A Note from Jim: The stated objective of this story is to introduce you to some notions that may be of imme-diate help to circle track engines operating under rules that allow non-OEM parts. It is based on years of design and experience dealing with flow passages on both sides of the cylinder. Likely, you won't agree with all the points. That's OK, but try them out on the flow bench. Build some sample flow boxes. Measure the indicators. U.S. Patents have expired on most of them, so we'll reach behind the patent claims and talk about how they were derived. You can take it from there.
Elementary physics confirms that when two different pressures exist, there will be some form of "activity" directed toward establishing an equilibrium condition. In short, nature would like for the two pressures to be equal. For example, when the inlet cycle of a normally aspirated engine begins, there is a difference in pressure between cylinder and atmospheric pressures. Air outside the engine is not "sucked" but forced by atmospheric pressure into the cylinder.
What you should remember is that this flow would like to take the shortest path into the engine, which is key to this discussion. In fact, on the exhaust side when cylinder pressure replaces atmospheric pressure and expands into lower pressure conditions outside the engine, the natural flow path is still the shortest route, even though this may not be possible at high rpm (simply based on fluid dynamics). If you take nothing else from this story, remember both these points. Now, let's get specific.
Intake Systems Suppose we walk through the basic pressure relationships in an intake path, beginning when the intake valve first opens. Initially, residual cylinder pressure tends to be higher than that in the inlet track (manifold and cylinder head ports). Containing residual and unburnable combustion by-products, cylinder pressure causes flow back into the inlet path. Over time, this has been called "reversion" or "back flow" into the induction system.
This dirt Late Model will...
This dirt Late Model will benefit from a new set of headers. bobby clark
For purposes of this discussion, we'll omit any comments about how valve overlap may play into reversion. It doesn't relate to what follows. Also, at least in this story, we'll not include the effects of "wave motion" and other unsteady flow conditions that exist in both intake and exhaust systems. Our topic here deals with port passage considerations in the presence of steady-state flow. It's a premise that forms the basis for a number of flow models.
Further into the initial stages of the inlet cycle, cylinder pressure will equal atmospheric, after which cylinder filling begins. Obviously, the shorter the reversion period the greater the potential for increased volumetric efficiency and power increases from additional fresh air-fuel charges. So it's critically important that flow toward the cylinder is optimized. It is here that port shape becomes important. And it's not just about port size or volume, as you'll soon see.
Now, remember we spoke about how airflow tends to seek the shortest geometric path, all else being equal. If we're dealing with a flow path that is without bends, pressure distribution across any given port section remains relatively the same (constant). Just as soon as the path provides a directional shift, however, conditions can change significantly (Figures 1, 2, and 3). Note how the pressure distribution profiles change, favoring (again) the shortest flow path.
In an intake system for carburetor-equipped engines, we're dealing with a "working fluid" comprised of both air and fuel, each of which has materially different mass (weight). In a steady-state flow environment, the ability of air and fuel to remain "in suspension" is not much of a problem, certainly by comparison with what happens during directional changes. When subjected to changes in flow direction, the kinetic energy differences between the two elements (air and fuel), based on differences in mass, causes them to mechanically separate, and the consequences of this are classic.
So, ideally, we'd like to have changes in flow direction at least "act" as if there have been no changes. In other words, it would be helpful if flow passage section areas could be configured (for directional changes) that encourage pressure distribution profiles resembling those in a straight path of flow (you may want to read that again, because it's important). It's additionally important to keep in mind that this same concept, in terms of pressure differential manipulation, also works on the exhaust side of the engine. But we'll get to that in a minute.
Exhaust Systems Here, we'll define "exhaust systems" as they pertain to exhaust ports and directional changes in tubular exhaust manifolds (headers). Although the working fluid in an exhaust system differs in both operating temperature and viscosity from an intake system, favorable passage section design is comparable. Not unlike on the intake side where the "short path" notion applies, exhaust passages respond to the same approach, except at high rpm when flow rates are high and the so-called "short side" path plays less of a role (except at low valve lift). Nevertheless, and certainly at low exhaust valve lifts, you'll find that the fundamental shape we'll be describing in a few paragraphs has merit as it does on the intake side.
