So far I have covered a lot of technical details that could be described as peripheral. It’s important, but is not actually the hands-on stuff. I am now about to set that right.
This Tech Tip is From the Full Book, DAVID VIZARD’S HOW TO PORT & FLOW TEST CYLINDER HEADS. For a comprehensive guide on this entire subject you can visit this link:
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Optimizing Cylinder Head Airflow
With the classic poppet-valve style engine, air cannot make a straight shot into the cylinder. There has to be a bend in the port to accommodate the valve stem and, worst of all, the air must make its way around the head of a valve. So we can confidently assume that the number-one skill a head porter must develop is the ability to get air to “corner” well. If you have any doubts about this, refer to Figure 10.1. What are we looking at here? It is a port for an all out-race, two-valve engine.
The color gradients indicate velocity. Red is 450 ft/sec, transitioning through yellow at 350 ft/sec, to aquamarine at 250 ft/sec, and then to blue at 150 (or less) ft/sec. Despite this port’s very substantial short-side turn, you can see that the air has a propensity to move in a straight line as indicated by the ghosted arrow in the port. We learned earlier that air was a lot heavier than is commonly believed and this is a manifestation of that property. Because of its mass and the speeds involved, the air really wants to travel in a straight line. Check the high speed exhibited on the long-side valve seat. This demon-strates that valve seat form is a factor even at this head’s 0.900-inch lift.
The left-hand arrow indicates the combustion chamber wall. Examine this closely because the flow is not following the curve of the chamber wall but is in a free stream. This tells us we could use a chamber wall that was flatter (not concave) here (as I describe in Chapter 9). Also, because this is a very active area we need to make sure that there is as near minimal valve shrouding as possible on this side of the port.
Let’s now move to the right-hand arrow. Here, we see that because there is much less activity, de-shrouding this side of the chamber to the same extent as the long side is unnecessary and possibly detrimental to the overall performance of the head. Keeping in mind all we have so far learned on valve seats and valve shrouding, let’s move on to the ports, chambers, and CR, which has all-important effects on initial head design.
Valve and Flow
Within either port, the valve is the only component that, as it opens and closes, delivers a variable geometry. At low lift, the flow is entirely dependent on the size of the gap between the valve seat, the seat in the head, and the efficiency with which they flow. At high lift, the port size and shape primarily (but not wholly) dictate the amount of flow delivered.
A key proportion in the valve’s opening cycle is a lift value equal to 1/4 of the valve’s diameter. This is commonly known as 0.25D, and Figure 10.2 shows its significance. At 0.25D the “curtain area” is exactly equal to the valve head area. By applying the same criteria, we can say that the valve seat influence, though continually diminishing, is the major influence on flow up to a valve lift equal to about 0.18 to 0.20D. In other words, the seat is the number-one priority until the valve has reached a lift figure between 18 and 20 percent of the valve’s diameter.
At this point, you may ask how much difference a valve seat shape can make to the efficiency delivered during the lower lift regime. The answer is: a lot. Up to the 0.20D lift point, a typical production three angle valve seat usually delivers efficiency figures starting at about 65 percent and quickly dropping to around 55 percent. Spend a lot of time on a flow bench, and those figures can be as high as 85 percent or more.
A point worth raising here is that there is a commonly held belief, among many successful engine builders who specialize in high output two-valve V-8 engines, that too much low-lift flow hurts power. I won’t go into a lot of detail here, other than to put forward a simple explanation. First, the cylinder does not know how far an intake valve is open. All it knows is how much flow is being presented. Therefore, if too much flow is being presented at any given moment, it is because the valve events are not what the head wants. That means the wrong cam is being used. Claiming a cylinder head has too much low-lift flow is a little like claiming your race team has too much cash on hand!
Remembering that everything in terms of flow starts at the valve seat, let’s move on to the subject of the ports themselves. Although a flow bench is a prime requirement for optimal results, there are many ground rules that can be applied to improve a typical stock cylinder head of two or four-valve design.
If we assume the port is paired with a good valve seat design, priority number-one is maximizing airflow. Second, high airflow must be achieved by high-flow efficiencies, not by an overly large cross section. If the cross section is too large, the port velocity is low; this reduces cylinder ramming and increases flow reversion. The result: poor low-speed output with possibly no benefits at the top end.
