Turbochargers and Diesel Engines
Here you will find a general explanation of the subject matter.
This second section will focus on a subject comparison.
Turbocharger #1 – High flowed factory turbocharger (grey line)
Turbocharger #2 – Designed turbocharger (blue line)
High Flowed Turbocharger
Generally speaking, a ‘high flowed’ turbocharger emphasises the turbocharger’s compressor ability to flow more volume than the factory turbocharger. These upgrades can vary from just adding a larger compressor wheel only, to also changing the size and characteristics of the turbine wheel to enhance throttle response, and compensate for the larger and much slower compressor wheel. Use of different materials and components to allow for higher shaft rpm is often integrated. A ‘designed’ turbocharger, is a turbocharger that has been specifically designed and built for each engine’s requirements targeting performance and addresses all correct flow criteria.
What is compressor surge?
To the turbocharger, compressor surge is a condition that occurs when the turbocharger’s compressor output air flow is higher than the engine’s flow requirements, combined with high boost pressure. When the conditions are right, a combination of lack of airflow, and the consequent rise in pressure, vs the high speed of the compressor wheel all cause the compressor wheel’s blades to effectively lose their “grip” on the air. Aerodynamically speaking, the air separates from the back of the compressor wheel’s blades, allowing some air to escape back out through the compressor intake. The resulting pressure drop in the pressure side piping allows the blades to “grip” again which, under the residual momentum of the turbocharger, will increase the pressure and the cycle repeats. This causes the characteristic fluttering sound of compressor surge, and also results in pressure pulses in the post turbo charger intake system.
In the above illustration, a clockwise rotating compressor wheel creates a low pressure area behind the compressor blade, that is quickly filled by the in-rushing higher pressure air flow. This subsequently causes the airflow to “grip” the back of the blade. If the compressor’s output flow becomes “static” at the blades this effect is lost (loss of grip), causing a back flow of built up pressure and volume from the intake system in a reverse direction.
Different types of compressor surge, and its cause.
For these examples the PR (Pressure Ratio) has been changed to PSI.
As viewed on a compressor map – Example 1A (surge under acceleration load).
Here we can see turbocharger #1 rides up in roughly the 60% efficiency island right on the surge line. At around 14psi of pressure the turbocharger isn’t able to produce anymore volumetric flow (Volume) indicated by the vertical line as flow stops increasing but the pressure continues to rise. At around 16psi with low volumetric flow (Volume), it crosses the line and enters the phenomenon known as surge (under load). Forcing its way through this surge area as the shaft speed increases, it is able to recover after some time. Once it reaches the maximum desired boost pressure, it can then start to flow (Volume) and move across the map through better efficiency islands away from surge. This is the typical result when a turbocharger (Turbine) is too small for a diesel engine’s output after fuel is corrected, and/or the turbocharger (Compressor) is too large resulting in a “miss match” of components. When surge occurs in these conditions, it can be particularly severe.
Different types of compressor surge, and its cause.
As viewed on a compressor map – Example 1B (surge under acceleration load)
As viewed on a compressor map. Here we can see turbocharger #2 pushing more volume for less compressor pressure keeping it away from the surge line. It rides up through around the 65% efficiency island before the maximum boost pressure is achieved and it continues to push volumetric flow (Volume) and move across the map in the same area as turbocharger #1. Turbocharger #2 is unable to reach surge under “full load” conditions.
Care should be taken when selecting a turbocharger so that it will not operate in the surge region at full throttle (WOT) or at any RPM.
Different systems for reducing compressor surge
Anti-surge compressor covers
Also known as Ported or Map width extension covers, feature either a bullet or slotted type whereby, the holes or slots intersect a groove just before the compression area of the compressor wheel. These styles are designed to bleed off air pressure to reduce compressor surge with in the compressor wheel blade area. This style works to some degree under acceleration surge conditions when the blades lose their grip and the exducer flow stalls, allowing the back flow of air to reach and flow out through the ported compressor groove back into the intake system (through either the bullet or slotted type holes). Ported compressor design covers, is primarily to target acceleration load surge and has minimal effect on deceleration surge conditions.
