unshrouding valves, and polishing combustion chambers


Staff member
some excellent info well worth reading thru
head gaskets are rarely completely round, nore are combustion chambers
you,ll want to place a head gasket you,ll use on the heads and mark the area inside the opening as the only areas you can change,
(notice the gasket fire ring is NOT a perfect circle like many people assume)
ideally you,ll want to un-shroud the valves while opening up the combustion chamber volume, but not extend the combustion chamber past the front edge of the gasket fire ring as that usually causes gasket failure














you might have somehow gotten the IN-CORRECT idea that the valve center-line and CYLINDER center-line are exactly matched in a sbc, they are NOT!
plus the valves are angled at a 23 degree angle so the outer valve edges close to the cylinder walls do not drop strait down into the bore as the valves open at the point of the cylinders largest diam. moving the head center-line to move the valve center- line to maximize flow at max valve lift potentially helps performance.



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
the most common mistake made by many people is that they fail to look at an engine as an interconnected group of component sub systems and they don,t realize that changes to a single component, no mater how much potential that component has is not going to allow that component or change in the potential to be realized until all the matched and supporting systems have similar potential.
the heads may be capable of flowing (x) on a stock engine but with careful selection of a cam with the correct duration and lift, and with a tuned header, and matching valve train mods along with some port and bowl clean-up the resulting improvements can be significantly more impressive.
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,

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.










it should be rather obvious that youll need to know the exact distance the piston deck sits at TDC ,above or below the block deck surface and the valve notch recess or pop-up dome volume of the piston to do accurate quench or compression calculations



keep in mind any valve clearance recessed areas must have the areas shrouding flow blended to increase rather than restrict air flow and to reduce the potential for detonation that sharp exposed edges tend to have









notice how the valve seat supporting casting in the cylinder head, throat extends out into the port and restricts the valve flow, a critical area that port and bowl clean -up can usually gain significant flow improvements

horsepower is a mathematical expression of how effectively you can use TORQUE or the rotational force an engine produces
torque x rpm/5252=HP
torque is the basic result of, the efficient use of,
your engines DISPLACEMENT and COMPRESSION, ratio and the fuels octane limitations, as your effective pressure above the piston is the result of both the volume of fuel/air mix being burnt and how efficiently that burnt fuel, can produce cylinder pressure.
changing the cam timing and duration will change both the rpm range and that cylinder fill efficiency at any specific rpm.
the cam timing obviously effects both the cylinder filling efficiency and port flow and obviously increase lift and longer duration allows more potential flow, but a bit less volume trapped above the moving piston, as it can,t compress anything until both valves seat







you can,t guess or assume, a stock push rod length will work, especially if you used an aftermarket, cam, lifters or rockers or changed cylinder heads ,you,ve got to measure correctly and get the correct length, failure to take the time and effort required usually results in valve train durability issues or engine damage, yeah you can ignore the info provided, but taking the time to do it correctly will save you a whole lot of potential problems & grief later on.

http://www.circletrack.com/enginetech/1 ... ch_engine/

http://www.gofastnews.com/board/technic ... uding.html

http://www.hotrod.com/how-to/engine/061 ... d-porting/


http://www.eurospares.com/graphics/engi ... 0Time!.pdf

http://www.gofastnews.com/showthread.ph ... rt-Volumes




http://users.erols.com/srweiss/tablehdc ... _Big_Block

http://www.j-performance.com/index.php? ... view&id=48






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


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



restrictive, as cast

less restrictive after unshrouding the valve


notice the tight bath tube shape and shrouded valves but decent quench areas in the fuelie head above
smog era heads like the #416 or #624 that came on the 1984 corvettes were less closed , because of the laid back chamber wall but they were designed for lower compression low octane fuels not performance



Potential HP based on Airflow (Hot Rod, Jun '99, p74):
Airflow at 28" of water x 0.257 x number of cylinders = potential HP
or required airflow based on HP:
HP / 0.257 / cylinders = required airflow

Every factor has some effect on the total combo, polish the combustion chamber and round all sharp edges and get the quench correct and use aluminum cylinder heads and you can run a bit more compression than stock iron heads , run a richer FUEL/AIR mix and a bit slower ignition advance curve and you can run the same total ignition advance on the same engine with out getting into detonation nearly as easily, run a 4.11:1 rear gear and a larger aluminum radiator with a 180F t-stat and theres obviously other factors, heat range on the spark plugs, and you can use a system that sprays E85 only under higher loads if you want to run a separate fuel tank , etc. so theres no 100% set compression to octane rules.
but generally the higher the combustion chamber temps, the higher the air temps, the higher the engine coolant temps and oil temps and the higher the compression,the faster the ignition advance, the MORE likely youll experience detonation under high engine loads


A HARD number that has held pretty true for conventional BBC on gasoline , with compression ratios up ner optimum, near 12:1-13.5:1 to predict peak HP from head flow is .25-.27 x intake flow rate @ 28" x 8 (# of cyls). Like others have said, a lot of variables,like efficiency of exhaust scavenging,compression ratio and valve lift VS port potential flow, but it has been within 20 or 30 HP on several different BBC's I've seen being dynoed.