Another factor that can greatly affect low-speed and, to a lesser extent, high-speed output is swirl or otherwise-induced mixture motion within the closed cylinder.
Next it helps to make sure no big problems exist with fuel management within the port. This comes under the heading of “wet flow.”
Finally, we have to look at the combustion chamber form as defined by both the combustion chamber and the piston crown. That simply means getting the required CR without producing a poor-burning chamber in the process.
The most basic port we could have is a round port (see Figure 10.4), which has a bend in it to accommodate the valve. Using this as a starting point, we can develop a port using some simple logic.
With any port, the radius of the short-side turn is usually the number-one obstacle to achieving good midto high-lift flow figures. F1 engines have a very large short-side turn radius, and the port’s downdraft angle is only about 30 degrees off vertical. This makes for a very simple port that requires very little in the way of Band-Aid fixes to make it work extremely well. Unfortunately, the dictates of less-than-ultimate power on a less-than-ultimate budget and low hood lines mean ports that are severely compromised. Figure 10.5 shows the basic steps that need to be taken for improved efficiency.
If you are modifying a typical parallel-valve head with two valves, the last step in the illustration is a very important one. Understanding that the port more than likely needs a bias is the key to getting those big high-lift flow numbers. This leads ultimately to high-lift flow efficiency figures that exceed those delivered by many four-valve heads.
The recovering flow efficiency of a parallel two-valve head is one of the reasons why this type of head responds to high valve lift, so bear that in mind when it comes time to spec out the cam and valve train. The target lift to shoot for is 0.30D for a hot street machine and as much as 0.35D for an all-out racer.
The optimal cross-sectional area for a given size of cylinder can vary somewhat, depending on the size of the intake valve, how tortuous the port is, and the bore/stroke ratio of the cylinder. A good starting point for the intake is to have a nearly parallel section of the port, about 1½ to 2 intake valve diameters up from the intake valve itself, sized to a cross section equal to 77 to 80 percent of the area of the intake valve.
The optimal area of this parallel section gets bigger as the port gets steeper and vice-versa (see Figure 10.10). This parallel section needs to extend, if necessary, into the intake manifold for about two to three times the diameter of the intake valve for a two-valve head and about four to five times the diameter if we are dealing with a four-valve head. For the record, it is better for the widest power-band and best torque if you err on the smaller side because this produces a “punchier” driving experience. Making the port too big can hurt output everywhere, as the test results from a small-block Chevy in Figure 10.13 show.
Although the sidebar on page 106 gives a theoretical picture of port area requirements, it does assume that the stream of air in the port can reasonably make it around the turn. Remember, the air has mass and is inclined to stay in a stream a little like water from a hose. What this means is the throat or bowl area before the valve becomes really critical. That is where the flow bench is needed to make the best job possible of getting the air around the bend and on past the valve.
One last point: The reason the different head configurations in the Sidebar “Calculating Peak Power RPM” below have differing maximum port velocities is tied to the fact they have differing port utilization characteristics. The more uniform the velocity is at the area considered, the higher the peak-power port speed is.
Limiting Port Velocity
Air drawn into the engine is done so in a stop-start fashion. There is a price to be paid in terms of the energy taken to accelerate the air up to speed for each induction stroke. As you may suspect, there comes a time when the energy required to accelerate the air to higher speeds costs more in power than any extra power the additional air may have produced. During the World War II era, Charles Fayette Taylor, a highly respected tech pioneer of the day, deduced that the peak port speed, at peak power RPM, reached a limiting value at about Mach 0.5 to 0.55 (at standard temperature and pressure, that’s 580 to 640 ft/ sec). Although this limiting number has been pushed up on really high-tech engines, such as F1 units (I have heard of numbers as high as Mach 0.62), those numbers hold really well when we are considering a high performance two-valve-per-cylinder engine and four-valve engines with ports of less than about 30 degrees of downdraft to the valve axis.
By the time the port’s Mach number has reached about 0.6, the power curve is very much on the way down for most engines.
So how do you determine the Mach number of your engine’s ports, short of measuring it? Well there are some formulas that were developed by Taylor and others but they almost all relate to slow-revving aero engines. The fact that we have so much of a better handle on tuned lengths today, especially the exhaust, means that determining the Mach number with any useful accuracy involves way more than just a single, simple formula. My recommendation here is that you get an engine simulation program, such as those from Motion Software (Dynomation) or Performance Trends (Engine Analyzer). With these programs, you can experiment with different port crosssectional areas to determine what works the best for your application, to within some pretty close limits.
