head gasket to lower compression on a quench head engine. If you have 10:1 with a proper quench and then add an extra .040" gasket to give 9.5:1 and .080" quench, you will create more ping at 9.5:1 than you had at 10:1. The suitable way to lower the compression is to use a dish piston. Dish (reverse combustion chamber), pistons are designed for maximum quench, (sometimes called squish), area. Having part of the combustion chamber in the piston improves the shape of the chamber and flame travel. High performance motors will see some detonation, which leads to preignition. Detonation occurs at five to ten degrees after top-dead-center. Preignition occurs before top-dead-center. Detonation damages your engine with impact loads and excessive heat. The excessive heat part of detonation is what causes preignition. Overheated combustion chamber parts start acting as glow plugs. Preignition induces extremely rapid combustion and welding temperatures melt down is only seconds away!

For a successful performance engine, use a compression ratio and cam combination to keep your cylinder pressure in line with the fuel you are going to use. Drop compression for continuous load operation, such as motorhomes and heavy trucks, to around 8.5:1. Run a cool engine with lots of radiator capacity. Consider propylene glycol coolant and low temperature thermostats. Reduce total ignition advance 2 to 4 degrees. A setting that gives a good HP reading on a 5 second Dyno run is usually too advanced for continuous load applications. Normally aspirated Drag Race engines have been built with high RPM spark retard. The retard is used to counter the effect of increased flame travel speed with increased engine heat. "Seat of the pants" spark adjustment at low RPM will almost always cause detonation in mid to high compression engines once they are rung out and start making serious horsepower. Set spark advance to make best quarter mile speed not best ET, usually 34 degrees total advanced timing.

Top Ring End Gap is often a major player when it comes to piston problems. Top ring butting under high load and heat conditions can destroy the piston top land. Most top land damage on race pistons appears to lift into the combustion chamber. The reason is that the top ring ends butt and stick tight at top-dead-center. Crank rotation pulls the piston down the cylinder while leaving at least part of the ring and top land at top-dead. Actual end gap will vary depending on the engine heat load.

Lean mixture, excessive spark advance, high compression, low capacity cooling system, detonation and high HP per cubic inch all combine to increase an engine's heat load. Most new generation pistons incorporate the top compression ring high on the piston. The high ring location cools the piston top more effectively, reduces detonation and smog, and increases horsepower. If detonation or other excess heat situations develop, a top ring end gap set to the close side will quickly butt, with piston and cylinder damage to follow immediately. High location rings require extra end gap because they stop at a higher temperature portion of the cylinder at top-dead-center and they have less shielding from the heat of combustion. At top-dead-center the ring is above the cylinder water jacket.

If a ring end gap is measured on the high side, you improve detonation tolerance in two ways. One, the engine will run longer under detonation before rings butt. Two, some leak down appears to benefit oil control by clearing the oil rings of oil build up. Clean, open oil rings are necessary to prevent from reaching the combustion chamber, which is also why we do not like gapless rings. A very small amount of chamber oil will cause detonation and produce significant horsepower loss. Top ring gaps can be increased 50% with hypereutectic pistons.

Ring Options of 1/16" or stock 5/64" are offered on most performance pistons. The 1/16" option reduces friction slightly and seals better above 6,500 RPM, while being considerably more expensive. Stock, (usually 5/64" compression rings), work well and help with the budget.

Piston to Bore Clearance for hypereutectic pistons were Dyno tested at wide open throttle with .0015", .0020", .0035" and .0045" piston to bore clearance. After 7-1/2 hours the pistons were examined and they all looked as new, except the tops had normal deposit color. Even with 320 degrees Fahrenheit oil temperature, the inside of the piston remained shiny silver and completely clean. Excessive clearance has been shown to be safe with hypereutectic pistons. Loose Hypereutectic pistons over .0020" do make noise. As they get up to temperature they still make noise because they have very restricted expansion rate and do not swell up in the bore. The Hypereutectic alloy not only expands 15% less, it insulates the skirts from combustion chamber heat. If the skirt stays cool piston expansion is drastically reduced. Running close clearances is beneficial to piston ring seal and ring life. A small short term HP improvement can be had by running additional piston clearance because friction is reduced. To obtain actual piston diameter, measure the piston from skirt to skirt level with the balance pad.