For example, I had an 357 AFR-headed 540, the heads flow 425cfm @ 28". So 425 x .27 x 8 = 918 HP. It made 940. Another motor calc'd at 1030, made 1040 on the dyno.

Using the same logic, 50 cfm x .27 x 8 = 108 hp. It's not that simple, it depends on combination, how optimized everything is etc. But if you are looking for round numbers, 50-75 hp is probably realistic, 100 hp possible

ITS A COMMON MISCONCEPTION,THAT YOU MEASURE PORT CROSS SECTION AT THE PORT ENTRANCE,BUT ITS NOT the port area at the entrance , you need to use in the calcs, ITS the MINIMAL port cross section at the SMALLEST point in the port, usually near the push rod area.
LIKE a funnel, its not the largest part of the opening but the smallest thats the restriction to flow
heres some old fuelie heads with the chamber slightly un-shrouding the intake valve pocket walls
polishing the combustion chambers and smoothing contours tends to reduce detonation and improve power, combine that with port and bowl area clean-up and careful blending of the port walls. and a back cut on valves with a multi angle valve job, etc. and its not unusually to gain 25-40 hp or more, from port work and combustion chamber mods that improve air flow rates









obviously a bit of port, runner and bowl clean-up port work , gasket matching and valve pocket clean-up along with a well machined 3 or 5 angle valve job ,would go a long way to potentially improving the total flow numbers



yes cc ing the chambers and getting the quench right helps a great deal






Last edited by a moderator:
Porting School #6 - Secrets to reduce valve shrouding

#6 Port Appraisal and Valve Shrouding

In part #5 we identified a primary flow restriction- now is time to get into more detail.
David Vizard

Porting School #5 established that the valve and valve seat is a prime restriction point - now let us take a closer look at what we are dealing with here. In part #5 I dissected the port to make a point. I believe the point that the valve seat, not the port itself, caused potentially the biggest restriction to flow’ was satisfactorily demonstrated. However much that analogy may simplify things, it does in fact contradict one fundamental truth about engines in general and cylinder heads in particular. That fundamental truth is that in the process of developing a high performance engine almost no part of it can be considered in isolation. Although (in terms of air flow) the interdependence of one section of the head on another in the drawing below varies we cannot take any one part of it and truly consider it in isolation.

This section was shown in Porting School #5 but I want to bring your attention back to it here. What we are going to do is both progressively put the head back together and lift the valve further.

But before we consider the implications here let us establish a couple of geometric factors concerning a valve and the hole it occupies. To make life easier check out the drawing below.

What you see in the above drawing is a valve lifted to the point where the open curtain area surrounding the valve is equal to the valve diameter. What this means is that if the valve seat was zero width, opening the valve any further than 0.25 of it’s diameter will not present any more breathing area to the cylinder. However because the seat occupies space under the valve we find in practice that the curtain area reaches the effective area under the seat and minus the valve stem at about 22.5% of the valve diameter. So for our example 2 inch valve lifting it more than 0.45 inches won’t actually deliver any more breathing area.

So does this mean that valve lifts above .225 D are a waste of time as they won’t deliver any more flow? No – because of losses and the areas surrounding the valve, lift values, especially for a parallel valve engine, can show increasing flow up to as much as 0.35 D. Although the 0.35 D may well represent a limit some special things start to happen at about the 0.25 D – especially in a typical two valve design of cylinder head so remember that relationship for later.

Re-assembling the Port.

OK now we have that bit of nomenclature concerning the lift to diameter ratio covered let’s go back and look at section ‘C’ of the above cylinder head. If we consider this section in isolation and the valve lifted to say 0.35 D or more we will find that, because of the uninterrupted straight approach, this section will flow a lot of air. In the case of our model, that’s about 275 cfm. However if we add section ‘B’ to ‘C’ then things change dramatically. The air now has to make a turning approach to the valve seat area.

Look at the throat section on the left. If we take the valve and lift it from about a mid lift point to something equal to say 35% of it’s diameter (0.35D) we can expect a big increase in flow. If the valve is lifted even further to say 100% of it’s diameter it has become sufficiently far removed from the seat to be about out of the sphere of influence of the seat. But this situation makes some assumptions. The first is that the air approaches the back of the valve from a near optimal direction and secondly that there are no further impediments to flow once the air has passed the seat. Unfortunately these circumstances might not represent the real world. If we re-assemble our sectioned port with the part that houses the guide we find that the air is now constrained to approach the seat area in an arc. At higher lift figures the air on the floor of the port, that’s the short side turn, is going too fast top make it around the turn. This results in skipping across the back of the valve and attempting to go out on the top (the long side) side of the port. This means the short side and the valve circumference in the immediate area is under-utilized. But our problems don’t stop there. Check out the red arrow labeled ‘Problem!’ This situation is called valve shrouding. Only by lifting the valve to extreme lift values can the head shown here avoid the restriction cause by this shrouding. For a clearer picture of what valve shrouding is check out the next drawing.

With the arcing approach to the back of the valve the flow situation changes dramatically. Now we are into port design rather than seat design. Although we can never truly separate these two aspects of the induction tract our focus will turn more toward the optimization of the shape of the port. But the approach to the back of the valve is not our only concern if we are dealing with a 2 valve per cylinder, parallel valve head, having either a bath tub or wedge chamber shape. We also have to deal with valve shrouding. Instead of using 1000 words here I am going to refer you to the drawing below and limit my words to 900 or so!

