We dig into some aspects of intake manifolds that you may not have contemplated, fully understood, or thought worthy of further thought. You may discover some useful information.
When you consider that a chunk of metal standing between a carburetor or fuel injection system throttle body can affect so many engine functions, maybe we should look a bit deeper into how it’s able to become so involved. But before we lay out some of these reasons, causes, and what we might do about them to optimize power and efficiency, let’s define the basic types. We’ll be referencing the following fundamental manifold designs as we develop the balance of this story.
As a rule, although not always a requisite, all intake passages (runners) that are joined to a common volume (plenum), may be defined as a single-plane arrangement. Even such a design for which the plenum is partially or fully divided can fall into this category. In most applications, depending upon runner cross- section area, single-plane manifolds are associated with a higher range of engine speed than other types. In fact, so-called high-rpm “Tunnel Ram” manifolds are of single-plane design. So also was the Smokey Ram that Edelbrock manufactured some time ago (that’s an entirely different story and a recollection for some of you “older timers”).
This design contemplates two different levels (planes) to which runner sets connect with a common volume (plenum). Both V-type and inline cylinder arrangements can accommodate a two-plane (or single-plane) concept. Whether such a plenum is partially or fully divided, these manifolds are usually noted for volumetric efficiency boosts in an rpm span somewhat lower than a single-plane design.
Independent Runner design
If none of a manifold’s runners connect to a common volume, even if some sort of a “pressure” or “pulse” balancing tube joins one or more runners, the design is defined as an independent runner (IR) design. Some mechanical and/or electronic fuel injection systems use this concept. Typically, they can be rpm limited and prone to fuel “stand-off,” a condition related to reversion and to be discussed later in this story.
Since we know that flow in a intake manifold, regardless of design, is unsteady and bi-directional (during some phases of the intake cycle), we next need to examine what such instabilities are, what causes them, when they occur, and what may be done about mitigating their negative influence on net power and combustion efficiency. In one way or another, they are all related and need to be understood.
Manifold Pressure History, the Intake Cycle and Reversion
Let’s examine events and conditions that occur throughout a complete intake cycle. Try to visualize these events in very slow motion. Depending upon the efficiency (completeness) of a given exhaust cycle, pressure in the cylinder is higher than atmospheric pressure at the time the intake valve begins to open. At this point, the most prevalent material in the cylinder is exhaust gas (byproducts of combustion). So until cylinder pressure becomes equal to manifold pressure (approaching atmospheric in a normally-aspirated engine), exhaust residue flows back into the intake manifold. There are various terms that describe this event. Perhaps the most common one is “reversion.” Once these two pressures (cylinder and manifold) pass beyond being equal, and while the piston continues its descent during the intake event, fresh air/fuel mixture charges and un-burnable exhaust residue from the previous event are passed into the cylinder. When the intake cycle is complete, we have a cylinder with fresh air/fuel charges diluted with uncombustible exhaust residue. So you can see that the extent of the reversion event can materially affect both combustion efficiency and net power.
What is Intake Manifold Runner “Cross Talk”?
Reversion also introduces another of its influences on an engine’s overall combustion efficiency as what we’ll call “cross talk” among the manifold’s runners. Remember we said that for a brief period there will be some flow back into the manifold? Suppose during this brief time, another of the engine’s cylinders is already into its intake cycle. At this point, some of the reversion material passes out of the runner experiencing reversion and is inducted by that other cylinder.
Until cylinder pressure becomes equal to manifold pressure (approaching atmospheric in a normally-aspirated engine), exhaust ...
Stepping back from all this for a moment, this “cross talk” follows the engine’s firing order over and over again as engine speed increases. But as it does increase, there is less time for these reverse pulses to participate in “cross talk” or even influence the pressure differential across the carburetor. Don’t forget, carburetors are pressure differential devices, so they’ll deliver fuel whether flow is normal or reverse (leading to the fuel vapor hovering above the carburetor), leading to the “stand-off” condition previously mentioned. Stated another way, as rpm increases, contamination of adjoining cylinders gives way to increased contamination of each cylinder in its intake cycle. Reversion becomes contamination, as a function of higher rpm
Over the years, a number of devices and techniques have been employed that are applicable to an engine’s tendency toward reversion. We won’t delve into them here, but the type of intake manifold can affect its containment. For example, a single-plane manifold tends to dampen reversion pulses more effectively than a two-plane design and certainly better than an IR version. Of these, the IR design (having no plenum volume to absorb reversion pressure pulses) is often characterized by a “fuel cloud” or “vapor cloud” hovering above the runners. Such conditions play havoc on attempts to develop an efficient air/fuel charge calibration (or fuel curve). And in some instances, reversion conditions can be so severe as to leave exhaust gas residue in the intake manifold, all the way to the underside of the carburetor. The stuff just doesn’t burn well, a second time around.