Pin Oiling should be done at pin installation, whether it is pressed or full floating, prelube the piston pin hole with oil or liquid prelube, never use a grease. If you are using a pressed pin rod be sure to discard spiral pin retainers. A smooth honed pin bored surface with a reliable oil supply is necessary to control piston expansion. A dry pin bore will add heat to the piston rather than remove heat. Pistons are designed to run with a hot top surface, and cool skirts and pin bores. High temperature at the pin bore will quickly cause a piston to grow to the point of seizure in the cylinder.

Marine Applications require an extra .001"-.003" clearance because of the possible combination of high load operation and cold water to the block. A cold block with hot pistons is what dictates the need for extra marine clearance.

"Compression Ratio" as a term sounds very descriptive. However, compression ratio by itself is like torque without RPM or tire diameter without a tread with. Compression ratio is only useful when other factors accompany it. Compression pressure is what the engine actually sees. High compression pressure increases the tendency toward detonation, while low compression pressure reduces performance and economy. Compression pressure varies in an engine every time the throttle is moved. Valve size, engine RPM, cylinder head, manifold and cam design, carburetor size, altitude, fuel, engine and air temperature and compression ratio all combine to determine compression pressure. Supercharging and turbo-charging can drastically alter compression pressures.

The goal of most performance engine designs is to utilize the highest possible compression pressure without causing detonation or a detonation related failure. A full understanding of the interrelationship between compression ratio, compression pressure, and detonation is essential if engine performance is to be optimized. Understanding compression pressure is especially important to the engine builder that builds to a rule book that specifies a fixed compression ratio. The rule book engine may be restricted to a 9:1 ratio but is usually not restricted to a specific compression pressure. Optimized air flow and cam timing can make a 9:1 ratio but is usually not restricted to a specific compression pressure. Optimized air flow and cam timing can make a 9:1 engine act like a 10:1 engine. Restrictor plate or limited size carburetor engines can often run compression ratios impractical for unlimited engines. A 15:1 engine breathing through a restrictor plate may see less compression pressure than an 11:1 unrestricted engine. The restrictor plate reduces the air to the cylinder and limits the compression pressure and lowers the octane requirements of the engine significantly.

At one time compression pressure above a true 8:1 was considered impractical. The heat of compression, plus residual cylinder head and piston heat, initiated detonation when 8:1 was exceeded. Some of the 60's 11:1 factory compression ratio engines were 11:1 in ratio but only 8:1 in compression pressure. The pressure was reduced by closing the intake valve late. The late closing, long duration intake caused the engine to back pump the air/fuel mix into the intake manifold at speeds below 4500 RPM. The long intake duration prevented excess compression up to 4500 RPM and improved high RPM operation. Above 4500 RPM detonation was not a serious problem because the air/fuel mix entering the cylinder was in a high state of activity and the high RPM limited cylinder pressure due to the short time available for cylinder filling.

Before continuing with theory, a little practical compression information is in order. If you have a 10:1 engine with a proper .040" assembled quench and then add an extra .040" gasket to give 9.5:1 and .080" quench you will usually experience more ping at the new 9.5:1 ratio than you had at 10:1. Non quench engines are the exception to this rule. Some racers make the effort to convert non-quench engines to quench type engines, as with our Mopar Squish Deck Heads. Compression ratios that work are as follows:


8.5:1- Non-quench head road use standard sedan, without knock sensor.

8.5:1- Quench head engine for tow service, motorhome and truck.

9.0:1- Street engine with proper .040" quench, 200� @ .050" lift cam, iron head, sea level operation.

9.5:1- Same as 9:1 except aluminum head used.

Light vehicle and no towing.

10:1- Used and built as the 9.5:1 engine with more than 220� @ .050" lift cam. A knock sensor retard is recommended with 10:1engines.