Applied Basic Porting
Having plowed through the basics in theory, it’s worth asking: What’s it worth in terms of power? This obviously varies from head to head and engine to engine. The better the head is to start with, the more your efforts are likely to resemble a basic porting job. The following tests should give you a good idea of the worth of any porting or compression raising you are likely to do. Figure 10.16 shows what can be achieved with just an easy weekend’s porting exercise on a set of production 1989 Ford 5.0 heads.
But let’s go one step further. We start with a pair of airflow-tested heads on a Ford 5.0 Mustang mule engine and develop a set of baseline power and torque curves. From there, we look at what a set of as-cast Dart Pro 1 170-cc port heads can do. And then, the 170s are swapped out for a set of 195s and a set of basic ported 170s, milled for 1.2 points of extra CR.
The cam used for all the test heads was a 280-degree Comp Cams High Energy, single-pattern street roller (profile number 1474). The 1.6 rockers, after lash, delivered 0.560inch lift at the valves. Other than Comp Cams rockers and Icon pistons, all the other parts, including the timing chain, oil pump, water pump, etc., were stock. We also used a Moroso pan, as its greater volume and surface area help dissipate the heat of repeated dyno runs.
For induction, we used an Edelbrock Performer RPM Air Gap intake along with a 650-Holley-style carb and a billet Petronix distributor with a mechanical advance curve to suit the cam spec, provided ignition. You can see our dyno mule was far from exotic, and actually had considerable test time on the card already. At this point, making power might look simple—just make the heads as big as possible to flow as much air as possible. Unfortunately, there is far more to it than just plain old flow as measured on a flow bench.
As I have tried to stress, air has considerable mass and is heavier than you might think. If that air is moving fast, we can utilize the kinetic energy it contains to not only ram the cylinders at high speed but also reduce flow reversion at low speed. Getting that port area just right for a particular engine combination means more power at the top end, where a previously air-starved engine can now breath, and also at the low end. And we use the low end 95 percent of the time on the street, so it should at least be a realistic priority.
Another item on the agenda is swirl. The stock 5.0 head is very poor in this respect. Here, the test Dart heads show excellent swirl for a Ford-style casting, though it is still less than a typical small-block Chevy casting. Without good mixture motion, the combustion process is compromised. If a head has good swirl, it helps improve combustion quality, especially at low speed. In a nutshell, good swirl most often equates to good low-speed torque.
The last factor to consider is the compression ratio, which I cover later in this chapter. At this juncture, you are getting an idea of the importance of understanding just what the CR, especially when suitably high, can do for an engine’s output. This is as important for a street-driven machine as it is for an all-out race car. The compression ratio has a considerable influence on the size of cam that can be used before low-speed output becomes unacceptable. The higher the compression ratio, the greater the cam duration that can be used. Also more compression equates directly to improved fuel efficiency, and that is something that cannot be overlooked these days.
Now let’s look at a few physical dimensions of the heads under review here. First, the as-cast 170-cc Dart Pro 1 has a 165-cc measured intake port volume and a 62-cc exhaust volume. The flow figures of the heads tested are shown in Figures 10.17 and 10.18.
As a reference point, the stock heads are typically in the low to mid 120 cc on the intake and low to mid 50 cc on the exhaust. The 170 Pro 1 heads are intended to be used as a direct replacement on an engine that has an otherwise-stock bottom end. That means they must have valves that are not too big to be accommodated by the stock piston’s valve cutouts. To do this, the intake valve is 1.94 inches in diameter. That’s up appreciably from the stock 1.84 inches but significantly less than the typical 2.02 inches that can be used when aftermarket pistons are in the engine.
To make flow comparisons fast and easy, we’ll look at two reference lift points that, for all practical purposes, define the head’s ability to get the job done. These two points are the flow at 0.250 lift, and the flow at peak valve lift as delivered by the cam and valve-train. In this particular case that is 0.56, but to make life easier here we use the 0.550 lift point because it is close enough.
At these two key lift points, the intake on a stock head delivers 121 and 155 cfm, respectively. On the flow bench, the 170-cc Dart Pro 1 in as-cast form delivered 151 and 255 cfm. Let’s put that into perspective here. First, even at as low a lift figure as 0.250, the Dart head delivered flow numbers almost as good as a stock head at maximum lift. At the peak valve lift point, the Dart Pro 1 head produced a full 100 cfm more than stock!