Flow Passage Area vs. Torque Output Because section area has a direct bearing on port flow velocity and port velocity is tied in with volumetric efficiency (potential torque output), we need to examine how area relates directly to torque. This is important, independent of port shape.
One approach to examining the importance of section area suggests that at an engine's peak torque (volumetric efficiency), there will be a value for mean flow velocity that appears in virtually any normally aspirated internal combustion engine. Typically, the value assigned to this velocity is on the order of 240-260 feet/second. Of the ways that have been used to verify this concept, building intake and exhaust systems sufficiently "tuned" at a wide difference in rpm has shown this value for mean flow velocity to occur at peak volumetric efficiency. Interestingly, the lowest values for brake specific fuel consumption can also be observed at this point, in the majority of cases.
Regardless, as briefly mentioned in a previous CT tech story, you can use port section area to create this flow velocity at rpm where you'd like the intake (or exhaust) system to provide a v.e. boost. Using this same "tool," you can also design/modify these systems at different engine speeds to help tailor torque curves to specific track requirements. And while the following equation (also previously offered) is a greatly simplified version of a more mathematically complex method of analysis, it is useful for designing/modifying components, selecting intake and exhaust components, or evaluating existing parts packages in a surprisingly accurate way. As previously indicated:
torque peak = (cylinder volume x port section area) / 88200
In this case, "torque peak" (or v.e. boost) is the rpm at which you'd like the intake or exhaust to be particularly volumetric efficient; "cylinder volume" is the calculated swept volume (displacement) of one cylinder; "port section" is the actual section area of the passage in question (don't rule out the option of using multiple-sized passages in a given system for a broader range of torque boost); and the "constant" (88200) handles conversions to allow you to use the English measuring system (feet, inches, seconds, and so on). If you're calculating header pipe section area, don't forget to account for tubing-wall thickness.
We've included this again to emphasize its usefulness and, in fact, direct relationship to the port shape concept that follows. It has proven its merits over a long period of time.
Port passage taper vs. no taper Actually, this subject could morph into a complete story, but not here. We'll limit comments to the fact that both can have merits, under the proper circumstances. On the notion of "no taper," however, we'll offer one thought for you to at least consider because it's linked to the mean flow velocity issue.
Even though there is considerable "pressure excursion" activity along intake and exhaust paths, you can generalize these conditions as being akin to kinetic energy systems. As such, their kinetic energy content may be mathematically described as one half the product of mass times velocity squared (1/2mv^2). Since we suggested mean flow velocity at peak torque (v.e.) to be on the order of 240-260 feet/second, would you not want as much flow mass as possible moving at this critical velocity when it is reached? If there is taper in a passage (and velocity is a function of section area), less flow mass will be moving at this velocity when its associated rpm is reached. Multiple "degrees of freedom" intake and exhaust systems (those tuned to multiple rpm points) frequently contemplate the use of no-taper passages. It's something to think about.
What happens when flow direction changes? Here again, we're attempting to simplify a rather complex issue. Fundamentally, under any given conditions of flow in a passage, there will be a "pressure distribution" profile that represents the distribution of kinetic energy in the passage. While some amount of fluid friction at the interface between the "working fluid" and passage walls is unavoidable, the effects flow-direction changes have on pressure distribution across any given cross-section along the path can be more problematic.
Dissimilarities in mass (weight) between air and fuel exacerbate the problem, keeping fuel suspended in air (wet flow), and the effects of gas dynamics (dry flow) can combine to make passage design for directional changes a real challenge.
Consider some of the conditions of flow during a 90-degree turn (elbow shape), using a constant circular section area (see Figure 1). Note that the "short path" in this instance is along the passage's floor and, in fact, amounts to a single line of virtually no width. In addition, because of this feature, you can see how the pressure distribution "profile" is skewed to favor this shortest path. As such, the highest flow velocity is toward and along the passage floor. Neither the surface nor shape of the "outside" surface plays much of a role.
Were we to flatten the floor (Figure 2; still maintaining the same section area) to provide some additional "short paths" that are equal to the one in Figure 1, the pressure profile becomes less concentrated along a single line, essentially making the passage wider and more efficient, as compared with the one in Figure 1. The result is a time-honored D-shape sometimes found in exhaust ports. This D-shape concept can be applied to many flow passage requirements where the short-side can (1) be designed accordingly or (2) modified to produce the same results.