12.5:1- Is the highest compression ratio suggested with unrestricted race gas engines.


15.5:1- Is the highest compression ratio suggested for unrestricted alcohol fuel engines.

Satisfactory use of 14:1 - 17:1 compression engines can be made when restrictor plate or small carburetor use is mandated by the race sanctioning. High altitude reduces cylinder pressure so if you only drive at high (above 4500 feet altitude) a 10:1 engine can be substituted for a 9:1 compression engine. General compression rules can be violated but is usually a very special case such as a 600 HP normally aspirated engine in a 1500 lb. street car with a 12:1 compression ratio. The radical cam timing necessary for this level of performance keeps low and medium RPM cylinder pressure fairly low. At high RPM detonation is less of a problem due to chamber turbulence, reduced cylinder fill time, and the fact that you just can't leave the above combination turned on very long without serious non-engine related consequences.

Piston temperature and horsepower are interrelated. High horsepower per cubic inch engines not only make more horsepower but they make more heat. How the excess heat is handled has a significant effect on total engine power and longevity.

Various piston, cam, valve, chamber and port configurations have been and are currently being tested to optimize engine internal temperatures. Some engines have ceramic exhaust port insulation coatings that allow cooler cylinder head operation while keeping exhaust temperatures elevated for efficient catalytic converter operation. The same ceramic type insulation on a piston top has been quite successful. Ideal piston temperatures in an operating engine would suggest refrigeration during the intake and compression stroke, and incandescence during the combustion and exhaust stroke. The advantage of a hot piston on the power stroke is that less combustion energy is going to be absorbed by the piston. So far, it is not practical to heat and refrigerate a piston 6000 times a minute. However, if the incoming air is not heated by the piston and the piston reflects the heat of combustion, you start to approach ideal conditions. A polished hypereutectic piston will reflect combustion heat back into the combustion process. This reflection, combined with the insulating qualities of the hypereutectic alloy, keeps the heat in the cylinder during the power stroke. A smooth polished piston runs cooler than a non-polished piston, even after combustion deposits have turned both pistons black. A cool, smooth piston will transmit a minimum of heat to the incoming fuel air mix.

The Hypereutectic piston gives the racer a real out of the box advantage with smooth diamond turned piston heads. A polish is relatively easy to achieve and does improve the already excellent reflectivity of the hypereutectic piston. If a buffing wheel is used, you will note a gray cast to the finished piston. The gray results from the exposure of the Silicon particles that are dispersed through the piston.

Experimental work to reduce piston heating of the incoming fuel mix has been very limited but, in theory, a thin ceramic coating may prove to be beneficial. A thin, smooth coating over a polished piston should still reflect combustion heat while reducing carbon buildup and protecting the piston polish. It is easier for a thin film to change temperature with each engine cycle than it is for the whole piston to do the same. A thin film can be cooled by the first small percentage of inlet fuel mix, allowing the main quantity of fuel mix to remain relatively cool. Tests have shown that polishing the combustion chamber, valves and piston top can increase horsepower and fuel economy by 6%.

All this polishing probably sounds counter to the practice of cimpling the combustion chamber. Dimpling has been show to put wet flow back into the air flow and improve combustion. We do not recommend dimpling, but do suggest cutting a small discontinuity close to the valve seat to turbulate wet flow. Some bench flowed cylinder heads encourage fuel separation at the inlet pot. If a small step is added at the valve seat to force the wet flow over the resulting sharp edge, fuel will reenter the air stream and give you the same affect as dimpling only without losing the benefit of a completely polished chamber. As you reduce wet flow you will improve combustion and most likely need to install leaner carburetor jets. Leaner jets compensate for the excess fuel that is available when wet flow is put back into the air/fuel mix. Significant additional horsepower gains can be had with careful attention to cylinder-to-cylinder fuel distribution by allowing all cylinders to be set "just right".

