Mind you, a similar (although not exact) set of circumstances are playing out on the exhaust side of the engine that will contribute to the overall torque curve (a blend of intake and exhaust system contributions to the net), but we’re presently talking about the torque curve as produced by the intake manifold. (On a personal note, I’ve designed and tested intake manifolds on engines for which the exhaust system was tuned beyond the speed range of the engine, thus producing a torque curve only contributed to by the intake manifold. So the concepts we sharing with you are valid.)

Let’s get back to this idea of setting the induction’s peak volumetric efficiency point (roughly at peak torque) where we’d like. And, by the way, note that we’ve not yet addressed the issue of runner length. Well get to that in a couple of paragraphs downstream.

What we’d like is a simplified mathematical equation that, although perhaps not hair-splitting accurate, enables us to manipulate the intake manifold’s torque curve, using piston displacement, rpm and runner section area. Well, in fact, there is one.

Runner section area = (peak torque rpm x cylinder volume) / 88,200

This format will enable you to determine intake manifold section area as a function of where peak torque occurs, given a known cylinder volume. If you already know the manifold’s runner section area and want to calculate where peak torque rpm should occur, you can algebraically transpose this equation into the following:

Peak torque rpm = (runner section area x 88,200) / cylinder volume.

Ideally, this approach assumes no runner taper, although most intake manifold runners employ taper. In such cases, you can calculate an “average” runner section area, using the runner entry area and exit areas. In addition, if you have the benefit of an engine dyno data sheet and apply the “peak torque” version of this equation, it’s possible to determine if the manifold is too large or small for the engine by simply comparing your calculated peak torque rpm to the actual peak torque rpm from the dyno sheet. If the demonstrated peak torque rpm is higher than the one calculated, the engine is “over ported.” And, of course, if the peak torque rpm is lower than the calculated value, it’s “under ported.” Just make certain you apply the same mathematical approach to the exhaust system’s primary pipe area (using pipe i.d.) so that the exhaust system’s contribution to net peak torque isn’t overlooked.

Before we move on to another topic, it’s worthwhile knowing that the engine’s exhaust system (header sizing) can contribute to the overall torque curve. And because you can apply these equations to headers as well, you may want to run the math on the exhaust primary pipes to see where they fit into overall torque output. They will influence how the intake manifold is performing, in the total scheme of things.

**The Effects of Runner Length**

OK, let’s address this issue. We’ve tried to emphasize how runner section area affects torque output, indicating that the area sets the peak torque rpm point. Runner length does not change that. However, by increasing or decreasing runner length, the torque curve tends to “rock” about that rpm point, either clockwise or counter-clockwise, depending upon whether the runner length is increased or decreased. Decreasing the length adds torque above the peak rpm point, increasing the length adds torque below the peak. In fact, this is no different from primary pipe length in a collected exhaust header system. Hopefully, by now you’re beginning to discover some similarities in how comparable dimensions between intake manifolds and header exhaust systems affect their respective performances. More of this will appear a bit later.

**Some Thoughts About Plenum Volumes**

Years ago, using far more complex and comprehensive mathematical models than the little equation shared earlier, we conducted some engine dyno tests that demonstrated the influence of plenum volumes on torque production. It turned out that a volume far in excess of what is customary in commercially available intake manifolds was required, if these are to be “tuned” in a fashion similar to runners. All this boils down to what plenums in a conventional intake manifold can affect and how they impact overall engine performance.

We know that smaller volumes (particularly the distance between carburetor bases and plenum floors) tend to produce higher fuel metering signals than smaller volumes. We also know that the larger volumes tend to dampen reversion pulses better than smaller ones. In addition, particularly in two-plane designs, we know that the upper plane has a greater influence on the carburetor than the lower one. So-called “stagger jetting” (smaller jets for the upper plane than the lower) can account for this.

There is also a body of evidence that the proximity of plenum floors to carburetor bases can impact the quality of air/fuel charges entering runners. The shear mechanical separation of air and fuel upon colliding with plenum floors can show up as disrupted brake specific fuel consumption (b.s.f.c.) numbers, particularly at the higher engine speeds. Carburetor spacers are one path to addressing this problem; e.g., the greater the carburetor base distance from the plenum floor, the less the tendency for air/fuel separation…and the weaker the fuel metering signal.

**The Effects and Benefits of Spacers**

Before we depart from this subject, some further discussion might be helpful. Essentially, the exit velocity of air/fuel charges entering a given plenum chamber is a function of carburetor size, engine speed and piston displacement. All three can come into play when deciding whether or not to use a spacer. And while the use of a four-hole spacer (for 4V carburetors) can be helpful in curtailing the effects of reversion pulses, the air/fuel charge discharge velocity will remain largely the same as without the spacer. An open spacer will allow some decay of the exit velocity as the air/fuel charges change their basic direction from vertical to something approaching horizontal upon entry into the manifold’s runners. It’s this change in direction and the fact air and fuel are of dissimilar masses (air is also compressible and fuel is not) that they tend to separate.

This same condition (air/fuel separation) can also occur along the intake path and combustion space, so let’s take a few minutes to examine how that can occur and what can be done about it. But first, suppose we conclude this part of our discussion with some comments about runner shapes.

**Thoughts on Determining Proper Runner Shapes**

While dry airflow benches are of value and wet flow versions have their benefits, remember we said that a running engine’s induction system operates in an unsteady, bi-directional fashion. Flow benches don’t replicate these conditions. So, we are left with the necessity of finding ways to at least approach “real world” airflow that’s as stable as possible.

First of all, when you create a pressure drop across a flow path (be it on a flow bench or running engine), air tends to find the shortest path. Keep that foremost in your mind because of the factors included one deals with port shape, particularly during flow directional changes. What you’d like to create is a shape that allows a minimum of pressure differential between the short- and long-sides of a turn. Early on, this is pretty much what led to D-shaped passage turns, where the flat side of the “D” was the short-side of the turn.

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