Imagine that for some reason you have a valve, seat and port at the bottom of a cylinder as shown in the left hand drawing. Because the walls of the cylinder are snug to the circumference of the valve it matters little how high the valve is lifted as no air will flow around the edge no matter what. Now let’s pull one side of the cylinder wall away from the valves circumference as per the drawing on the right. With the wall this far from the edge of the valve there is a clear path for the air to flow around that side of the valve.

From the preceding drawing it can be seen that if either the combustion chamber or the cylinder wall is too close to the edge of the valve flow will be impeded – and all this brings us to the $64,000 question - how close is too close??? If you supposed that there might not be a simple answer to this question go to the top of the class. There are two ways we can consider shrouding of a valve. The first is just by consideration of the through-flow areas involved as the air attempts to navigate it’s way around the valve. The second is shrouding relief that is dictated solely by the degree to which that part of the valve is utilized due to the amount of air flow at that part of the circumference. That may sound a mouthful but I will expand on that in a moment. For now let’s consider the first situation – the through-flow area’s.

Understand that the design of the ports and chambers of a cylinder head for optimal flow/performance does not lend itself to overly simple mathematics. However a little relatively simple math applied here and we can come up with dimensions that, for want of a better description, produce a shrouding factor that is geometrically minimized. When applied to the fullest extent it produces a design that cannot be geometrically de-shrouded any further without mechanical compromise to the system. An example here is valve shrouding caused by the cylinder bore. If this is cut away then the valve may be un-shrouded further but the rings won’t seal up! What follows is an example of a real world situation.

What you see here, drawn to scale, is a typical head design for a small block Chevy. It is of a wedge configuration with the deep side on the plug side of the chamber. For the valves to be fully geometrically un-shrouded at 0.25D lift there must be a completely clear area from the edge of the valve right out to the black circle surrounding each valve. Anything within that circle is, to some degree, reducing the area of the flow path around the valve. The chamber shown is a stock Chevy design as per a 186, 049 etc high perf casting of the late 60’s through to about the mid 70’s. This chamber design was, for the most part, the starting point for just about every after market small block Chevy high performance head. In this drawing you can see that any part of the chamber colored green is, at 0.25 D, shrouding the valve. Assuming the casting thickness is there these area’s can be cut away. The shrouding caused by the bore is something we are stuck with when using this style of head with parallel valves.

To understand the next part of our discussion you need to appreciate that air does not flow in a neat and orderly fashion within the port. This is so far removed from reality and the only way you will really appreciate how far out that may be is to spend time on a flow bench and see for yourself. The drawing below shows how the bulk of the air enters the cylinder on a small block Chevy head.

When the valve is lifted to about the 0.450-0.500 range the air, although with streamlines far less tidy than shown here, enters the cylinder about like this. The bulk of the air passes out of the port and into the cylinder through the half of the valve labeled ‘A’.

Zeroing Out Geometric Shrouding.

When addressing valve shrouding with the intent of minimizing it we need to make a start somewhere and ascertaining what the form of a chamber may be, if it was geometrically un-shrouded, is as good a place to start as any.

The breathing area presented to the chamber by a valve moving through it’s lift envelop is not quite as simple a geometry problem as it may first appear. The reality is that as the valve lifts it moves through three distinct regimes, each of which requires it’s own particular set of math formulas to produce an answer as to what the through-flow area is. We are not going to deal with this now as it is more advanced stuff. However, even if we ignore that we can still come up with a very good approximation of what it takes in the way of chamber form to produce a geometrically un-shrouded chamber. What we find is that at low lift the angle of the chamber wall as it leaves the valve seat needs to be very close to 45 degrees and as the lift progresses up to the critical 0.25 D lift point the angle needs to increase to about 52 degrees from horizontal.

The drawing below gives us a good guide to the form that needs to exist around a valve as it progresses through it’s lift envelope to ensure that the flow area around it is at least equal to the effective curtain area beneath the valve head.

Look closely at this drawing. The green line represents the angle of the chamber wall as it comes off the seat. For all practical purposes this is right around 45 degrees. As the valve lift progress the point of zero shrouding of the edge of the valve in relation to the chamber wall gets slightly steeper until at 0.25D the wall angle is close to 38 degrees off the vertical (52 from horizontal)as represented by the blue line. Although not totally accurate we can say, within close limits, that when the valve is at 0.25D lift the gap between it and any possible obstruction should be equal to a minimum of 0.20D. Above 0.25 D valve lift the chamber wall can be vertical for zero geometric shrouding as the valve has reached the limit of the area it will present to the cylinder.

Minimizing Real World Shrouding.

It is entirely possible to eliminate valve shrouding at the design stage of a cylinder head. The classic Hemi layout shown above does just that. But such a design is not always practical for reasons of cost or vehicle installation. The hemi head design configuration and to a lesser extent a four valve per cylinder type head have minimal to zero valve shrouding. The drawing below shows why.