Combustion chamber design work has increased steadily the last ten years. Some of the work is mandated by the EPA and some is the result of race engine development. Some of the smog work has actually enhanced race engine development. Combustion chamber science is now more concerned with the effects of swirl, tumbling, shrouding of the valve, quench, flame travel, wet flow and spark location. A combustion chamber shaped dished piston can improve the flame travel in the combustion chamber. A dish allows the flame to travel further and expand more before it is stopped by a metal surface. This rapid flame travel makes it unnecessary to run big spark advance numbers. Ideally, we would like to be able to initiate ignition at top dead center since this would reduce negative torque in the engine that is now cause by spark advance. We are some time away from a practical spark ignition system that will make optimum power with a TDS setting. Some day it will happen. Don't go out and buy dished pistons for your open chamber 454. The advantage in flame travel is more than offset by the low compression ratio this combination yields. Small combustion chambers respond well to dished pistons, especially reversed dome or "D" cups. A 400 small block Chevy can use a 22CC D Cup piston and still have 10.4:1 compression. The trend in modern engine design seems to be smaller combustion chambers with recessed piston tops for more HP per cubic inch.

Ignition timing on most installations should be 34 degrees total with full mechanical advance dialed in. More advance may feel better off the line but the engine lays down as the combustion chamber components come up to temperature. At the drag strip set timing for maximum MPH not best ET. Too much spark advance will shorten the life of any performance engine, sometimes drastically.

Nitrous oxide can double the horsepower of most engines with less effort and money being spent than any other modification. Even the "smog people" are usually happy, as the nitrous is activated only during full throttle "open loop".

A nitrous engine can be built as a stock rebuild or it can be a dedicated effort to maximize the total performance package. As more power is generated, more waste heat, exhaust air flow and combustion pressures push the limits of engine strength. Often more beef is needed in the drive train and tires.

All stock factory engines are built with a safety factor when it comes to RPM, HP produced, cylinder pressure, engine cooling, etc. If you are only going to use a 100 HP nitrous setup on a 300 cubic inch or larger engine, built in factory safety factors are probably sufficient. As power output levels are raised engine modifications are usually prudent.

The most common mistake made when using nitrous oxide injection concerns ignition timing. A normally aspirated engine makes its best power when peak cylinder pressures occur between 14 and 18 degrees after TDC. Pistons usually require 34 degrees BTDC ignition timing at full mechanical advance to achieve proper ATDC peak cylinder pressure. The total time from spark flash to the point of peak pressure is typically 48 to 52 degrees. If an engine is producing 30% of its power from nitrous, the maximum cylinder pressure will occur too close to TDC to avoid run away to detonation. If ignition does not get retarded, good-bye horsepower and head gaskets. The key to getting max HP from a max nitrous engine is to shift the maximum cylinder pressure event progressively further after TDC.

Cylinder pressure of 1000 PSI at TDC, (FIG. 1), can drop to 500 PSI with less than 3/8" of piston travel, (FIG. 2). If you can manage to get 1000 PSI in the same engine after the 3/8" travel, (FIG. 3), the pistons will have to travel an additional 3/4" to lower the cylinder pressure to 500 PSI, (FIG. 4). Work is defined as a force times distance. An average pressure, (750 PSI X 12-1/2 sq. in.), times distance in feet, (3/8" divided by 12), equals 293 foot pounds of work. Our second example, because it has twice the chamber volume above the piston location, must move twice as far to lower the cylinder pressure by 1/2. Since all the other numbers, by our own definition are the same, the force multiplied by a distance twice that of the first example will equal twice the work done, 586 foot pounds of work. There is no free lunch in horsepower equations because to get 1000 PSI above the piston in the second example takes twice as much fuel and energy as the 1000 PSI in the first example. What this offsetting of the peak pressure does is allow us to use the extra fuel mix available to a nitrous engine without breaking and melting things. The system that allows us to postpone maximum cylinder pressure is ignition timing retard. To a lessor extent short rod ratios, lower compression ratios, high RPM, aluminum heads, a tight quench, a rich fuel mixture, a small carburetor and hotter cams tend to delay maximum cylinder pressure.