Further to the issue of fuel curves, if we define “metering signals” as the positive net pressure (atmospheric and other carburetor-influencing pressures), the reversion phenomenon and fuel delivered to the engine varies with manifold type. In particular because of its larger overall internally-connected volume, single-plane manifolds cause weaker (softer) metering signals than either a two-plane or IR, in that order. In other words, if you calibrated a carburetor using a single-plane manifold, you’d likely be pulling jet size out of the carburetor if you switched to a two-plane design, simply because of the stronger metering signals typical of a two-plane design, no other changes involved. In fact, when plenum dividers come into play (single- or two-plane manifold design), metering signals typically increase as well.
The Effects From Changes to Runner Cross-Section Area
Even though we stated that intake manifold flow is both unsteady and bi-directional (we just discussed reverse flows or reversion), there is ultimately a net rate and volume of flow. Let’s define this net rate as “mean flow velocity.” A fair amount of research and data collection has been conducted over time supporting the concept that the m.f.v. at peak volumetric efficiency (essentially peak torque), is around 240 ft/sec. Hold that thought for a moment.
Now visualize that we have a flow passage across which we place a pressure differential. In reality, this is what we have during the main part of the induction cycle; e.g., atmospheric pressure pushing air and fuel into the engine. Since atmospheric pressure (for purposes of this discussion) is relatively constant, the rate of net flow will be governed by the cross-sectional area of the passage. If we increase this area, still maintaining a constant pressure drop across the passage (same rpm), the net flow rate decreases. If we make the area smaller, flow rate increases. Hang on. This is pretty important stuff.
Now, remember we said that the m.f.v. at peak torque will be pretty close to 240 ft/sec, so we’ll use that value. Visualize an engine on an engine dyno, running at peak torque, constant rpm. Two factors will affect this m.f.v.; runner cross-section area and cylinder displacement. But if we fix cylinder displacement to a given value, only runner section area remains to move this m.f.v. to wherever we want in the rpm range. The concept becomes a tool, either for the design or modification of any of the type intake manifolds we’ve included in this story.
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.
If the passage had a circular section area, there would be only one “shortest path” in a turn. Using the “D” section shape provided multiple shortest paths along the flat portion of the “D” (see illustration). During one particular manifold design, we discovered that a trapezoidal section not only helped control flow pressure on the large side of this shape but provided additional pressure control by varying the length of the shorter part of the shape (see illustration). And even if you cannot produce a true trapezoidal section, trending toward this configuration helps minimize the pressure differential from short- to long-side of the turning passage. If you’ll spend a little time will the sketches, this should appear more apparent.
It all comes down to creating (or modifying) manifold runners that minimize energy differences across the sections of a passage, particularly during changes in flow direction. Obviously, air will follow a change in direction much easier than the heavier fuel. You’d like for both to change direction as if no change had occurred, and trending toward a trapezoidal shape can help resolve the problem. As you would expect, once air and fuel separate, reuniting them in a fashion that supports combustion can be problematic.
Air/Fuel Charge Distribution vs. Dry Flow Measurements
Several factors are involved with this. First, we need to seek ways of making cylinder-to-cylinder dry flow as equal as possible. This at least provides us a shot to balance out comparative cylinder-to-cylinder power. And while it doesn’t account for the dynamics of flow direction changes and the differences in momentum between air and fuel, balancing dry flow to all cylinders (manifold runner paths) is a starting place.
The use of “downstream” and “upstream” Pitot tubes connected to a water manometer will help identify regions in a runner where boundary layer separation is occurring. It is in these areas where a degree of flow restriction and air/fuel separation can take place. (Note sketches showing two type of Pitot tubes.) While the “J-tube” shape tube will show where airflow shear or separation is occurring at or near flow surfaces, the “I-tube”) can be used to determine comparative flow rate (velocity) at any given point in the path. You can even perform flow section “mapping” by using this tube at the runner entry as well as further into the passage.