As I have previously pointed out minimizing geometric valve shrouding is only a starting point. For us to apply whatever it may show us we need to ask ourselves two questions. The first is: do we want or need to have an obstruction to flow that is at least as un-streamlined as a valve seat. In other words do we want to present the air with a second bottle neck about equal in flow capability to the gap between the seat on the valve and it’s counterpart in the head. Secondly we must ask if the part of the valve we are considering the shrouding factor of is over or under utilized as far as the airflow is concerned. Here I refer you back to the drawing showing that on a small block Chevy (and the same will apply to any head of similar configuration) most of the air goes out the long side of the port. From this it is obvious that we will want to unshroud this side of the valve more than the other for optimal flow.

If we look at what we have learned so far and apply it to a real world cylinder head we can begin to see some porting development value of drawing a zero shrouding circle around both the intake and exhaust valves. The illustration below does just that – study it and read and absorb the caption .

Take a look at the blue circle first and note it’s relationship with the cylinder bore and the combustion chamber. The most noticeable aspect here is that at 0.25D lift there is a substantial amount of valve shrouding caused by the bore as represented by the area of the blue circle overlapping the cylinder bore circle. There is little, in a parallel valve design of head, that can be done about this although it’s negative impact can be partially offset by utilizing a very high valve lift. The second point of note is that the chamber wall on the plug side has been cut away far more than whatever amount it would need for zero geometric shrouding. The reason for that is because this side of the valve is a high flow area and is utilized far more than the side opposite.

Turning our attention to the exhaust valve we can see that at the critical 0.25D lift value there is virtually no shrouding of the exhaust valve. The spark plug side of the chamber is cut away far more than is needed just to achieve zero geometric shrouding. Why is this? Here we see the chamber form having to satisfy the needs of the combustion process so the shrouding, since it is none existent in that area, becomes academic.

However there is one aspect that does show up here that had we not drawn the zero shrouding circles around the valves would have been much harder to appreciate. With any cylinder head the most important valve is the intake as (in the simplest of terms) it is a lot harder to suck a full charge in than to blow or push one out. What we see here is an intake with appreciable shrouding and an exhaust with virtually zero shrouding. This immediately tells us that both the intake and exhaust valves should be re-positioned so that the intake valve moves away form the bore wall and the exhaust towards the bore wall on it’s side of the cylinder.
So why did this head manufacture miss the mark on the relative position of the valve centers in relation to the bore? Well really they did not. The intent with this head was that it would use the same valve centers as a production small block Chevy so that it would fit all the piston valve notches as used for off-the-shelf pistons and things such as pushrod guide plate centers would remain stock. In other words everything that fit a stock head would also fit this one.


I will guarantee that this is a more comprehensive lesson on valve shrouding than you could have got any where else – but are we finished on this subject yet? No – and rest assured the next level will be more advanced and reveal factors little appreciated by even most pro head porters. For now let’s consider how we can apply what we have learned so far to the development of a parallel 2 valve style cylinder head. Here’s the scenario: you have your flow bench set-up and have flow tested the head and are now amazed at how bad it is (makes you wonder what these factory engineers did with their benches eh?). What now? Take a look at the head and as far as possible identify what are likely to be the ‘busy’ areas around the valves. Next, using a valve with a center-pop in the middle scribe out, with a set of dividers, a circle around each valve that is 0.2 D out from the edge of the valve. (that’s a diameter equal to 1.4 times the valve diameter). The line you are left with across the head face is a good starting point for de-shrouding the valve.
Below is an example I worked on an 850 BL ‘A’ series engine back in the 70’s to demonstrate this point. The flow numbers show the effect of de-shrouding and how eventually a limit is reached.

The valve head in this example is 1.093 inches diameter. The green circle in the drawing represents the radius around the valve with the head in stock form. This is approximately 0.68 inches. The black curve in the graph is the flow delivered with the stock chamber form (shown by the black line in the drawing). The blue line is, at 0.80 radius, near the zero shrouding limit. When the valve was de-shrouded to this line the flow increased as per the blue line on the graph. As you can see there was a substantial increase in flow. Increasing the de-shrouding from 0.2 D to 0.3D as per the purple line increased flow further but going out to 0.4D netted only minor gains at high lift values. Remember – this is a small valve and 0.400 represents a lift to diameter ratio of 0.366. On a small block Chevy that would be equivalent to 0.740 lift.

The ‘busy’ area for this port is the lower side (port approaches from the top) of the chamber in this drawing. This particular head responds to de-shrouding because there is a reasonable straight run of port prior to the valve so and this, in conjunction with a small valve allows the short side turn to wok relatively well.

Ok – we have made a start on our investigation on ports and valve shrouding but, as important as this may be what we have so far considered is but a tip of the iceberg. Not only must we look at the ports in terms of shape but also size. All too often the Stroker McGurk syndrome eeks its way into the racers mentality. Although a big port may look like the way to go and may indeed net more flow than a smaller one it is, more often than not, a case of ‘if some is good more must be better and to much just right’ thinking. Well the news here is that it just doesn’t work that way. In our next Porting School feature we will look at tests of four different size intake ports on a 383 small block Chevy – after you have seen the dyno tests you may well have a little more regard for the port volume (size) why it is rated that way for a typical US V8 and what you may want to choose for your next build.

