Selecting the proper turbocharger for your engine involves many considerations. Not only are the facts about your specific engine necessary, but equally important is the intended use for that engine. The most important approach to these considerations is a realistic mindset. In other words, if you’re turbocharging an engine that is presently rated at 200 hp in its naturally aspirated form, you’d probably love to have it produce 600 hp. However, that may be unrealistic inside of the additional collection of modifications you intend to do. If you’re looking for a nice power increase for all-around street driving, a 50-percent increase is more realistic and matching a turbo to this level of increase will produce more satisfactory results. A 300 percent power increase (200 to 600 hp) is possible in many engines, but increases like that are reserved for competition engines that have an array of additional modifications, both internal and external, that all work together to achieve this level of power. One of the most important factors in determining which turbocharger is most appropriate is to have your target horsepower in mind. But you have to be realistic about what you’re shooting for.
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The application and intended use of the vehicle is extremely important as well. An autocross car, for example, would require a rapid boost rise for fast acceleration, whereas a Bonneville car running long straights is more concerned with horsepower at higher engine speeds. Indy cars frequently adjust the turbo for short tracks versus long tracks because of how critical the turbo match is to optimize flow at specific engine and vehicle speeds. Tractor pull applications will likely see the highest engine speeds right at the start of competition, and as the pull progresses, the load is progressively increased much like a prony brake until the engine is maximum loaded down by the pulling sled. These different uses require different turbo matches.
The term Volumetric Efficiency, or VE, is a very important term and concept to understand. Maximizing engine VE raises it’s potential for horsepower and RPM. With the exception of fuel and ignition modifications, most of the traditional aftermarket high-performance engine parts essentially raise the engine’s VE. Forced-air induction is all about increasing VE. But what is Volumetric Efficiency exactly?
An engine’s VE is a comparison of an engine’s calculated, or theoretical, volumetric flow rate of air, versus its actual capability. An engine has a fixed displacement, for example, 300 cubic inches. That displacement will theoretically flow 300 ci every two engine revolutions (a four-stroke engine must rotate twice for all cylinders to complete all four cycles). In theory, there would be a linear relation to airflow and engine RPM where doubling the revolutions per minute would double the air displaced by the engine. If an engine were able to flow exactly as much air during operation as the theoretical calculation says is possible, that engine would have a VE of 100 percent. However, in reality that rarely happens.
While there are some engines that achieve 100 percent or higher VE, most do not. There are many factors that impede the engine’s ability to meet 100 percent volumetric efficiency, some intentional, some unavoidable. For example an air cleaner housing and filter will typically impede intake airflow, but you don’t want to operate your engine without air filtration.
The common methods used to increase VE in naturally aspirated engines are things like larger valves, more valves per cylinder, improved intake manifold designs, enhanced valve timing using different camshafts, free flowing exhaust systems and more. But perhaps the most limiting factor to VE is time.
The critical factor here is simply how long the intake valve can be held open. The calculation for 100 percent VE assumes a full cylinder charge of air without regard to time. During actual engine operation, the intake valve is only opened for a short time. The higher the RPM, the less time the intake valve is off its seat. Therefore actual VE is not a constant. It’s typically an efficiency ratio at one spot in the engine’s operating range.
If we did a simple calculation to express, in real time, just how long an intake valve is opened, it places the challenges to make horsepower into perspective. Let’s assume an engine was operating at 3,000 rpm, and it had a camshaft that contained 230 degrees of intake valve duration, measured at the crankshaft. The intake valve would only be opened for 0.0255 second. The reduction in valve open time is inversely proportional to engine speed. This same engine now operating at 6,000 rpm would only have the intake valve opened for half as long, or 0.0127 second! As you can see, volumetric efficiency will limit the upper engine RPM potential of any engine design, as it simply can’t breathe well enough in the short period of time dictated by the mechanical events. So, hot-rodders have been working for years to maximize the time we have.
The reason turbocharging has such a dramatic impact to engine performance can be better understood using this concept of volumetric efficiency. In a turbocharged engine, time still limits how long the intake valve is open, but if the intake pressure is greater than atmospheric pressure (boosted), then we can force more total air volume in during the valve opening. The quality of that air is improved for combustion purposes because its density has also been increased. The combination of boost pressure and air density compensate for the time-limiting aspect of the valve events and allow boosted engines to achieve well over 100-percent VE. But when maximizing total horsepower output, even turbocharged engines will benefit from many of the very same design improvements done to enhance VE on naturally aspirated engines.