Blow off valves
Blow off valve type systems are a more effective way to eliminate all types of compressor surge (both acceleration and deceleration). Compressor efficiency is also increased because slotted systems continuously bleed off air during non surge conditions due to their design. A blow off valve only works when surge conditions are present. It does this by sensing differential pressures from the post turbocharger induction system to within the volute part of the compressor housing /pressure around the compressor wheel, and/or either side of the intake system’s butterfly valve.
Eclipse Turbo Systems ACRV system
ACRV is a bypass system designed and incorporated into the compressor housing to eliminate the need for anti-surge ports. Working on pressure differential, under normal operating conditions the valve remains closed. Pressure is fed from the lower compressor volute to the diaphragm side of the valve (example 25psi). This side of the diaphragm is also aided by an adjustable spring, while the outlet pressure of the compressor is also 25psi. These forces counteract themselves and the valve remains sealed. Once the differential pressure from the compressor wheel to the intake system lowers more than the preset adjustable spring tension, the valves opens. Pulse pressure and flow from the compressor outlet and intake system can now bypass the compressor wheel entirely through the ACRV valve into the area before the compressor wheel. The ACRV system reduces torsional loads on the compressor wheel, shaft and bearings, which can cause damage resulting in a short life for both journal and ball bearing type turbochargers.
The best way to eliminate compressor surge is by compressor wheel design!
Eclipse Turbo Systems have gone to great lengths to design each compressor wheel not to surge. The ACRV system was developed as a backup system with security in mind for zero probability of compressor surge/damage to the turbocharger.
Pressure vs Volumetric Flow vs Corrected Mass Flow.
Mass flow is a calculated total volume factoring in the air pressure, temperature and volumetric flow – leaving a corrected air density volumetric flow called “Mass Flow”. Before this correction we are using CFM (cubic meters per minute), once the “mass flow” density has been calculated it becomes a new format of Lbs/m (Pounds per Minute).
Different types of compressor surge, and its cause.
As viewed on a compressor map – Compressor pressure vs Volumetric flow vs Corrected mass flow
For this example we have stopped both lines at a point where both turbocharger #1, and turbocharger #2 have the same amount volumetric CFM (cubic feet per minute) flow. In reality turbocharger #2 will be pushing a higher volume of flow than turbocharger #1 at any point of the map.
Turbocharger #1 result is 14psi of compressor pressure.
Turbocharger #2 result is 11.5psi of compressor pressure.
Remembering both have the same volumetric flow. Only in this format do they look the same and comparable. After both turbochargers have been corrected to mass flow, the compressor efficiency (air density) results reveals a different story. Perfect efficiency is 100% as calculated by the ideal gas law. 60% efficiency means a 40% loss of energy from this perfect calculation. 65% efficiency means a 35% loss of energy from this perfect calculation. This means turbocharger #2 will have a 14% increase in air density over turbocharger #1 for a lower compressor pressure output, ensuring that compressor surge cannot be reached. Which is why the corrected mass flow of turbocharger #2 is higher. This is how important compressor efficiency is to a turbochargers!
Different types of surge, and its cause.
As viewed on a compressor map – Example 2A (deceleration surge)
With the lines now moving down the compressor map (from right to left). Deceleration compressor surge is the turbocharger’s ability to disperse the pressure and flow of the compressor, after the accelerator pedal has been released resulting in an instant removal of fuel and the diesel engine’s flow requirements. For this example, we assume the engine being discussed has no butterfly valve to close in the intake system. Here we can see turbocharger #1 shaft speed is unable to drop fast enough, along with the pressure being produced as it enters surge and “bounces” in and out of surge, down the surge line. The main cause of this surge in around 90% of cases is high turbine back pressure, and a turbine that is simply too small and not matched for the engine’s requirements. Turbine back pressure is a buildup of pressurised exhaust gasses before the turbine, that the turbine and waste gate are unable to flow. This results in a delay, from the time the accelerator’s pedal has been released, until the turbine back pressure reduces enough to NOT hold drive to the shaft and subsequently compressor wheel that’s producing the pressure and flow the engine now cannot use. Turbocharger #2 with low turbine back pressure and a correctly matched turbine is able to disperse the pressure and flow avoiding deceleration compressor surge.