David Vizard
he chart below,is supposed to point out LSA, but its mis-labled



which can,t be changed once the cam is manufactured
Which can be adjusted by advancing or retarding the cams index to the crank rotation, as desired
with bushings or an adjustable timing set.







notice its right where the roller cams lobe design maximized the extra air flow potential that is the most effective flow area during the whole valve flow curve
and yes it frequently helps to match a roller cam to roller rockers as the reduced friction further helps the engines durability and ability to easily cope with faster valve train component acceleration, that tends to reduce heat and wear.
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, thus it makes a great deal of sense to push the valve lift a bit over .500, and have an intake port that is at least 3.2 square inches in cross sectional area, if you want to maximize flow on a 2.02" intake valve




keep in mind that the piston moves a great deal slower per degree of rotation
near the top and botton of the stroke than it does near the mid stroke in the bore, so thers more time for flow and presure to build per degree of crank rotation,making those areas of rotation more critical to performance


keep in mind the goal here is to increase or decrease the overlap , that occures as that has a major effect on the efficiency of the headers ability to efficiently scavenge the cylinders while both valves are simultaneously open




Last edited by a moderator:
5 Golden Rules to Goof-Proof Porting

#9 Follow the five rules discussed here and you will be sure not to fall foul to a power breaking move.
David Vizard

The title of this, part #9 of our Porting School series, is self explanatory – but why now? Why was this one not part #1? I gave this much thought when I started this series and came to the conclusion that it would be best to get, to an extent, immersed into our subject so to speak first. By introducing examples early on I felt that any ‘general rules’ that may be made from there on out would have more significance. For instance I won’t need to explain the importance of getting the port size right – you will have already seen how that effects things as shown in PS#7. Really what I am going to do here is take a breather and sum up the implications of what has been covered so far – so here goes.

As obvious as Rule #1 seems the big problem for the novice is almost always a question of recognizing exactly where the greatest restriction in the induction/exhaust tract is. Primary restriction points are dealt with in PS #5 and 6 so if you have not read these two features yet now would be a good time. One aspect that the novice porter will be pleased about is that tackling the most restrictive part of the system and freeing up some flow potential delivers the best power return for the time invested. For the record pocket porting heads is all about focusing on Rule #1 too the exclusion of almost all else. At the end of the day pocket porting may not produce the fastest looking set of heads or the most photogenic but the results can be very satisfying.

Any time you constrain the air to flow along a path that you are dictating the total flow will almost certainly drop. If where the air in a port is flowing is investigated it will be found that there are two distinct situations which determine it’s path. In the first situation we find that a substantial amount of air is flowing in a certain part of the port because the route along which it is flowing has minimal flow resistance. In the second situation we find that a lot of air is flowing at a certain point/area because of the shapes involved upstream, downstream or both of that high flow or ‘busy’ area. It is important to be able to recognize the difference between these to types of busy area’s. With the first situation there is a strong indication that the area involved needs to be enlarged to make room for more air to flow along what can be seen to be a flow efficient path. The roof of a typical port is a good example here. In the second situation we find that the fix for more airflow is to add material at and around the point of fastest flow. A prime example here is the very high speed flow that can occur on, or just in front of, the short side turn of a relatively low angle intake port (SB Chevy and Ford are prime examples). The trick here is to recognize one source of high speed flow from the other as they require totally opposite responses. So before I get a ton of questions here let me tell you this is a subject we will get into later.

Once a head porter or head designer appreciates just how heavy air is they tend to have a whole different prospective on the importance of port velocities and cross sectional area’s. The dyno tests covered in PS #7 are a good demonstration of the need to have the ports appropriately sized for the job. When we get to the stage of flow testing ports we find that not only is there a need to know how much air is flowing but there is equally a need to know where it is flowing and how fast it is going. All this comes under the heading of velocity probing and the cost of the equipment necessary to do that falls into the peanut category. We have looked at how to build a flow bench and down the road we will look at what it takes to make and calibrate a velocity probe for just a few bucks. And one last point before moving on – I had better give at least some explanation as to what redundant port area is. As the term ‘redundant’ suggests it is an area of the port where little flow is taking place. If this is the case it is redundant to requirements. The best action to take here is to fill it in. Redundancy in a port makes for a lazy port and that results in a less than optimal torque output every where in the rpm range.

A charge that has little motion not only burns slower but also less effectively. This is most noticeable at low engine speeds. Lack of adequate mixture motion can cut torque output at say 1000 to 2000 rpm by as much as 25%. When engine speeds are high (5-6000 rpm) the need for port/chamber induced mixture motion is far less. Mixture motion from quench action between the piston crown and the cylinder head face can be instrumental toward increased torque at all engine speeds. At part throttle lack of mixture motion can also have a direct negative impact on mileage. Another desirable engine characteristic to suffer when low mixture motion is involved is throttle response.

This is a big one here. The flow capability of a head absolutely cannot be judged by it’s reflectivity! Heads with a rough finish the right shape will always out-power heads with a shiny finish the wrong shape! This being so don’t be in too much of a hurry to start work with those 180 grit or finer emery rolls.