As mentioned above, a given engine will have better or worse VE over the RPM band. Every engine will have its sweet spot, which is the point in an engine’s design where, at full throttle, the volumetric efficiency is at its highest. This is typically the point where peak torque will be found on the torque curve. Since VE will be at its highest point, maximum fuel efficiency or BSFC, measured in pounds of fuel per horsepower, per hour, will also be at its peak efficiency.
When calculating the proper turbo match, VE is an important element to consider, as it is an important contributor to determining the airflow demand of a given engine. This will be discussed further in this chapter’s section on compressor matching calculations.
Ballpark matching isn’t an official term, but a rough approach that can be used due to the more sophisticated way in which today’s turbo suppliers are marketing turbochargers. Historically, turbos weren’t marketed in the automotive performance aftermarket as widely as they are today and they were rated only in their mass-flow capability, not in terms of horsepower compatibility. However, a reasonable turbo match can be obtained by using a realistic horsepower target, and simply choosing a turbocharger from a manufacturer where your target horsepower is located right in the middle of the horsepower band it is capable of supporting.
For example, in the above situation where your target is 300 hp (a 50 percent increase over the stock 200 hp) the Garrett brand GT2860RS with a 76 trim, also known as the “Disco Potato,” might be an excellent choice. That turbo is rated to be compatible with horsepower applications of 250 to 360, so 300 is right in the middle of its capacity. This ballpark method allows for both surge margin and broad flow range of operation, as will be discussed in the following paragraphs. This is not the most scientific way to apply a turbo, but it can be done because much of the science of the horsepower match has already been done for you by the manufacturer rating a given turbo in its horsepower compatibility. Also, because many of the turbos offered for sale today have been built with an internally designed wastegate, subtle adjustments in boost pressure optimization can also be performed after the car is test driven. While most turbo experts will raise their eyebrows at this overly simplified approach, it can be very successful for the average turbo car project and it makes the confusion of turbo matching a rather simple process indeed.
The process of properly matching a compressor to your engine uses a set of assumptions, all of which are intended to be close and approximate in order to reach a reasonably close turbo match. An approximation is necessary because of an engine’s varied RPM, the entire manifolding system (and its efficiency), corresponding fuel flow, and many other aspects that will affect the final match. This is why most consider turbocharger application both an art and a science. If it were a pure science, then the engine dynamometer cells used to refine turbo matches for production engines would not be necessary.
There is no shortage of reference materials on the subject of how to match a turbocharger to an engine. Most all are very sound, and some make it very difficult to follow along. We’ll attempt to progressively work our way through the math involved and explain what we are doing as we go. This way the concepts of what we are doing, and why, are best understood. After all, the objective of this book is to help you apply a turbocharger to your project and obtain good results, not confuse you with overly complicated math.
Let’s begin with some fundamentals of air flowing through an engine. As discussed in Chapter 1 an engine, in its simplest form, is an air pump. The turbo is simply going to help increase the total airflow through that engine. Now, when we think of the engine as something that makes horsepower, we need to consider the flow of air and fuel through that engine as mass flow. The more mass of fuel and air, in their correct proportions, that flows through the engine, the higher the horsepower. But for now we’re just going to concentrate on the airflow portion of that mass.
That being said, we must next understand that in the world of turbochargers, the airflow is measured in pounds of mass per minute. In a naturally aspirated engine, airflow is typically measured in cubic feet per minute (CFM). But once you begin compressing the air to pressures above ambient conditions, the correct method of measure is to rate it in pounds of mass. This is because once the air is compressed, it has an entirely different amount of oxygen content, or air density, than the same CFM at ambient pressure. Further, under ambient conditions, CFM will not always have the same oxygen content due to temperature and atmospheric pressures. That’s why jetting changes are so commonplace in carbureted engines, while today’s EFI systems use mass-flow sensors to correct for this variation.
Let’s assume we intend to turbo a 3.0-liter gas engine rated at 190 hp at 6,500 rpm in our street/strip car. The next assumption is that we wish to operate the turbo with 10 lbs of boost pressure (a reasonable boost level for which we can obtain proper fuel octane levels to control fuel ignition and avoid detonation). The first thing we do is to convert its displacement into cubic inches, allowing easier conversion into our ultimate objective of determining lbs/min of air mass flow.