Understand that, in our quest to delay cylinder pressure's peak time, more is not necessarily better. Instead, consider that the ideal cylinder pressure would be just short of detonation pressure and this pressure would be maintained from top dead center, and as long as possible after TDC. If timing is really late, you won't build enough cylinder pressure to start the car, let alone drive it. The 1000 PSI pressure in the example is not the maximum allowable combustion pressure but, rather, a comfortable pressure for illustration of the work principle.

Some nitrous manufacturers recommend, "retard the timing two degrees for each fifty horse power of nitrous". Other nitrous kits have the flame speed artificially slowed by the intentional use of a rich fuel to nitrous ratio. The maximum performance engine with a heavy nitrous load must achieve peak cylinder pressures, with the combustion chamber size being drastically increased due to the piston being on its way toward bottom dead center. The strongest engines have less compression ratio, less spark advance, and more nitrous.

Many people just don't like the idea of any retard. They say, "retard timing and exhaust heat goes up". It usually does in a stock non-nitrous engine because lower peak cylinder pressure slows the burning. If the timing is retarded in a non-nitrous engine, the exhaust opens before the fuel mix is finished burning and exhaust temperatures go up. Piston temperatures usually go down and exhaust valve temperature goes up. In the nitrous engine, exhaust temperature goes up for several reasons. The first is that the power output has gone up considerably. More power usually produces more waste heat. Second, the need to keep maximum cylinder pressures within reason has dictated that the biggest part of the fire happens closer to the exhaust valve opening time. There just isn't enough piston travel to extract all the energy out of the charge before the exhaust valve opens. Now, we could and sometimes do, open the exhaust valve later so more combustion pressure energy can be used to turn the crank. The trade off is negative torque on the exhaust stroke. If we still have significant cylinder pressure in the cylinder as the piston moves from BDC to TDC on the exhaust stroke, your net HP falls drastically. A real problem at higher RPM.

You can improve maximum power stroke efficiency and minimize exhaust pumping losses by running the engine at lower RPM and/or improving the exhaust valve size, lift and port design. A big nitrous engine likes everything about the exhaust to be big. If it flows good enough the cylinder will blow down by bottom dead center, even at high RPM with relatively mild exhaust valve timing. There are many variables in the design and development of an all out nitrous engine. A mistake will cause the melt down of any piston. The high strength of the hypereutectic piston will withstand detonation and severe abuse. Unfortunately, all pistons, even forged will melt and when cylinder pressure limits are exceeded, run away detonation can occur. The excess detonation heat makes the plugs, valves and pistons so hot the ignition system alone cannot be used to shut the engine down. Continued operation worsens the situation to the point of a total melt down. Designing a maximum performance nitrous engine is more of an exercise in heat management than it is in engine building. Serious nitrous users should review our list of ceramic coatings.

A lack of a sufficient fuel supply is probably the most common killer of the nitrous engine. If you add a 300 HP kit to your present 300 HP engine, your fuel requirements roughly double and a shortage doesn't just slow you down, it melts things. An electric fuel pump and fuel line devoted entirely to the nitrous equipment is recommended. If you are using a diaphragm mechanical pump to supply fuel to the carburetor, it is worth while to increase the fuel line I.D. If the carburetor goes lean while the nitrous is on, the pistons can melt even with a rich fuel line trick (1/2" dia.) only makes a major improvement in the operation of diaphragm mechanical pump is not recommended and does not do well at high engine RPM. A large size line is effective with a mechanical pump, even if you use smaller fittings at the tank, fuel pump and carburetor. The advantage of the 1/2" large line is not related to the steady state flow rate of the line.

The advantage relates to the acceleration time and displacement of the pulsating flow common to the mechanical pump.

High compression ratios can be used with nitrous but shifting the maximum pressure after top dead center becomes more and more difficult. I prefer to use street compression ratios and then just work with adding more nitrous to get desired horsepower levels.

We are currently testing some pistons specifically designed for Nitrous use. Current "off the shelf" pistons have been successfully run with a 500 HP nitrous kit combined with a nitrous control system. Most of our effort has been to develop new ideas that will push the limit of nitrous technology. More testing is planned with a piston especially coated to reduce detonation.