David Vizard

the fact that peak lift is not where you gain the vast majority of the ports airflow potential
think about the math a bit , at 6000 rpm, the valve is going from seat to full lift and back to its seat,3000 times a minute, thats 50 time's a second
but that 50 times a second includes the full 360 degree rotation, peak lift rarely lasts more than 40 degrees of that rotation, or about 1/10th of that time.so how much air flow can occur in 1 /500th of a second the valves are near peak lift








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


(1) open throat to 85%-90% of valve size
(2)cut a 4 angle seat with 45 degree angle .065-.075 wide where the valve seats and about .100 at 60 degrees below and a .030 wide 30 degree cut above and a 20 degree cut above that rolled and blended into the combustion chamber
(3)blend the spark plug boss slightly and lay back the combustion chamber walls near the valves
(4)narrow but dont shorten the valve guide
(5) open and straiten and blend the upper two port corner edges along the port roof
(6) gasket match to/with intake and raise the port roof slightly
(7) back cut valves at 30 degrees
(8) polish valve face and round outer edges slightly
(9)polish combustion chamber surface and blend edges slightly
(10) remove and smooth away all casting flash , keep the floor of the port slightly rough but the roof and walls smoothed but not polished.
(11) use a head gasket to see the max you can open the combustion chamber walls
(12) blend but don,t grind away the short side radius
notice how the valve seat supporting casting in the cylinder head, throat extends out into the port and restricts the valve flow, a critical area that port and bowl clean -up can usually gain significant flow improvements




a basic but effective valve job with blended port bowl area clean -up helps flow rates






aluminum cylinder heads tend to allow you to run about 1/4-to-1/2 point more effective compression, IE, if iron heads get into detonation at 10:1 ALUMINUM might ALLOW YOU TO RUN 10.3-10.4:1 BEFORE GETTING INTO DETONATION, BUT ON THE PLUS SIDE AT LEAST IN THEORY IRON HEADS AT ANY GIVEN CPR WILL HAVE A SLIGHT ADVANTAGE IN HP
but in my real world testing the difference is much closer almost non-existent
the main advantage I see in aluminum heads is lighter weight and their much easier to repair when damaged
an aluminum cylinder head allows heat transfer to the engine coolant at a significantly higher rate than an iron head, and generally about .25-.50 higher compression can be tolerated but theres No absolute real definite answer, depends on quench, cam LSA, cylinder head design, advance curve, fuel/air mixture, intake temperature, coolant temp, spark plug design, piston design, heat barrier coatings, combustion chamber surface texture, and a bunch of other parameters.

#2 Turbulence and Combustion Dynamics

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 wellâ€. 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 levelâ€. 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.
Last edited by a moderator:
just some related bits of info

A "burned valve" is a valve that has overheated and lost its ability to hold a leak-free seal. Valve burning is usually limited to exhaust valves because they run much hotter than intake valves.
its not common but a badly adjusted valve will not seat properly or allow exhaust heat to dissipate to the valve seat and can rapidly burn a valve

The diagnosis of a burned valve is usually the result of a compression test. If a cylinder shows little or no compression, it frequently means the exhaust valve is not sealing. The valve may or may not be actually burnt (melted), but have other physical damage such as cracks or areas where pieces of metal are missing or eroded away from the valve face.

The cure for this condition is to remove the cylinder head, replace the bad valve and reface (or replace) the valve seat. As a rule, the head is usually given a complete valve job at the same time because the rest of the valves and guides probably need attention, too. If one exhaust valve has failed, the rest are probably on the verge of failure if they haven't already started to leak.
Why Valves Burn

There are several reasons why valves burn. One is normal wear. As an engine accumulates miles, the constant pounding and thermal erosion wears away the metal on the face of the valve and seat. The exhaust valve sheds most of its heat through the seat, so when the face and seat become worn and the area of contact is reduced, the valve starts to run hot. Eventually the buildup of heat weakens the metal and pieces of it start to break or flake away. Once this happens, it forms a hot spot that accelerates the process all the more. The valve begins to leak and compression drops. The result is a weak or dead cylinder and a noticeable drop in engine power, smoothness and performance.

A bad exhaust valve will also increase exhaust emissions significantly because it allows unburned fuel to leak into the exhaust. High hydrocarbon (HC) emissions, therefore, may also be an indicator of a burned valve.

An exhaust valve can also burn if the valve lash closes up for some reason (improper lash adjustment, cam or lifter wear, a bent push rod, worn rocker arm or cam follower, etc.). The lack of lash (clearance) in the valvetrain prevents the valve from closing fully, which causes it to leak compression and overheat.