Swirl was also measured during the flow tests. The Dart heads all showed excellent swirl characteristics. In spite of having bigger ports (which, all other things being equal, typically reduces swirl) all the Dart heads tested had significantly better than-stock swirl performance.
As far as port velocity is concerned, we have somewhat more challenging concepts and principles to get a handle on. The Dart head has a significantly bigger port than stock; so for a given flow, the velocity is less. However, the valve is bigger and so, for a given lift and flow bench depression, more air is pulled in and this boosts the velocity.
Okay, that all looks simple enough, but the engine does not inhale air in the same way that a flow bench does. Let’s consider the situation at low RPM. If the valve is open to the extent that it can more than satisfy the instantaneous need of the cylinder, the depression drawing air in could be lower than we see on the bench. The net result is that the port velocity could actually be lower even though it was (as in our case here) higher on the flow bench. If we throw all the variables of port velocity, swirl, and flow into the melting pot, and do not have many years of experience, it is difficult to confidently predict the outcome of this type of proposed head change.
Although swirl, because we are considering the exhaust, is out of the picture, the exhaust ports flow and flow velocity are also factors affecting the low-speed output as well as the top end. One of the worst enemies to low-speed torque, when a big cam is used, is exhaust flow reversion. Higher uniform port velocities on the exhaust side can have a considerable positive influence on low-speed output. For the exhaust side, our 170-cc Dart Pro 1 had, with its measured 63-cc exhaust port at the two key lift points, flowed 123 and 185 cfm, respectively. The stock 52-cc port, at those same key lift points, delivered only 92 and 121 cfm. This means that even though the port was bigger, the Dart exhaust had more velocity because of the greater quantities it could flow.
The 170s on the Dyno
With the flow, swirl, and velocity characteristics so far discussed in mind, let’s see how the as-cast 170 Dart Pro 1s fair on the dyno. Check out the curves in Figure 10.27. The green lines are the ones to look at, compared to the stock head results shown in black. Even though it had a much-bigger-than-stock intake runner, the combination of swirl, flow, and runner velocity was such that output was improved right down to 2,200 rpm. At this point, the Dart 170 put out a creditable 14 ft-lbs more than the stock heads. This combination also led to the peak torque going up by 16 to 17 ft-lbs and peak power by a very satisfying 68 hp. The usable top-end RPM figure also rose by about 700 to 800 rpm. That is pretty good for just a cylinder head change by any standards, especially ones that are still as-cast.
The Bigger Sibling
The 195-cc Dart Pro 1 differs from the 170; the intake measured 25 cc more and the exhaust 2 cc more. Also the 195 had 2.02-inch intake valves instead of the smaller 1.94s of the 170 heads. At the two key points of 0.250 and 0.550, the 195s flowed 159 and 268 cfm. That’s up by 8 and 13 more than the 170-cc head. For swirl, the bigger port was a little lower until about 0.400 lift, then it picked up to numbers very similar to the 170-cc head variant.
Here is how it plays out for port velocity. By enlarging the port by 25 cc, the mean cross-sectional area of the bigger port is 13 percent greater. However, the flow increases we see are about 5 percent at the 0.250 and 0.550 check points. The net result is a decrease in port velocity by a nominal 8.5 percent. In simple terms, this means that this head would be better on an engine that was either 8.5 percent bigger or turned 8.5 percent more RPM. The exhaust port, 3-cc bigger, was marginally better on flow between 0.300 and 0.500 lift. This makes them close to the same as the ports in the 170-cc heads.
The 195s on the Dyno
On the dyno, the 195s produced the output curves shown in dark blue in Figure 10.27. Notice the torque at low RPM is down compared to the 170-cc heads. This is how it remains until about 5,000 rpm. From there to 5,500 rpm, the bigger port heads matched the smaller ones. It was not until the RPM exceeded 5,500 that the bigger, higher-flowing ports produced any additional output. Even then, the power increase only amounted to 4 or so HP. Increased output at 2,200 rpm amounted to just 1 ft-lb more than stock. Peak torque was up by 12 ft-lbs and peak power by 72 hp. This means that although the 195-cc head is still a very effective piece on a relatively big-cammed 302, it appears much more suited to a 331or 347-ci engine where its port size is more appropriate.