3.0 liters =
3,000 cubic centimeters (cc)
Displacement in cc / 16.387 =
Cubic Inch Displacement
3,000 cc / 16.387 = 183 ci
Now that we have our displacement in cubic inches we need to calculate the engine’s airflow in CFM in its naturally aspirated state. This is very easily done as follows.
CID x 0.5 x Max RPM / 1,728 = CFM
This formula simply takes the engine’s size in cubic inches times .5 because in a four-cycle engine it takes two complete engine revolutions for all cylinders to complete their cycles. This value is then converted from cubic inches per minute to CFM by dividing it by 1,728, the number of cubic inches in one cubic foot (12 x 12 x 12 = 1,728).
So, this formula with our engine looks like this,
183 x .5 x 6,500 / 1,728 = 344 CFM
Next we take the 344 CFM and we adjust it to reality. The above calculation contains an assumption of 100- percent VE. Since most engines in a naturally aspirated state do not operate at 100 percent VE, we must adjust our calculation to compensate for this. While VE varies with engine speed, valve, etc., timing experts have historically agreed that an average of 80 percent is logical in most cases. However, in many of today’s engines operating with four valves, variable valve timing, and computer designed induction systems, it may be more logical to use a higher figure. For the sake of your calculations, you can use the following suggestions regarding what VE value to assume.
Typical two-valves-per-cylinder, push-rod engine: 80% VE
Four-valve engine: 85% VE
Four-valve engine w/variable valve timing: 95% VE
So for our example, we’ll assume it’s a four-valve engine: 344 CFM x .85 = 292.4 CFM
Now that we have the CFM flow potential of the engine in its naturally aspirated state we need to adjust it by the rise in actual airflow once turbocharged. We do this using a density ratio chart. It is possible to calculate a density ratio for each level of compressor efficiency at a given pressure ratio. This calculation says for any given compressor efficiency line there will be a resulting increase in air density at a given pressure. The density ratio chart plots air density as a function of compressor efficiency and pressure ratio. The pressure ratio is easily determined by taking ambient pressure, which is 14.7 lbs at sea level, adding it to your boost pressure, and then dividing it by ambient pressure. This gives you the absolute pressure ratio.
Boost Pressure + Ambient Pressure, P2
Ambient Pressure, P1
= Pressure Ratio
10 + 14.7 / 14.7 = 1.68 PR
Now that we know the pressure ratio is about 1.7 we can obtain the density ratio (DR) from one of our density ratio (aftercooled or nonaftercooled) charts. If boost is expected to exceed 7 lbs, it’s advisable to use an aftercooler. In the example, we’re talking about 10 lbs, so I’ve used the aftercooled chart.
By looking across the X-axis to the 1.7 PR and extending up to the 74 percent compressor efficiency line, we can then read across to find a density ratio of 1.5. The 74 percent number is used because it ends up being a conservative and realistic efficiency average to expect.
CFM x DR = CFM with our turbo
292.4 x 1.5 DR = 438.6 CFM turbocharged
Next we simply convert the turbocharged CFM into lbs/minute mass flow by multiplying it times standard air density, which is .069 lbs mass/cubic foot of air.
438.6 CFM x .069 = 30.26 lbs/min of air mass flow
The entire formula look like this, (0.5 x CID x Max RPM / 1,728) x VE x DR x .069
It’s really just that easy. One of the double checks is a basic ratio of 1 to 10. It takes about 1 lb mass of air per minute to make 10 horsepower. To double-check our example, we need about 30 lbs/min of mass flow and 30 lbs/min x 10 = 300 horsepower, so we should meet our goal.
To improve the accuracy of your calculations it may be worthwhile to use the 1:10 air mass-flow to horsepower ratio initially to basically “shop” for the potential turbo of choice. Then, using compressor maps (see Chapter 5), determine the turbocharger that has your basic mass-flow and pressure ratio located right in the middle of its highest efficiency. Using that efficiency number, go back through the calculations for mass flow to verify your match.