Intake Manifold Performance vs. Cylinder Head Port Flow Performance
It is sometimes helpful to think of an intake manifold’s runners as extensions of the cylinder head ports. This is certainly a consideration when attempting to design an intake “tuned” to certain engine speeds; e.g., the total intake path. But for purposes of manifold selection and use of existing designs, you can focus on intake manifold runner section areas and apply the previously shared equation about mean flow velocity, piston displacement and peak torque rpm.
Depending upon piston displacement and anticipated range of rpm, circle track racers have discovered it’s sometimes beneficial (particularly for corner exit torque) to use intake manifolds of slightly smaller runners than cylinder head ports. Generally speaking, as engine displacement increases, the internal size of intake manifold chosen should grow, accordingly. The same rule of thumb applies to increases in operating rpm, especially higher levels. However, since many racing intake manifolds have longer runners than the cylinder head ports, it’s possible to have a material influence on torque range by, once again, applying the little “sizing” equation mentioned earlier.
There is also another option to modify manifolds in a way that broadens the torque curve, using runners of two different section areas (every other cylinder in the firing order). Some of you may recall an Edelbrock intake manifold once named the Victor 4+4. In this case, the four inboard runners (comparatively short) were intentionally sized with slightly larger section areas than the out-board four (longer than the inboard four runners).
Depending upon piston displacement and anticipated range of rpm, circle track racers have discovered it’s sometimes beneficial (particularly for corner exit torque) to use intake manifolds of slightly smaller runners than cylinder head ports
Once again, based on the equation we provided, you can see that the desired m.f.v. required a slightly higher rpm to be reached for the inboard runners than the outboard runners. The result was a somewhat flattened and broadened overall torque curve, at least as contributed by the intake manifold. And, by the way, when matched to a set of headers of two-size primary pipes (linked to the different in intake manifold runners), the combination was particularly effective for corner exit torque and then again at the flag stand.
Flow Surface Texturing
We decided to include a few thoughts on this subject, even though there may be some skeptics about the value of working certain flow surfaces in an intake manifold. Actually, the concept can be carried out beyond the intake manifold to such places as piston tops and combustion chamber walls. When properly applied, the notion is to provide some boundary layer excitation, especially in locations where air and fuel tend to separate. By so doing, some of the separated fuel becomes re-suspended in the inlet air stream. We know for a fact that this technique is particularly beneficial when also applied to piston crowns and certain areas of the combustion chamber.
Inside the intake manifold, roughening the entire length of the runners (at least to the extent surfaces can be reached), plenum floors and around the entries to the runners can help, depending upon the fundamental design of the manifold. You’ll be able to read the effects by slightly reduced b.s.f.c. data. The best barometer is on-track performance because some of the benefits show up only in a transient mode of operation, as provided on the track.
Dynamic Cylinder-to-Cylinder Mixture Distribution Fixes
Particularly in circle track applications, you will likely discover that distribution fixes sorted out on an engine dyno will not always satisfy what is required on the track. We recall years ago working on an intake manifold project with Smokey. In fact, it was a cross-ram, single 4V design he’d done for Chevrolet and their Trans-Am program. After a fair amount of work on the engine stand, it turned out the “fixes” did not apply to the track at all. Especially in this particular design, there was a considerable amount of fuel running around in the manifold as a function of not only air/fuel separation but from the sheer centrifugal forces present on the track.
Even in contemporary single-plane, single 4V racing intake manifolds, such forces can upset distribution patterns otherwise worked out on the dyno. That’s not to say the static engine method is of no value. Just don’t be surprised if what works on the dyno fails to translate directly to the track.
Some Concluding Thoughts
It’s important to understand how influential an engine’s intake manifold can be to the total volumetric efficiency landscape. If you consider that v.e. and net torque are closely linked (and they are), then manifold selection becomes a focal point to having an engine that delivers power where you’d like and need it to be. In particular, the decision needs to include the range of engine speed where power will be required most often, and then you can make intake and exhaust system selections that match this range. The little equation provided earlier will help you through this process as well.
One suggestion is that you also seek the advice of either your favorite engine builder or the parts manufacturer. The chances are good that both have knowledge they’ll share and intended to help with the manifold selection process.