Inliners International Forums General Discussion Engines headers on 261

I slept on my earlier reply and thought that one last post on the use of these Vizard CFM numbers was is in order:

His chart is based on flow-bench data and dyno testing.

The challenge is to estimate the CFM output of a particular engine so that the data contained in the chart can be put to use.

My personal take on that estimation is what I have offered here.

As I stated above – panic is spot-on about the arithmetic in converting cubic-inches to cubic-feet given a 4 cycle engine. So I went back and looked at my notes from when I was designing the exhaust for my early 50's Road Job. The use of 1728 as a constant is a sizing convention that also accounts for temperature – resulting in a handy formula using just displacement and RPM range. These models/conventions have been created over time to avoid what is actually a messy and cumbersome calculation when one tries to become more precise.

If an engine were simply an air pump (2-cycles) the amount of air entering the cylinder and the amount exiting would be equal - making the divisor for converting cubic inches to cubic feet 1728. But a gasoline engine is doing more than just pumping air – there is a compression and power stroke in between the intake and exhaust strokes – so only every other revolution is in play at a given RPM – making the correct divisor 3456. In addition, what is entering the engine (air + fuel) is not what is exiting the engine after compression & combustion (carbon dioxide + water). Moreover, the gaseous mixture is entering the engine at temperature A and exiting at temperature B - where the difference between B & A involves many variables.

When one tries to get rigorous in calculating a theoretical exhaust flow rate for a given displacement – one enters the world of combustion chemistry/efficiency and thermodynamics where temperatures are actually a gradient both on the way in and on the way out . . .

Here’s a good discussion on an engineering site ENG-TIPS that shows what happens when one attempts a more formal approach. Notice how quickly the debate splinters off in all directions depending the engineering focus: stoichiometry, volumetric efficiency, combustion efficiency, throttle position, effect of ignition timing on exhaust gas temperature (and therefore whether or not the engine is turning fuel into mechanical motion or just throwing heat – unproductively because of poor tuning). And then the thermodynamics guys chime in that exhaust gas is not an “Ideal Gas” and all of sudden what looks like some really cool applied-thermo formulas turns into many countervailing variables and caveats.

Which brings us full circle to the use of sizing conventions that are easy to compute and get us close enough to choose an exhaust pipe diameter and get on with building a system.

If you spend some time reading on this topic you will find several conventions in use. The guys from the bigger-is-better-camp immediately zero in on the effect of temperature deltas and peak EGT – for example the extreme 80-degree-intake/1800-degrees-exhaust(used in the post linked above) which yields an expansion factor greater than 4x - suggesting that 50 CFM of displaced flow becomes 200 CFM of exhaust flow due to heating!

But this maximum temperature exists for just a short time in a water cooled exhaust port. Dissipation begins as soon as the exhaust valve opens and continues as the exhaust slug moves further into the header and ultimately the exhaust system. Visualize a balloon blown up and tied-shut on a hot Sunny day – now toss it into the deep freezer . . . with a temperature differential of just 90 degrees the volume of air trapped by the balloon quickly collapses causing the balloon to shrivel. The same thing happens to a slog of exhaust as it moves from the exhaust port to the tail-pipe – and the temperature differential is much greater so the collapse is even more pronounced.

As a result the sizing convention needs to consider average flow rate – not the max or the min – which is a function of the displaced volume and the temperature differential. For a long tube header primary or head-pipe the mid-point or 2x can be used. Which is where the sizing convention comes from (Cubic-Inches/3456)*(RPM)*(TeF)=(Cubic-Inches/3465)*(RPM)*2=(Cubic-Inches/1728)*RPM after combining the constants in the formula.

Some suggest that if one is building say a shorty header or a log-manifold then a thermal expansion factor of 2.25x or 2.5x should be utilized because of the close the proximity to the exhaust valves. Still others argue that the tube should be sized closer to the peak expansion say 3x – but using super high exhaust gas temperatures is a whole other kettle of fish – starting with why is the EGT so high in the first place? Some will have you believe that the 2x factor is good enough for any application.

Others will remind that 100% volumetric efficiency is unlikely so the Cubic-Inch-Displacement value should be reduced to .9 or .85. And then the chemists will chime in and remind us that the combustion process for gasoline actually increases the expelled content by 7-10% depending on combustion efficiency. Considered in concert the VE and Combustion effects essentially cancel each other out.

For me the thermal expansion factor remains open for debate – but 2x seems a great starting point.

For the curious there are many scholarly papers out there discussing this EGT topic in extreme detail. The emissions police of the past have funded a lot of research in this area whilst trying to engineer catalytic converter placement into then contemporary exhaust system designs. This piece out of Aristotle University is an interesting albeit difficult read at times. For me the most important take away from the paper is that long-primary-tube exhaust systems retain the least amount heat – and failed to lite the catalyst in testing. The comparison charts are interesting – using the exact same 1.8L inline 4 they compare EGT and Tube Temps for a 4 into 1; a 4 into 2 into 1; and a 4 directly into the Cat design.