Now, how do we use this 30 lbs/min mass flow data? First let’s discuss the use of compressor maps. To build a compressor map, a turbocharger is operated on a performance test stand, where a diesel fuel fired burner feeds hot air into the turbine and drives the compressor. The test stand operator has complete control over the burner’s air supply and fuel flow to simulate virtually any engine exhaust flow parameter. Meanwhile, the test stand operator also has complete control over the compressor discharge stack leaving the turbocharger’s compressor. Fully instrumented, the test stand is able to collect all of the flow parameters of a turbocharger’s compressor and turbine. The data points collected at various speeds and choke conditions allow for the mapping of compressor and turbine flow characteristics. These maps are useful as each compressor and turbine combination is matched and applied to engines.
The compressor map has pounds of mass flow on the X-axis with the absolute pressure ratio on the Y-axis. At each speed, compressor minimum flow is found by restricting the outlet stack leading from the compressor to find the smallest amount a compressor will flow without going into a condition called surge. Those points at various speeds and pressures form a line called the surge line, which is where the flow potential begins for each compressor’s map.
As the test operator opens the compressor discharge stack, the pressure ratio at a constant turbine speed will begin to drop off as mass flow rises to the point of choke flow, or where the compressor ceases to efficiently compress the air to a useable degree of adiabatic efficiency. The compressor’s discharge temperature is measured and calculations relative to inlet temperature and pressure ratio are performed to measure adiabatic efficiency. The adiabatic efficiency is displayed by the efficiency numbers expressed on each compressor map. A 100 percent adiabatic efficient compressor is where no heat is added or taken away from the job of compressing the air. In reality there is no 100 percent efficient machine, but the efficiency of each compressor is rated relative to that figure. As a general rule, you should not match a compressor to an engine if you expect it to be less than 65 percent efficient. Below that level of efficiency, the compressed air becomes too hot and increases the thermal loading on the engine and rapidly increases the likelihood for detonation.
Today’s turbocharger compressors will commonly hit efficiencies as high as 78 to 80 percent. But the wide operating RPM band of gasoline engines will operate any given compressor through a wide range of flow efficiencies. Each compressor has areas of varying efficiency called efficiency islands.
Find the nearby compressor map of the Garrett “Disco Potato.” As we can see our 30 lbs/min requirement at max RPM of 6,500 rpm shows up in a nice high efficiency island of about 72 percent, at a 1.7:1 pressure ratio. If we consider the fact that we’re going to be running our wastegate at 10 lbs boost, and that we want full boost at about 4,000 rpm, we can also calculate this point using the same formula as above except we use 4,000 rpm instead of 6,500 rpm. This will calculate about 19 lbs/min at a 1.7:1 pressure ratio, which is also in a high efficiency island over 76 percent. Once you plot both points on the map, you can see that from 4,000 rpm through 6,500 rpm with boost controlled at a 1.7:1 pressure ratio, the turbo will be running around 110,000 rpm turbine speed and it will flow right through the highest efficiency part of the compressor map. This would represent a very good match.
As the above match points show, there is a good operating margin between the match points and the surge line as well as the other direction of max flow, or choke. It’s best to stay away from matches that allow the turbocharger to operate near the surge line. Airflow instability can happen near surge, known as choppy air noise. This can adversely affect engine performance. A turbo simply won’t operate in surge. Once a turbocharger compressor hits surge, a loud chirping or bark is heard as the pressure head tries to revert its direction and the turbo instantaneously stops and starts again. This process can happen in a very rapid succession and hammer the turbo thrust bearings, while the compressed air gets superheated due to rapid recycling. A turbocharger operated in surge will likely fail very soon.
The trim of a wheel is a ratio used to describe both the turbine and compressor wheels. Trim is used in reference to the basic flow potential of a given wheel’s machining dimensions. Every wheel, both compressor and turbine, may have several trim configurations within that wheel casting. A given compressor wheel casting for example, could be “trimmed” such that it has two, three, or more trim configurations available. It is the trim that determines the wheel’s flow range and pressure characteristics.
Trim is calculated using the inducer and exducer diameters of the wheels. Note that the inducer of the compressor wheel is the smaller inlet diameter where the fresh air is induced into the compressor wheel, while the exducer is at the maximum diameter. The exducer has two primary components, the overall diameter and the tip width, sometimes called the tip height. In a trim calculation, however, the tip height is not part of the trim calculation.
Trim is calculated as:
(Inducer / Exducer) x 100
Example: Inducer diameter = 88mm
Exducer diameter = 117.5mm (88 / 117.5) x 100 = 56 trim
As a rule, as the trim is increased, the wheel can support more mass flow of air. Turbines are typically not as sensitive to flow changes as compressors. Because of this, it’s common for there to be more selections of compressor trims within a turbo model family than turbine trims.