Different types of surge, and its causes.
As viewed on a compressor map – Example 2B (deceleration surge)
For this example we have stopped both lines at the point of deceleration surge for turbocharger #1, and at the same volumetric flow for each turbocharger as a direct comparison. Leaving only the shaft rpm and compressor pressure to display. Lower compressor pressure and coinciding lower shaft rpm allows turbocharger #2 to avoid all deceleration surge conditions.
Note: This deceleration surge noise is not to be confused for blow off valve sounds as this example is based on a open intake diesel engine system, with NO butterfly valve body. The characteristic “choofing” sound otherwise described as “dose” does has the potential to harm bearings, compressor wheels and shafts of your turbocharger – especially at very high shaft rpm. The violent action of “compressor surge” also causes premature wearing of a journal type thrust bearing system that relies on a thin layer of oil for separation of its metal components.
Compressor Boost Pressure
Compressor boost pressure, is the measurement of restriction (or resistance) at which the volumetric flow is flowing. Higher compressor efficiency will produce the same volumetric flow, with a greater density (mass) at a lower psi (pound per square inch). Compressor wheels with a different number of blades, configurations, sizes and weights, will yield different results. Generally a turbocharger with a larger compressor wheel will produce more volumetric flow for a lower psi at a slower shaft speed, increasing the compressor’s density ratio by both volumetric and corrected mass flow. There are three types of boost gauges generally used to read this measurement of restriction.
Is a measurement of pressure measured from vacuum, and is an absolute reading. Meaning that atmoshperic pressure (14.7psi at sea level) is subtracted, resulting in a negative (vacuum) gauge start position.
This is the most common gauge used to read boost pressure. It reads from atmospheric pressure (14.7psi at sea level) and therefore only reads positive boost pressure.
Is a measurement of pressure read from vacuum, and is also a absolute reading. How ever the units of measurements are read in Bar. With 1 Bar being equal to 1 atmospheric pressure, and 0 Bar being equal to zero atmospheres (vacuum). Compressor maps are also read from this absolute reading.
Turbine Back Pressure
As viewed on a turbine flow map – Example 4A
All turbochargers have some turbine back pressure, this is usually calculated in a ratio format between the turbine pressure, divided by the compressor pressure (ie, 60psi divided by 23psi = 2.6 ( 2.6:1 turbine back pressure).
Here we can see the thin solid lines of the compressor pressure for both turbocharger #1 and turbocharger #2, with similar results in this aspect for this explanation. The thick solid lines are the corrected mass flow for each turbine for flow differences due to size and design. The dotted lines are each turbocharger’s turbine back pressure. Turbocharger #1 is able to flow 18lbs/m. Turbocharger #2 is able to flow 25lbs/m. As the turbine back pressure for turbocharger #1 increases, it reaches a critical point where the buildup of pressure stops and even reduces the turbine’s ability to flow. Where pressure increases – flow reduces.
If the turbine back pressure is severe, the exhaust valve’s ability to extract the combustion flow (as the turbine back pressure equalises with the used combustion pressure – limiting flow) is significantly reduced. In some cases, the turbine back pressure will overpower and hold exhaust valves open causing them to burn out. The design requirements of any engine is to extract the engine’s hot gases with as little impedance as possible. Anything that will cause a disruption in flow will, in turn, significantly raise the temperature throughout the engine system and result in a loss of potential power and torque. Turbocharger #2 has reasonable turbine back pressure which has a lower effect on the turbine’s ability to flow.