In both examples (Figures 1 and 2), you can see that there is a lack of flow activity on the "long side" wall, thereby causing some pressure distortion in the pressure distribution profile. One feature that can improve flow efficiency is to create a shape that provides a more uniform distribution of pressure across each section in a passage, including times when the direction of flow changes. Configuring section areas that work toward uniform pressure distribution is one possible solution. This leads us to the next topic for discussion: trapezoidal section shapes.
Trapezoidal passage sections Take a look at Figure 3. You'll notice how the pressure distribution pattern has become more uniform during the bend. Had the passage begun with a square cross-section, it would have transitioned (gradually) into the trapezoidal shape you see at the apex of the bend, up to the apex of the bend, and transitioned (again gradually) back into a square section at the passage's exit. The same would have occurred regardless of the entry and exit passage shape; e.g., circular, rectangular, and such.
Not surprisingly, gradually widening the short-side base and making the long-side base more narrow (forming the trapezoid) caused velocity pressure along both paths to become more uniform, causing the flow (wet or dry) to transition through the turn without realizing the full impact of a directional change. Both wet and dry flow conditions stand to benefit from the method.
Here's where you might want to build a test flow-box. It needn't be fancy. One way is to whittle out a wooden "plug" that's shaped like the passage you plan to evaluate. Using 2-inch-square patches of fiberglass cloth and some casting resin, wax-coat the plug and layer it with resin-impregnated patches of 'glass, randomly overlapping them to provide coverage and strength. A couple of layers are sufficient.
When the 'glass hardens, split it down its length and remove the plug, covering the cut with a few more patches of resin-soaked fiberglass cloth. Now you can attach your new "flow-box" to a mounting flange that's connectable to the bench, form a molding clay radius for an entry, and you're ready to take some measurements. Although somewhat crude, this model will help you visualize the pressure distribution relationships being described.
Using a manometer and Pitot tube for downstream flow velocity data, you can begin profiling sections through the length of the passage (see sidebar on pressure mapping). Note the similarity in velocity pressures, comparing the short- and long-side paths. Compare these same two areas at the entry and exit (assuming you are using the recommended entry radius). By adjusting the transition from square to trapezoid and back to square again, you'll be able to approach equality of pressure maps at the entry, midpoint of the bend, and exit. At that point, the flow's recognition of a directional change is minimal.
This concept is particularly applicable during conditions of wet-flow where the mechanical separation of air/fuel charges can lead to a number of consequences, virtually all of which are undesirable and include lost power and potential damage to parts. By minimizing pressure distribution differences that can lead to energy imbalance and losses in flow efficiency, it's possible to optimize volumetric efficiency and reduce problems associated with intake and exhaust flow paths.
Understand that what we're discussing is outside the intentions of what you may have heard or read about the benefits of "controlling" air or mixture motion in induction systems. This technique (sometimes referred to as "swirl" and "tumble") is more directed to exhaust emissions and fuel economy benefits, particularly at engine speed (flow rates) well below what you find in a circle track engine. Taken to extreme, these features in the inlet track can lead to poor mixture control and unwanted variations in mixture ratio as a result of excessive motion of fuel particles.
Determining the proper trapezoidal shapes Experience has shown that unless you have access to a comprehensive computational flow analysis (CFA) or comparable software package, experimental investigation is probably the best approach. Depending upon specific dimensional characteristics of the flow path change you're confronting (space limitations, abruptness of the change, and such), you'll likely find the width relationships between the upper and lower bases of a trapezoid won't become the same from passage to passage. In some cases, the smaller base width may be roughly half that of the larger, but don't expect one rule to hold for all cases.
What you're striving to accomplish is a pressure distribution pattern that shows flow rates along the short-side of the turn to be as equal as possible to the long-side. Again, to the working fluid, this presents an environment where a pressure differential (change in flow direction) has not been experienced . . . or at least the negative effects are minimized. As in virtually any parts design or modification process, there are unavoidable compromises. The idea is to accept the fewest number of them that have any impact on a project's outcome.