For the most part, compressors with large inducers and smaller exducer diameters will flow large volumes of air (mass flow) at lower pressures, while smaller inducers with larger exducer diameters will flow less mass at slightly higher pressures. A large inducer with a large exducer diameter is a high-pressure, highflow compressor used in competition applications such as tractor pulling.
The trim characteristics of a compressor wheel will greatly influence the compressor map and thus that turbo’s flow potential. An example can be seen in the comparison in maps between the GT2860R and the GT2860RS, on page 66 in Chapter 5. They are very similar turbos from the same family except the 2860R uses a 55 trim while the 2860RS uses a 62 trim. Both compressor wheels are 60mm in diameter, but the 2860RS has about a 3mm larger inducer diameter. So, when tweaking your match and you wish to change compressors, if your pressure ratio is working fine, but you suspect you’re a bit short on air, a larger trim may be what you want, but make sure the trim you are switching to has a larger inducer by comparing the new wheel OD (outside diameter) to your wheel’s OD, if they are the same, but the new wheel has a larger trim, then the inducer is larger and this formula will tell you by how much.
The turbine end of the turbocharger is quite a bit more challenging to understand than is the compressor end. On the compressor side the compressor takes in air that is relatively predictable and only slight changes in barametric pressure and temperature affect what a given compressor with a given efficiency will produce at a known pressure ratio. In other words, there are not so many variables involved. It is quite common for turbines to easily confuse people due to the more complex considerations involved in what makes them react. But why is that? It’s because turbines have all of the complexity of the compressor side variables, plus you’re running the air feeding the turbine through an engine which further complicates matters. Air to fuel ratios, actual VE, charge-air coolers, manifolding, and more all go into turbine matching. This is why engine and turbocharger manufacturers use dyno cells to verify matches, the math for turbine matching is accurate and specific, but the variables are many and almost never known for sure. The following section on turbines will not necessarily give you a cut and dry manner in which to match a turbine to your engine. It is intended more to increase your comprehension of what is going on inside of a turbine and how to consider adjusting your match for optimum performance.
Perhaps the very best thing you can do for your turbo project is to not forget to place a pressure tap in the turbine inlet just past the turbine foot. Most people only focus on boost, but if all you’re looking at is boost you’re only measuing one side of the equation. You must know your turbine pressure or you’re flying blind. This can be easily accomplished by creating a 1/8-inch pipe tapped hole and installing an air tight fitting as a pressure tap. Run a couple of feet minimum, of steel tubing from the turbine to separate it from the heat of the exhaust before connecting it to a high-temp flexible line. This line will essentially be your “turbine boost gage” line, or actually turbine inlet pressure. You want to make sure that your turbine inlet pressure does not dramatically exceed your boost pressure or you’re developing too much back pressure and killing your power.
Applying a turbocharger to an engine, while involving quite a bit of science, is still an art. The lower the amount of test equipment you have, the more art it becomes. While turbines are perhaps the most confusing part of turbocharger matching, to make matters even more difficult, manufacturers don’t tend to publish turbine maps. They prefer to keep them a bit closer to the vest. In order to help you achieve a better “feel” for matching and tuning a turbo to your engine, the following discussion is designed to help you understand some basic principles and important nomenclature.
The most important aspect of matching the best turbine, and therefore the best turbo, is to be realistic as to what you will be doing with the application. If you want low-end stoplight-to-stoplight, headjerking acceleration, that is one thing to consider. If you want midrange power and economy, that’s another. But, if you want high-end horsepower without concern for low-end response (like a Bonneville racer), that is something altogether different. However, done properly, the realistic sizing of the compressor will place you in a model series that will likely have a suitable turbine selection for your application. This is because turbine and compressors are relatively matched in flow range because of the power balance necessary between the compressor and turbine ends of the turbocharger.
While the process of mathematically performing a turbine match is not really logical for an aftermarket application, understanding why and the elements necessary do help in understanding turbines in general.