If Some Is Good—More Is Better
From the tests so far, it looks like port velocity is instrumental in delivering more area under the curve. If keeping port velocity up to some key value, such as we are doing here, can produce a fatter curve without sacrifice at the top end, then we effectively make a smaller engine run like it has more inches. Let’s take the 170-cc heads and pose a question. How would they perform if they were given a basic porting job, plus the bigger 2.02-inch valve as on the 195-cc heads?
A basic porting job, as defined earlier, means limiting the metal removal to skinnying the guide bosses, blending in obvious irregularities (there were only a few of those), making the most of a progressive radius on the short-side turn, and, in this case, replacing the 1.94 intake valves with 2.02-inch ones. Installing the bigger valves also allowed for a more sophisticated-radius valve seat to be used.
All things being equal, this should improve the low-lift flow, which has the effect of improving the top-end output and helping power hang on longer after the peak point has been passed.
Unfortunately, flow graphs do little to show the extent of improvement at low lifts on the finished 170-cc Darts, so let’s look at some numbers to bring the point home. First the intake: The new valve at 2.02 inches diameter is 4 percent larger, so it has 4 percent more circumference.
This means, if it is utilized at exactly the same efficiency as the valve it replaces, it should be 4 percent better. If the seat form is also more efficient, that also increases the flow. Countering this is the fact that as the valve size increases, so does the shrouding caused by the cylinder wall.
Let’s see how the numbers shape up here. First, at 0.025 inches lift the flow with the new, bigger valve was up from 16 to 19 cfm, for a 14-percent improvement. Not bad for a starter. At 0.050-inch lift the gain was from 34 to 36 cfm for almost 6-percent improvement. We see this trend all the way up to the peak valve lift that we are going to use. On average, the flow increase is about 8 percent, and this has been achieved with a port volume that is only up from the measured 165 to 170 cc. That’s just 3 percent. What this means is that for any given flow rate, the port velocity has also increased.
Now we have heads that deliver both more flow and more velocity. If we have not altered the basic port form, we should also see a little more swirl activity. Our swirl meter confirmed this was in fact the case. So far, so good—now on to the CR.
The stock heads and the as-cast Dart heads delivered a measured compression ratio of between 8.87 and 8.94:1; so, for the sake of simplification, we average this out to 8.9:1. The ported Dart heads were machined 0.0050 to reduce the chamber size to a final 52 cc. This, on the mule’s short-block combination, gave a CR of 10.1:1. Using a basic equation for thermal efficiency, this increase in CR should deliver 2.9 percent more torque everywhere in the RPM range.
However, the formula assumes that the intake valve opens at top dead center (TDC) and closes at bottom dead center (BDC). We actually have a pretty big cam in this engine and the valves open and close way before TDC and way after BDC. This means the effective compression ratio, the one that applies using the intake valve closing point, is much lower than the theoretical, or static, CR. As a result, an increase of 1.2 actually gives the engine a bigger percentage increase of the CR than it initially seems.
The result is, with a big cam, the low-speed torque can increase considerably more than you might otherwise expect. Also increasing the CR causes the exhaust gas speed to increase at just the time when it’s most important to do so—right around TDC in the overlap period. This reduces the tendency for the exhaust flow to go into reversion and delay the onset of intake flow into the cylinder.
Modified 170s on the Dyno
Maintaining (or even slightly improving) the swirl and port velocity at low speed should mean we don’t lose anything at those lower speeds. While the increase in flow won’t really pay off until higher RPM, the increase in compression should return dividends everywhere, but most especially at low RPM. A check of the red output curves in Figure 10.26 confirms just that. If this had been a very short cam, we would have seen only about 8 ft-lbs increase at the 2,200 rpm point; but because the cam was quite big (for a street cam), an extra 16 ft-lbs was realized. This little exercise should bring home the importance of matching cam and compression; the bigger the cam, the more compression is needed to make it work.
Moving up the RPM scale, we see that the ported 170 Darts pushed up the peak torque to 20 ft-lbs better than the best of the un-ported heads. And peak power went up by some 30 hp more than the bigger-port 195 Darts and hung on longer, allowing shift points to be raised to near the 7,000-rpm mark. About 10 of those extra horses are from the compression increase and the other 20 from a combination of airflow and port velocity improvements.