Valve burning can also be caused by any condition that makes the engine run hot or elevates combustion temperatures. This includes cooling problems, abnormal combustion like detonation or preignition, loss of exhaust gas recirculation (EGR), retarded ignition timing or lean fuel mixtures.
Valve Recession

A condition known as "valve recession" can allow the valves to recede or sink into the head because of excessive seat wear. This causes the valve lash to be lost which allows the valves to leak and burn. It occurs primarily in older engines (mostly those built prior to 1975) that were not designed to run on unleaded gasoline. When leaded gasoline was still around, lead acted like a lubricant to reduce valve seat wear. But when lead was eliminated, it meant engines had to be made with harder seats. These older engines didn't have hard seats, so many began to experience valve wear problems when switched to unleaded fuel. If you're driving an antique or classic car, therefore, you should either use some type of lead substitute fuel additive to protect the valves or have the seats replaced with hard seats when the engine is overhauled.
one of my friends is building a big block Chevy 496 and like most of us moneys rather tight so hes going to be using a set of used aluminum rectangular port heads that he bought rather cheaply because they needed rebuilding, new valve guides and valve seats inserts, and they came with no valves and one chamber damaged needing welding then the heads need to be surfaced, and port and bowl clean up work done..
I suggested he ask the machine shop he deals with what the cost difference will be to install .100 longer valves and use 2.30" intakes and 1.9" exhaust valves vs the 2.19" and 1.88 " valves that were previously used
you can purchase a 1/2" thick 6" x 6" lexan sheet you can drill and modify to cover a heads combustion chamber or a blocks bore to cc a combustion chamber volume or piston dome volume for under $16 on the internet/amazon /ebay




http://www.popularhotrodding.com/tech/1 ... ewall.html

the difference in cost was under $200 according to the machine shop so in my opinion its a no brainer as the larger valves and slightly longer length give him more options in valve springs installed height and the larger valve curtain area will easily be worth an extra 20-25 hp but at about 400-500rpm higher engine speeds.
minor upgrades and thinking ahead in parts selection when your on a strict budget can save you a good deal of cash and most guys don,t know about or bother to ask if there are options in the head assembly process.
BTW aluminum heads are easily repaired with a tig welder in most of the better head refurbishing shops,
the same damage done to iron heads results in an expensive and heavy door stop for the shop

Aluminum does have advantages, like light weight, and easy of machining compared to cast iron, example,cracks in valve seats on iron heads ",usually the result of overheating,"tend to result in coolant leaks that are not easily repaired, so you need a new cylinder head even if you had hundreds of dollars in port work done previously.
but on aluminum heads a bit of tig welding and machining for new valve seats repairs the heads rather easily



related useful info














Last edited by a moderator:
hey grumpy do you have the rest of those porting school series from david vizard? i would love to read the whole series

heres two of the few side-by-side closed chamber vs open chamber pictures I have, the open chamber heads are the newer design, and while the combustion chamber is larger and less shrouding of the valves should be obvious , you can not always swap heads with domed pistons as both the chamber wall location and valve clearance notches cam provide interference or clearance issues in some cases AND there obviously going to be a resulting loss of effective compression if your going from a 95cc-101cc closed chamber design to the larger 112cc-122cc open chamber heads.
Yes the open chamber heads do tend to flow a bit more air, the same port type (rectangle or oval port )AND if the open chamber heads have the same valve sizes as the closed chamber head that your comparing it too, but it can be a trade off resulting in marginal upper rpm gains for a noticeable off idle, rpm speed loss in low speed torque , when swapping to the larger chamber design heads, which is not always to your advantage in all applications so don,t be in a huge rush to upgrade.
it should be equally obvious that the dome piston design matching the larger open chamber heads will cause serious interference in most cases if you attempt to use the closed chamber heads.
yes you can get huge gains from the better aftermarket performance head designs , and yes theres a dozen ways to improve the OEM heads performance, but do your research carefully and don,t get blind sided, its rarely a good value to have OEM heads extensively reworked as the cost frequently exceeds the cost of far better performing aluminum performance heads.
THE question about buying bare heads and having a local shop you trust add the valve train components and do the machine work , VS buying ready to run, off-the-shelf cylinder heads, comes up frequently.
the manufacturers can buy in volume and purchase valve train components a good deal cheaper than you can, and they can have heads machined in large batch jobs that also reduce machine work costs, but they are also trying to be competitive so they rarely select the top quality components, thus there's always a trade-off and you,ll rarely see top quality machine work, and the better components used in off the shelf heads for sale....that does not indicate the off-the-shelf heads are junk, but the manufacturers must balance quality and price to remain competitive.
Youll almost always find that the larger and better known brand name suppliers like

have listed upgrades as options
yes quality varies and your going to be doing research or your not likely to get the best value per dollar or best potential performance.it would be foolish to not ask about the options cost and potential benefits of those upgrades.

related threads and sub linked info










ID suggest you select from heads from these sources
Jegs; 800/345-4545; Jegs.com

Summit Racing; 800/230-3030; SummitRacing.com

Scoggin-Dickey Parts Center; 800/456-0211; ScogginDickey.com


http://www.trickflow.com/egnsearch.asp? ... 4294867081
1-330-630-1555 • 1-888-841-6556


Dart Machinery; 248/362-1188; DartHeads.com

toll free: 877-892-8844
tel: 661-257-8124

Patriot Performance
Patriot Performance; 888/462-8276; Patriot-Performance.com


Toll Free: 877-776-4323
Local: 901-259-1134

http://www.edelbrock.com/automotive_new ... main.shtml
Edelbrock; 310/781-2222; Edelbrock.com