Compressor maps plot mass flow as a function of pressure ratio. Turbines basically operate in a similar manner, but instead of plotting their mass-flow as a function of pressure ratio, it is plotted as a function of expansion ratio. The expansion ratio is basically the inverse of pressure ratio. The expansio ratio is basically the turbine inlet pressure in absolute value (gage plus ambient atmospheric pressure) divided by the ambient pressure. It’s a thermodynamic law that high pressure always seeks lower pressure just as hot always seeks cold. The reverse cannot happen. The higher pressure seeking the lower pressure is the principle driving force in the turbine. A turbine flow map will use expansion ratio on the X-axis. The following equation is how turbine expansion ratios are calculated.
EMP + Atmos / Outlet P + Atmos = Expansion Ratio
EMP = Exhaust Manifold Pressure
Outlet P = Turbine Outlet Pressure
Atmos = Atmospheric Pressure
Next, the turbine corrected flow is calculated in mass-flow for the Y-axis. This becomes a problem to correctly calculate due to the lack of accurate values available. The formula requires turbine pressure, which you will not know, and the engine’s actual exhaust temperature, another value you will not know. The standard formula for turbine corrected flow is as follows:
MF √ ([EGT + 460] / 519)
(BP + EMP) / BP
= Turbine Corrected Mass-Flow
MF = The Engine’s Actual Mass-Flow
EGT = Exhaust Gas Temperature
BP = Barometric Pressure
EMP = Exhaust Manifold Pressure
The numeric constants correct to absolute values. If you’re lucky enough to be developing your engine on a dyno however, access to turbine maps is not impossible. Usually, manufacturers will share turbine corrected flow maps with individuals who are engaged in advanced application engineering and have a need for this specific information. But in the hands of automotive enthusiasts, manufacturers have found that turbine maps create more questions than they answer.
The process of mathematically matching a turbine to an engine is quite a bit more involved than the relatively simple calculation of intake air mass flow. As with the compressor however, a ballpark match can be done for one simple reason. All turbochargers tend to run a power balance across the engine, meaning that the power driving the turbine is relatively equal to the power required to drive the compressor. Further, there is a pressure balance between the turbine and compressor sides of the turbocharger and the turbine selection options will tend to take that into account.
Even turbocharger application engineers do not use turbine flow calculations when applying today’s range of turbos to aftermarket retrofit situations because they are already cataloged and flow matched, compressor-to-turbine.
Up until about the last five years, turbocharger manufacturers did not have such a wide selection of turbo models specifically sized and catalogued for automotive retrofit. But today there are many choices to choose from and much of the work has already been done. Therefore, simply matching the anticipated horsepower to the midrange of what a turbocharger is rated for can be done successfully and make the job of creating your own turbo project much easier than in previous years.
The best way to successfully apply your turbocharger is to size it correctly by the mass flow demand for the compressor and select the model and trim accordingly. The turbine end will have trim characteristics that must be understood, however. As you select your choice of turbo for your application, it’s recommended that you select a turbo that has more than one choice of turbine housing size to help tune your match. For example, if your air mass flow and horsepower are right in the middle of a turbo’s capacity, and if there are three turbine housing sizes available, the middle size rating in A/R is a wise choice to start out with. This allows you room to go up or down as you tune your match.
When we think of an engine, we typically begin with the intake side. After all, that’s where the engine starts, with intake air. That’s also how we determine the relative size of the turbocharger (compressor matching). If you “think like a turbocharger,” then it all starts with the turbine inlet. The turbocharger can’t do anything until it receives its energy from the exhaust ducted into the turbine. While we’ve already discussed the fact that compressor housing size or A/R has little effect on the compressor, that is not the case with the turbine housing. Remember that the turbo starts with the turbine and the energy you are using is in the form of heat and pressure, just like the garden hose nozzle.
Let’s revisit the A/R ratio, or literally, Area divided by Radius. It’s a simple mathematical relationship between the size of the turbine wheel, which dictates the radius, and the throat square area of the turbine housing. You can think of it as a rating of the exhaust gas swallowing capacity of the turbine section. To calculate the A/R ratio the, area of the turbine housing is measured in square inches of a cutting plane line that passes through the turbine’s gas passage at the tip of the tongue, divided by the radius from the center of the turbine wheel’s axis of rotation, to the centroid of the volute. The tongue tip is the entry point of the turbine housing where exhaust gas flow begins to reach the turbine wheel inducer. The centroid of the volute can be thought of as the center of mass in the volute shaped exhaust passage, or the dynamic center (DC) of the exhaust passage. The dynamic center is that point which divides the volute area, or scroll as it is sometimes referred to, so half the flow passes above and half the flow passes below. It will therefore not necessarily be in the linear center of the exhaust gas passage when thinking of it in a two-dimensional manner. The centroid, or dynamic center is a threedimensional perspective and accounts for the potential non-symmetrical cross-sectional configuration of the turbine-housing volute because it concerns itself with flow dynamics and gas velocity as a function of the radius in the housing, hence the “Dynamic Center” relevance.