That’s the term to describe the engine speed required to produce enough exhaust gas flow (drive) in order to generate enough energy for the compressor flow to start creating a restriction (higher than atmospheric pressure) into the engine. This is a level of engine speed (based on fueling and thermal expansion) where the turbocharger has positive manifold pressure. Until this positive manifold pressure is reached, the turbine is acting as a restriction and causing drag, which in turn results in a slightly negative boost pressure reading in the intake manifold (worse than a naturally aspirated diesel).
On the other side of the turbocharger discussion, you’ll find turbo lag. This describes the amount of time between the point when the throttle is opened and the turbo spools up. You’re already operating the engine at a point past the boost threshold though, so that is not a factor in the turbo lag equation. Instead, turbo lag can be related to engine tuning, intake design, back pressure, and the size of the turbocharger itself. A larger turbocharger for a given engine capacity simply takes longer to spool up and deliver more boost and flow.
Boost threshold is the point at which your turbo is primed for spooling because the engine speed/exhaust gas volume is sufficient to create positive pressure. Turbo lag is the delay in time, until the turbocharger makes enough boost pressure once you’ve crossed the boost threshold (closed throttle to open throttle).
Chassis Dyno Readings
Chassis dynos are a great tool in the sense they are able to detect a change in results based on previous performance, providing the conditions are the same. “Dyno runs” are a dated way of portraying this, with many chassis dynos these days being able to mimic real world road conditions and results. There are many brands all with different functions, but generally fall into two different categories (Inertia and braked). Both are a great way of tuning a diesel under load! But like any machine, they can vary, be tricked, and give false results – so should be taken as such. From showing “derived” torque which is around 33% higher than the true result, to pre-spooling a turbocharger and manipulating the reading data. After all, the function of a chassis dyno is to try and replicate the driving experience.
Compressor pressure vs torque vs time
Here we see the constant of seconds in the bottom X axis, and compressor pressure in the left axis. Now we have added torque in newton meters at the wheels to the right for this explanation. Turbocharger #1 is making maximum boost pressure 1.5 seconds faster than turbocharger #2. Turbocharger #1 is making the same maximum torque (even with lower efficiency) as turbocharger #2 and has an advantage from 18 to 23psi. This is a product of the dyno being unable to accurately portray what is actually going on when the car is driven on the street. This maximum torque reading for turbocharger #1 has a very simple explanation – it is a result of the low speed at which the piston is moving away from the combustion source. Which is why it is much easier to make high torque numbers earlier in the rpm – fact. The down side of making maximum torque so early is the efficiency run down (diesel engines’ ability to maintain torque) also starts at a much lower rpm, and why some turbochargers tend to “fall over” after peak boost pressure/torque is made. Based on this, the only way to hold torque longer is to make significantly more torque at the same point of rpm, or to make the same or more torque later in the rpm – which is significantly harder to do. Turbocharger #1 makes more torque lower with a linear and predictable run up to its maximum torque all be it later in the rpm. The significant benefit is the amount of extra torque being produced much higher in the rpm. Which leads to two completely different driving experiences. As this is a “dyno based” result for explanation purposes, real world driving does not produce the same result with turbocharger #1 conceding to turbocharger #2 in every way. To the driver, a rush of torque that is short lived VS a sustained linear amount of torque throughout the entire engines rpm.
Turbocharger shaft Speed (rpm)
Turbocharger shaft speed is simply the rpm of the turbocharger’s rotating shaft assembly. Spool is the acceleration of the turbine shaft to its desired maximum boost pressure, or its typical maximum shaft speed to deliver that desired boost pressure for a given engine capacity and engine rpm (exhaust mass flow). Maximum shaft speeds and/or boost levels, are usually controlled by a waste gate valve, which bypasses gas flow and pressure around the turbine wheel. These shaft speeds can be above 200,000 rpm, with typical turbochargers usually reaching 140,000 to 160,000 rpm. It is usual for shaft speeds to be less, as the turbocharger size or wheel size increases.