BMP (world products)
Tel: 631-737-0372
Fax: 631-737-0467





http://airflowresearch.com/articles/art ... /A-P1.html

http://airflowresearch.com/articles/art ... /A-P1.html

http://www.jegs.com/p/Brodix/Brodix-Big ... 9/10002/-1

http://www.onallcylinders.com/2012/05/2 ... -assembly/

http://www.onallcylinders.com/2012/09/0 ... s-top-end/

http://www.onallcylinders.com/2012/10/1 ... o-results/

http://www.hotrod.com/how-to/engine/ccr ... ock-build/

http://www.superchevy.com/how-to/engine ... ine-build/

http://www.superchevy.com/how-to/engine ... big-block/
Last edited by a moderator:
ok you, should do extensive research BEFORE you spend a dime, but Im assuming you have gotten a deal on cylinder heads from some place and they just arived!
pull out that set of cylinder heads and closely inspect them for obvious damage ,missing or busted parts, its not that uncommon for two cylinder heads you purchase to either have machining mistakes flaws or have the heads not match,ok, so you have a set of new cylinder heads and you start thinking about doing the required machine work, to get the max performance, first step if your smart is to contact the manufacturer and verify that the heads you have purchased fit the application you intend to build, no I'm not kidding I see guys order heads that won,t even fit occasionally!
your not going to get some large valve performance cylinder heads to even fit and function on some of the smaller bore heads.
some SBC heads won,t seal the head gasket on all blocks either.


now youll obviously want to verify the valve spring clearances, load rates and rocker

before you spend a good deal of money porting and un-shrouding any iron cylinder heads, keep in mind aluminum heads are easily repaired in a skilled and experienced automotive machine shop thats equipped to do those repairs but damaged iron cylinder heads are either much harder to repair or good door stops

geometry and theres certainly several threads on those factors, but polishing the combustion chamber, laying back chamber walls a bit to un-shroud the valve curtain and having a 30 degree back cut on the valves are ways to increase the flow at lower lifts, and a MULTI ANGLE valve seat with a rather shallow 300 degree seat vs the rather common 45 degree seat can significantly increase the low lift flow below about .300 lift, and remember the valves at peak lift once its at moderate lift TWICE so that flow rates quite important.










Id also point out your selection of valve springs can make a big difference, the newer BEEHIVE designs can occasionally allow use of lighter retainers and the fact that the BEEHIVE designs don,t generally reach a harmonic vibration that causes valve float or loss of valve seat pressure as easily for any given load rate tends to allow them to control the valves to several hundred rpms higher.



Last edited by a moderator:
David Vizzard in the past used a 30 degree Intake Valve Seat Main Angle I recall Grumpy.
Same as Pontiac V8 1955-79.
Except for the 1969 Ram Air 5 & SD45 They used a 45 degree intake seat.
Vizzard used 30 degree seats on his SBC builds. Shared secrets back around 1998 in a Drag Race Magazine .
He built a Super Mouse 454 sbc Gen 1.
I recall the Article still. Also used Ti valves Intake side.
After reading through....Its a bit optimistic porting away heavy and being successful .
Need a Flow Bench and Velocity Probe to check your work.
No one in our small group is so equipped.
I removed casting flash from my 1970 RAM AIR IV Heads in the past.
Never had to gasket match because the Fel Pro intake gaskets line up perfect.
Chambers are fully machined and valves unshrouded nice by Pontiac factory .
I only performed a multi angle valvejob.
Never ported.
Talking to Ron Iskenderian last week with the T/A race cam , he said my success in the past is because I never ported my 614 heads. Right cross sections as is by Pontiac.
thats interesting brian, i bet theres power to be had tho just with a sandpaper roll cleaning any casting flash there may be and creating an EVEN surface. 60 grit tootsie rolls and a sundy you could probably do both heads... this way theres no weird eddies in the air flow stream. on some aftermarket cast aluminum heads theres as much as 10 cfm there waiting for you with that safe simple work... i assume the same could probably be that much in your 614's also... its hard to overdo it with a sandpaper roll n cast iron i think youll be safe..

just somethin to think about
Super Rare Heads today Phil 614's.
I have one of the few sets never ported.
Value is around $5-7 k.
10 cfm possible gain not worth it.
True NHRA Super Stock Certified Heads.
Bone Stock.
On the Flowbench in the psst the 614's never WHISTLED.
Sign of Turbulance if heads whistle.

Just pure clean airflow .
Only Vacuum cleaner motors heard.
If you look at Grumpys flow area percentage improvements...Nearly all is from the valveseat and throat.
60 %.
Valvejob is the Secret to Airflow .
Velocity with airflow.

The Hellcats have more than 780 HP.

MY 614'S FOR 455 PONCHO.
col and yea youre correct theres more to power than cfm and ive never seen airflow speeds published anywhere
I think your comparing apples to oranges here!
the hellcats engine is SUPERCHARGED ,
the Pontiac engine and heads your referring too are not,
and I know you could match a proper set of lets say 335cc big block chevy cylinder heads to an 871 supercharger,
a vortec or paxton centrifugal supercharger or a couple turbos and get very impressive results. build a decent BBC and supercharge it,
and leave that hellcat looking at tail lights thru tire smoke!



Last edited by a moderator:
Are You Footing the Bill Grumpy ?
Its Race Time.
All have been bench racing.
The Warbirds are here. X 3.