Turbine housing designs utilize what is called a free-vortex flow field. In this design, the exhaust gas tangential velocity will vary inversely with radius. The tangential velocity component approaching the wheel tip will have approximately the same value as the wheel tip peripheral velocity. This is why it would be a mistake to assume that the same A/R ratio used in different diameter turbine wheels would have similar volumetric flow patterns. A smaller A/R will raise exhaust pressure but only within that turbine family where the same turbine wheel is used.
1) 3-inch area / 3-inch radius = A/R of 1
2) 1.5-inch area / 1.5-inch radius = A/R of 1
In each of these examples the A/R is “1.” However the volumetric flow capacity of each is dramatically different from the other. Imagine a 3-square-inch orifice feeding a 6 inch diameter turbine as in example 1, versus only a 1.5-inch orifice feeding a 3-inch diameter turbine wheel. In this case it can be easily understood how different the volumetric flow of each combination is, yet the A/R of each turbine is the same. Remember that A/R is only a ratio of the values within a given turbine housing, not the values themselves.
The practical application of this perspective comes when you are tuning or retuning an application to understand that: a smaller A/R only has relevance to raising exhaust gas pressure inside of the same turbine family. If you change engine dynamics such that you switch turbo models, the same or similar A/R to what previously worked is not an indication of what A/R you may need on your new turbo match.
If there are turbine wheel trim selections in the turbo model you’ve chosen, the turbine wheel exducer diameter also provides for more subtle adjustments to optimizing exhaust gas pressure for a best match. Sometimes the jump from one A/R to the next available A/R for your specific model is too radical of a change. In this case, a slightly larger or smaller turbine wheel exducer may provide the desired result. Hopefully there are turbine wheel trims to choose from in your turbo selection. If not, more radical adjustments can be made by certain shops that have the knowledge and expertise to re-profile the turbine wheel and insert the turbine housing with a plug to cut a new contour. This is typically reserved for special extreme applications where unique compressor and turbine combinations are not readily available off the shelf from a turbocharger manufacturer. But these types of specific trim adjustments will not be necessary for the majority of high-performance projects.
Lastly, if your turbo selection uses an internally wastegated turbine housing, chances are the turbine trim of the wheel and the choices of turbine housings available have taken the wastegate’s presence into account by allowing exhaust backpressure to rise more quickly for drivability and response since the wastegate is present.
The wastegate performs a very valuable function. The wastegate is simply a valve that opens the turbine housing to relieve excess exhaust backpressure, triggered by a boost pressure actuator. The actuator is activated by boost pressure fed from the compressor. It is adjusted such that a small turbine housing A/R that provides boost pressure at a low to mid-range RPM, doesn’t over boost as the engine climbs in speed thereby over speeding the turbo and over boosting the engine.
Many off-the-shelf-turbo models will have a wastegate built integral into the turbine housing and that makes your application much easier. However, in-line wastegates are also available on those turbo models where a wastegate is not designed integral to the turbocharger. The adjustment and application of wastegates will be further discussed in Chapter 8. The wastegate is the simplest form of what is called a variable geometry turbine, meaning it has the capability to act like a small turbine under low-flow conditions, but then act like a larger turbine in a highflow, high-demand situation.
In many commercial diesel engines, turbos are applied without a wastegate. This can be done due to the relatively narrow RPM band of operation. Most commercial diesels will typically be operated at engine speeds ranging from about 1,100 to 1,200 rpm where they will hit peak torque, to about 1,800 to 1,900 rpm. A gasoline engine on the other hand will operate over a much broader RPM range. A high-performance automotive engine may see boost at 3,000 rpm, but continue to operate well above 6,000 rpm. With the exception of the loss in volumetric efficiency at the much higher RPM, the engine just doubled its airflow by doubling its rotational speed. When you think about this situation, it’s easy to see why a waste gate would be not just helpful, but necessary.
Written by Jay K. Miller and Posted with Permission of CarTechBooks