When choosing piston rings for an engine the most important factor is the intended use of the vehicle. A piston ring set that delivers excellent street performance may not be correct for an engine that will see competitive action, or for one that will be used with nitrous oxide.

Piston rings serve two purposes - to contain the cylinder pressure, and to prevent oil from getting into the combustion chamber. Sealing against pressure leakage, or "blow by", is the responsibility of the top ring. The keys to good ring sealing are cylinder wall finish and piston ring groove condition. If pressure gets past the top ring it is already "lost". Any such leakage will not be ignited by the spark plug, and is unlikely to produce any significant power, even if captured between the first and second ring. The second ring is primarily an oil control device. If the top ring is doing the job, the second ring will see fairly limited combustion pressure. Some companies sell second rings that use complex or fragile designs for sealing. These are prone to premature wear and have unpredictable behavior at high RPM levels. Cylinder leakage test percentages are only useful for comparison to data generated when an engine was fresh. Unfortunately this kind of information can be misrepresented to show very low leakage numbers by folks trying to sell "trick" parts. Leakage tests are steady state - they do not account for time, piston movement, or true operating pressures. "Blow-by" measurement will give a better picture of ring condition, but on track performance is the only real measurement of success. Our moly rings are intended for applications where cost is of prime importance.

Engines being built for serious competition will be far better off using Plasma Moly ring sets. These feature an extremely durable ductile iron top ring with Plasma Moly facing. This design allows the ring to seat quickly and to maintain its sealing integrity under the severe stress of racing. The second ring is a special low tension plain iron design. These taper faced rings are specifically designed to break in quickly and to keep oil from migrating into the combustion chamber. The SS50U stainless steel oil control rings are the absolute best in the high performance industry. This ring combustion gives dependable sealing and allows maximum power production.


Piston ring sets are available with either standard or low tension oil rings. The standard tension rings are recommended for street driven applications, and for race vehicles which may see frequent open to closed throttle transitions in use - such as road racing. They are also useful in engines that may experience cylinder bore distortion during operation.

Low tension oil rings deliver increased performance due to their reduction in cylinder wall drag. These are highly recommended for many racing applications. Engines using low tension rings should be built with particular attention to cylinder concentricity, and often benefit from the use of a crankcase vacuum system.


The piston ring's end gap can have a significant effect on an engine's horsepower output. Rings are available both in standard gap sets, and in special "file fit" sets. The file fit sets allows the engine builder to tailor the ring end gaps to each individual cylinder. Ring gaps should be set differently dependent upon the vehicles use, within the range of .003" (for the 2nd. ring) to .004" (for the top ring) per inch of cylinder diameter. The more severe the use, the greater the required end gap (assuming the use of similar fuels and induction systems). Engines having low operating temperatures, such as those in marine applications is too small. The chart below is a general guideline for cylinders with a 4.00" bore, adjust the figures to match your engine's cylinder diameter:

Top Rings (ductile iron, 4" bore)


Nitromethane .022 - .024"

Alcohol .018 - .020"

Gasoline .022 - .024"

Normally Aspirated - Gasoline

Street, Moderate Performance .016 - .018"

Drag Racing, Oval Track .018 - .020"

Nitrous Oxide - Street .024 - .026"

Nitrous Oxide - Drag .032 - .034"

2nd Rings (plain iron, 4" bore)


Nitromethane .014 - .016"

Alcohol .012 - .014"

Gasoline .012 - .014"

Normally Aspirated - Gasoline

Street, Moderate Performance .010 - .012"

Oval Track .012 - .014"

Pro Stock, Comp. .012 - .014"

Nitrous Oxide - Street .018 - .020"

Nitrous Oxide - Drag .024 - .026"



When installing new rings, the single greatest concern is the cylinder wall condition and finish. If the cylinders are not properly prepared, the rings will not be able to perform as designed. The use of a torque plate, head gasket, and corresponding bolts are necessary to simulate the stress that the cylinder head will put on the block. Main bearing caps should also be torqued in place. The correct procedure has three steps. First the cylinder is bored to approximately .003" less than the desired final size. Next it is rough honed within .0005" of the final diameter. Then a finer finish hone is used to produced the desired "plateau" wall texture. Use a 280 - 400 grit stone to finish cylinder walls for Plasma Moly rings.