A Few Concluding Thoughts If you persevered through this entire story, it's likely that you're wondering how you can apply what's been discussed. Having eaten my share of cast-iron dust while grinding on ports and heads, washing aluminum shavings out of my hair, and using eye drops to excess, I can tell you the foregoing is good information. Smokey once told me to try talking to CT readers as if they were sitting across the table from you, nursing a cold cup of coffee, and trying to learn from what you've learned. The material you've just read can be applied to your weekly racing engine. It's up to you to find out how well.
A "saturday night" approach to port flow mappingConsidering the large number of flow benches owned or accessible to weekly circle track racers, the following information is intended as an introduction to pressure distribution mapping in flow passages. Whether you are examining a wet- or dry-flow system (particularly if you're not using a wet-flow bench), the suggestions and illustrations provided are intended to get you off dead-center in these types of studies.
Essentially, you'll want to either build a "system" of three water manometers or one to which you can interchangeably attach two different types of probes. For the sake of simplicity, we'll call one of these a "velocity" probe and the other a "boundary layer" probe. As is typically the case with measuring devices or test equipment, each has benefits and limitations. From experience, you can expect the former outweighs the latter for these two methods.
You can build the probes in a variety of ways. One proven method uses malleable steel tubing of about 0.040-inch o.d. and 0.015/0.020-inch i.d. These can be attached to flexible plastic tubing connected to the manometer(s), the other ends of which are vented to atmospheric pressure. A suitable manometer(s) can be constructed with lengths of hard plastic tubing connected by flexible tubing that forms the U-bends and filled with colored water. The manometer(s) can be mounted on a plywood backboard on which some form of "grid" or "scale" has been placed so that you can record numerical differences in manometric readings.
Figure 4 is provided to indicate the types of manometer readings you can expect from the use of a velocity and boundary layer probe. Manometer No. 1 depicts fluid deflection direction when the boundary layer probe senses pressure less than atmospheric, as in the case of layer separation from the passage surface or (in other instances) turbulent or unstable flow. Manometer No. 2 is what you should see when there is no turbulence and/or when the probe only sees atmospheric pressure (as positioned in the figure). It will also appear this way at or near areas of high and stable port-flow velocity. Manometer No. 3 shows gauge deflection direction when using the velocity probe, higher readings being a function of increased flow rate.
Now, study Figure 5. This is a visualization of how you can "map" a passage, section by section. Using the velocity probe, record readings at each "corner" of a given section, along with the highest pressure read elsewhere in the port; e.g., ideally at its center. Each of these readings is taken in the plane of each section. By assembling the data from all sections in the passage, you can determine the pressure distribution profiles for the entire path, entry to exit. Although there will always be compromises, the goal is to create a passage (including directional changes) that trends the center of pressure distribution toward the geometric center of the passage. This is critically important in wet-flow conditions and significantly important during dry flow. Remember, the lower the dynamic pressure, the lower the flow activity and the greater the possibility for turbulence (in dry flow) and air/fuel separation (during wet flow).
Using the "boundary layer" probe (open end facing atmospheric pressure), you can navigate these same sections as were examined with the velocity probe. In this instance, you'll be looking for manometric gauge deflections opposite to what you saw using the velocity probe. When you find them, it's an indication the working fluid (air or exhaust gas) has separated from the passage walls, creating a lower than atmospheric condition that hampers dry-flow efficiency and can lead to air/fuel separation during wet flow. You'll also discover that at the point of greatest velocity (within a given section plane), the boundary layer probe will not cause the manometer to move. This is because both sides of the manometer (at this point) are experiencing atmospheric pressure. It's a quick way to also map maximum velocity pressure in a passage (section to section).
Now, you may decide all this isn't very scientific. And, by some contemporary standards, it's not. There are far more accurate and expensive ways. But for our purposes, you'll quickly discover that what you're attempting to do is compare readings on a test-to-test basis, so units of measurement can be grid-count, portions of an inch or virtually any linear unit of manometer gauge information. Using this rather crude system, you can make serious progress creating (or evaluating) changes to achieve stable, uniform, low turbulence port flow.
At the end of the day, you'd like to minimize manometer deflection, as measured by the boundary layer probe, and equalize section pressure distribution with an emphasis on trending the center of pressure toward the center of the section, using the velocity probe. You can also use this method to evaluate parts known to work well, in the sense you'll develop a better understanding why. "Getting there" is one thing. "Knowing how you did" is another matter, and you need to understand them both.