Cooler air can hold more molecules (mass) for the same volume! Although the “volumetric” flow remains the same, the density of the air has changed – storing more “mass” for the fuel to react with. This is called “Efficiency”. The higher the air density “mass” means the more fuel can be added for increased engine power output – for the same relative combustion temperatures. Compressor efficiency is usually measured in percentage, which is a measurement based on how much mass flow from ideal (100% – Ideal Gas Law). Most turbocharger compressors usually consider 60% to be the limit before excessive heating of the air charge reduces the mass flow significantly, enough to not be considered beneficial to engine power. At 60% efficiency the compressor is considered to be in a state of surge or choke.
The compressor map supplied by OEMs is essentially a plot of pressure ratio against the mass flow. The left-hand side of the operating envelope is referred to as the ‘surge’ line. At surge, the aerodynamic elements of the compressor wheel creates reverse flow effects leading to stress load reversals in the compressor blades and a ‘coughing’ type of sound as the airflow stalls. At the other side of the map, towards the right-hand side of the envelope, the limit is one of compression efficiency of 60% depicting choke. In between the map surge and choke lines consists of a series of contours connecting points of equal compression efficiency, rather like the height contours on a map. Superimposed on the map, are the lines of constant compressor wheel speed (rpm). These efficiency contours depict the regional islands of efficiency in a corrected percent of the compressor stage.
(IE. if the compressor efficiency is 65%, then the effective loss from ideal 100% is 35%. The 65% reading pertains to the level of efficiency remaining.)
The formula for comparing efficiencies from 2 different compressor wheels consists of comparing this loss of efficiency from “ideal”, and not the difference in compressor efficiencies shown on the compressor map.
(IE. 60% efficiency = a 40% loss. 65% efficiency = a 35% loss. 40/35 = 1.14 resulting in a 14% increase of air density between these 2 compressor wheels.)
Turbocharger bearings usually fall into two different types, journal bearings (which rely on a thin layer of oil to separate the metal components controlling both radial and axial movements) and ball bearings (which use angular contact and pre-load to control axial and radial movements) and also consist of rotating balls within a cage. There are pros and cons to both systems.
* Journal bearings require a high amount of oil flow (that can cause oil to bypass the bearing piston seals). While ball bearings require an oil restrictor for very little oil flow.
* Ball bearings run cooler than journal bearings due to the very small pin point load contact area of the ball bearing system.
* Journal bearings and ball bearings at idle are equal, with ball bearings requiring 50% less load at 150,000 rpm.
(We stock and sell both bearing systems in our turbocharger range)
Intercoolers and turbocharging go hand in hand. For diesels the addition of adding an intercooler increases some of the lost air density. Intercooling alone will result in a slight change in torque, as increasing air density without increasing the fuel won’t yield any changes in performance. This is why intercooler stability is very important! If tuned while the intercooler is working in optimal conditions, a drop in air density when suffering from heat soak or lack of cooling air flow will result in over fueling and excessive heat. Intercooler piping (length, diameter and bend radius) also plays a part along with the actual size of the intercooler, where bigger is not always better. The affects of intercooling on a turbocharger with low compressor efficiency will produce higher results than intercooling a turbocharger with high compressor efficiency, as the compressor with higher efficiency already has a greater density towards ambient temperature. For the intercooler, this means there is LESS thermal load and a smaller intercooler can be used.
Waste gates are used to bypass excess engine gasses around the turbine wheel – limiting the amount of drive the turbine receives restricting both the shaft speed and compressor flow and pressure output. The waste gate (flap or valve) position that controls this criteria can be controlled by positive boost pressure, vacuum pressure and electronically. VGT (variable geometry turbochargers) rely on open vane position to vent the engine flow through the turbine at low pressures controlling shaft speed. Actuators and rods are used to control many turbocharger models on the market (Boost and vacuum) via springs with varying spring rates and rubber diaphragms that this pressure or vacuum acts upon.