Note - the "grit" number we are referring to is a measurement of roughness, it is not the manufacturers stone part number (a Sunnen CK-10 automatic hone stone set #JHU-820 is 400 grit). The cylinder bores should be thoroughly scrubbed with soap and hot water and then oiled before piston and ring installation.

Piston ring grooves are also sealing surfaces, and must be clean, smooth and free of defects. Ring side clearance, measured between the ring and the top of the groove, should be between, .001" and .004".
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I'm interested in actually calculating the Squish Velocity of a few different combo engines to see how much difference an increase in stroke vs. quench distance vs. piston type (flat top vs d-dish vs. traditional dish), BUT I am having a TERRIBLE TIME finding any information on what is a reasonable "Squish Area %" to use a flat top SBC 350 or 355 vs. the same combo with a D-dish piston.

To just be able to understand the relationship and how significantly each factor changes the squish velocity, I don't need to go to crazy lengths to get a perfectly accurate value, I just need "ball park" numbers and I can't seem to find them anywhere.

Anyone ever do this before and have reasonable numbers they can share?
BUT I am having a TERRIBLE TIME finding any information on what is a reasonable "Squish Area %" to use a flat top SBC 350 or 355 vs. the same combo with a D-dish piston.
Is it because you don't have a piston and head to measure the surface area of the quench surface?
Is it because you don't have a piston and head to measure the surface area of the quench surface?
Partly given that my engine is assembled now, but also because what I'm looking to understand is how much changing different parameters changes the quench velocity and I definitely don't have a giant stack of different heads and pistons in front of me to just measure and find out.

I'm looking to use a Quench Calculator to model things like a 350 SBC with different piston-to-head clearances, changes to stroke, changes to the bore, rod length changes, flat top pistons vs. full dish vs. dish with perimeter quench band, vs. D-Dish / Inverse Dome pistons -to just see what the range of quench velocities is given these very typical changes.

I am pretty sure I understand the DIRECTION of each of those changes, but not the relative AMOUNT that changing each one impacts the quench velocity.

Play with even made-up data, it's obvious that the head gasket changes things very quickly, stroke not as much as I expected, but without decent %Squish Area estimates, it's hard to do things like compare the quench velocity between a 396 and a 406 SBC with roughly the same static compression. (I'd guess that the smaller bore and longer stroke of the 396 has a better quench velocity vs. a 406 with a piston that would match the CR, but I'm not certain.) --Do any of these combos end up with a quench velocity that's getting TOO HIGH and actually makes detonation MORE likely?

-IDK, but I'd love to use a quench velocity calculator to find out. Getting good Squish Area % data is my current "blocker".

Partly given that my engine is assembled now, but also because what I'm looking to understand is how much changing different parameters changes the quench velocity and I definitely don't have a giant stack of different heads and pistons in front of me to just measure and find out.
But if you are limiting this discussion to the 350 SBC, that simplifies the subject considerably. Most heads have a similar quench surface .... right?

With good top view photo of the piston, you can calculate the quench surface area. Same with the heads if needed. You don't actually need the components, just good photos. You would need at least one know dimension for each photo. For the heads, that could be valve size, for pistons that would be bore size.

I am pretty sure I understand the DIRECTION of each of those changes, but not the relative AMOUNT that changing each one impacts the quench velocity.
Do you have a formula(s) to use ?

-IDK, but I'd love to use a quench velocity calculator to find out. Getting good Squish Area % data is my current "blocker".
I did a quick search and it seems there are website calculators out there, but it also looks like a possible candidate for an Excel spreadsheet.

Do any of these combos end up with a quench velocity that's getting TOO HIGH and actually makes detonation MORE likely?
Well that's a new one on me, I've never read that quench could produce too high of a velocity or mixing.

Intriguing subject, how did you come up this !!!