From these results, you can see that having good airflow, port velocity, swirl, and compression results in a vastly superior power curve. The numbers speak clearly here. Not only was the low-speed torque better for street use, but the ported heads added a staggering total of 101 hp at the top end. That’s like having a 100-hp nitrous system installed that you never have to fill! Mission accomplished.
An increase in the CR was part of the equation that made the example head porting successful. It is a key ingredient toward making power, and understanding the implications involved helps make porting and cam selection decisions that put you at least one step ahead of your competition.
The Virtual Flow Bench
With anything that involves the passage of air through or over solid objects, it really helps to see where the air is going. As we all know, the problem is that air is invisible. All those tests we do to map out the velocity using the probes as described in Chapter 6 are to establish where the air is moving. Computational fluid dynamics (CFD) does just that, but until recently, it has been a valuable tool that only F1 and Cup Car teams could afford. But that is changing. Beginning in early 2011, I have been working with Design Dreams, a small company that has been developing ways and means to bring a virtual flow bench or wind tunnel to the market at a price that is little more than the most expensive commercial flow benches. The work here is being done by David Woodruff and his aim is to bring CFD capability to any serious head shop.
The downside to CFD is that you need to be able to develop a mesh that describes the form of the head. I am going to use some CFD results contributed both by David Woodruff and the ever-resourceful Dr. Rick Roberts, who could best be described as Edelbrock’s chief research engineer.
First, take a look at Figure 10.31. You can see that the area between the seats is the busiest but also how, on this shallow hemi-style head, the flow is virtually even out of both sides of the valve. This is a virtue of a head with nearly zero shrouding.
A CFD result from the same head as viewed from the side is also revealing (Figure 10.32). First, note the high velocity occurring on the beginning of the short-side turn. Immediately after the high-velocity plume, the flow separates from the port wall. This demonstrates the need to spend whatever time is required to get the short-side turn to be as effective as possible. In a real-world situation, that only happens if you flow the port at a working pressure drop, not a fixed 28 inches or so!
With CFD, you can flow the head not only with real-world pressure differentials, but you can also do so with the piston in the appropriate place. This means you have the ability to flow the head from the intake right through the exhaust port during the overlap period. When I am teaching students at a university or in seminars, I go to great lengths to emphasize the importance of optimizing the effectiveness of the overlap period. My time with flow benches, dynos, and in cylinder pressure measuring gear has shown that what happens during the overlap period, and the first 30 degrees of the intake stroke, is the making or breaking of a high output engine.
Using the traces of intake port, cylinder, and exhaust port pressures allows us not only to see, but to refine the intake and exhaust systems of test engines. So, we can feed a CFD program with real-world data, in order to realistically flow the head during the overlap period as well as the entire induction stroke. Tests, such as this, are very cumbersome, and in some respects impossible with a conventional flow bench. For instance, take a look at the interaction of the valve pocket with the valve during the valve’s and piston’s closest approach in Figure 10.31. Note that air wants to flow off the back face of the valve and into the cylinder.
Unfortunately, the process is short-changed by the fact that the valve cutout is shrouding the valve at this point. By modifying the valve cutout, as shown in sidebar “Piston Porting” on page 121, additional airflow into the cylinder can be had. I have tried pistons with and without the mods described here. On a nominal 520-hp mule engine, the so-called piston porting looks to be worth about 5 hp. Not much, but that’s 5 more that was not there before doing the mod.
Though in a different form, the same applies evxen more so to the tall piston crown on a big-block Chevy. On high-compression pistons, the dome that is near the wall, and is directly in line with the flow out of the intake valve during overlap, shrouds and reduces the flow. The same applies to the approach on the exhaust valve. I won’t go into more detail here except to say that if you want to modify big-block Chevys like a pro or better, my big-block Chevy book, How to Build Max-Performance Chevy Big-Blocks on a Budget, is what you are looking for.
It takes little imagination to appreciate the huge potential that an affordable virtual flow bench has for serious head development. By the time this book is published, David Woodruff of Design Dreams will be well on the way to having working copies of the virtual flow bench. If you are already into CNC head porting, applying the virtual flow bench is a straightforward job. If not, you need to do some drafting work to draw out what you want to test, or some digitizing on an existing port to create the mesh outline of what you want to test.
Written by David Vizard and Posted with Permission of CarTechBooks