Oil drain sizes and design are critical! If the oil drain has a vertical fall of over 150mm (6″) and a smooth curve into the sump with an inside diameter of 13mm (1/2″) it will be sufficient. The ideal size for this application (150mm/6″ fall) is 19mm (3/4″). If the fall is less then 30 degrees, 25mm (1″) would be ideal. Bigger is always better in these cases.
Crankcase breather size is dependent on the number of cylinders. On a four cylinder engine you can use 19mm (3/4″) in most cases. 6 cylinder engines require 25mm (1″). Catch cans are a major cause of oil smoke, as they restrict the engine breathing. If using a catch can, the hose and fitting sizes need to be doubled.
Fact Turbocharger seals do not seal engine oil pressure
Fact Turbocharger seals are there to seal crankcase pressure.
Fact Turbocharger seals are made of steel, not rubber. They are a metal piston ring design.
Fact Turbocharger seals do not fail. Bearings fail first then the seals leak.
Fact You cannot BLOW a seal, no matter what you do. If a section of piston seal is missing, it has been broken/chopped during the installation process
Fact You can easily damage the bearings, resulting in the seals leaking.
Fact Incorrectly designed oil drain = oil smoke
Fact Incorrectly designed crankcase breather = oil smoke
Fact If you have oil smoke, then you need to fix the oil drain or crankcase breather.
Fact Most ball bearing turbochargers all have oil restrictors internally. Normally 0.8 mm
Fact Most ball bearing turbochargers do not need an additional restrictor.
Fact You cannot suck the seals dry when dry sumping the oil drain.
Fact Turbocharger seals do not go hard or perish when the turbocharger is sitting on the shelf.
Q: What boost pressure can my diesel handle safely?
This question originated from, and is more directed at, the turbo petrol industry where an electric ignition system is used for timing and start up. Adding boost pressure to a petrol engine increases the combustion pressures, along with increasing compression temperatures. The fuel source (petrol) is introduced on the downward stroke of the intake allowing the fuel to mix with the air before compression. If the temperature from the subsequent compression reaches higher than the ignition temperature (flash point) of the petrol, pre-ignition is experienced which in some cases can cause engine damage. Diesel Engines, Firstly the basis of this question is flawed. As it is not boost pressure dictating the measure of increased combustion temperatures and pressures, it is the density of the air (corrected mass flow) as more density is harder to compress (cooler air can hold more molecules for the same volume). Diesel engines can handle extremely high compression with ease. It is a compression ignition and requires no spark unlike petrol engines. The timing is governed by the injector pump/ecu and the fuel source (Diesel) is injected at a point just prior to TDC (top dead center – maximum compression) to allow for a delay from when it’s injected, to the point of ignition combustion. Pre-ignition is greatly reduced, and high compression is safe on diesel engines (within reason) providing the injection timing is correct.
A turbocharger at 20psi with the correct fuel and timing will have combustion pressures of around 2000psi (far greater than many think) with the cold compression (no fuel source) of around 475psi. This combustion pressure is critical and the governing factor for reliability of any diesel. Timing has a very significant influence on peak combustion pressures and combustion temperatures, with a greater influence than boost pressure or fuel alone can (air density or diesel volume). It is not uncommon to run 30psi+ on diesel engines, but a turbocharger running a lower pressure with more air density (mass flow) yields best results and engine longevity. This combustion pressure and temperature can also be calculated back to a hp/torque figure as a maximum achievable result while still keeping the engine safe.
(It is very important to ensure the engine turbocharger’s configuration emulates all aspects of pressures, flows and efficiencies for reliability)
Along with high combustion pressures, comes high combustion peak temperatures. This is a maximum flash point ignition temperature within the cylinder, and controlled by timing and the volume of diesel being injected. These temperatures can exceed 1600deg Celsius and although these high temperatures are only for an instant, they still require adequate flow (low turbine back pressure, correctly sized turbine) to remove these temperatures. A turbocharger with high turbine back pressure retains this heat for longer resulting in higher thermal loads for the pistons, head and cooling system. This can cause significant damage to the pistons, head and valves.