How does a Turbo work? ...The role of AIT & BP vs EGT & EBP!
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10-05-2009, 01:45 PM
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How does a Turbo work? ...The role of AIT & BP vs EGT & EBP!
First I'll "paraphrase" an "incorrect" explanation of how a turbo works because I've seen versions of this explanation posted previously... a turbo is a pressure driven device that's powered by the "diesel engine" acting like an "air compressor" and forcing a flow of pressurized exhaust gas through the "turbine" just like an "air compressor" powers an "air tool" with room temperature "air pressure" and temperature has nothing to do with powering a turbo! ...and then I'll categorically state... "this isn't how a turbo works"! A turbo is a "Thermodynamic device" and its operation is governed by the differences in outlet versus inlet temperature at its compressor and turbine sections. These temperature differences can be characterized by the compressor inlet temperature and the compressor outlet pressure and by the turbine inlet temperature and the turbine inlet pressure.
I've developed equations that describe how a turbo works from a "Thermodynamic perspective" and for a given MAF=Mass Air Flow and AIT=Air Inlet Temperature at the compressor inlet and a given BP=Boost Pressure in the intake manifold, the EGT=Exhaust Gas Temperature and EBP=Exhaust Back Pressure at the turbine inlet must have a particular "EGT-EBP combination value" in order for the turbine to generate the "TSHP=Turbine Shaft HP" needed to drive the compressor hard enough to produce the observed "AIT-BP combination value"!
Unlike equations found in Thermodynamic text books which use complicated sounding parameters like "Enthalpy" my intuitive approach is easier to follow and understand. Since my equations use parameters like "FEHP=Flow Energy HP" and "IEHP=Internal Energy HP" they look different than the text book versions but my equations give the same results! How do I know my equations give the same results as those in Thermodynamic text books? Well I actually broke down and bought a new and expensive Thermodynamic text book just to make sure they do!
Why make an effort to understand from a "Thermodynamic perspective" how a turbo really works? Because turbo-diesel performance depends on the "particular EGT-EBP combination" that is producing a "given AIT-BP combination" and some "EGT-EBP combinations" are better than others and most posters seem to be looking for ways to get their EGT as low as possible for a given operating condition and this isn't the correct approach! Thermodynamics quantifies the "EGT-EBP combinations" that can produce a "given AIT-BP combination" and some of these "EGT-EBP combinations" are "undesirable" because they result in the "EBP being larger than the BP" while other "EGT-EBP combinations" are "highly desirable" because they result in the "EBP being less than or equal to the BP".
There're two reasons why having the "EBP less than or equal to the BP" is highly desirable. First during the exhaust stroke the piston pushes up against the EBP and this requires a mechanical "work energy" input from the crankshaft for each exhaust stroke and during the intake stroke the piston is pushed down by the BP and this provides a "work energy" input back to the crankshaft during each intake stroke so as long as the "EBP is less than or equal to the BP" there's no net pumping loss! When this condition for no net pumping loss is met it should then be clear that the turbo isn't receiving a "net input" of "work energy" from the crankshaft during a combined intake-exhaust stroke and that the turbo isn't being powered by the "exhaust strokes" pushing a flow of exhaust gas through the turbine like an "air compressor" powers an "air tool"!
The second reason for wanting the "EBP less than or equal to the BP" is this condition avoids "unintended EGR" by not allowing exhaust gas to backup into the intake manifold during the valve overlap interval and produce an "effective decrease" in engine displacement which is what happens if the EBP is greater than the BP! The reason I traded my trusty Ford 7.3L in for a "used" CAT 7.2L engine was to avoid having "designed in EGR"! On the other hand if the "EBP is less than the BP" then the "clearance volume" and the "swept volume" are both completely filled with oxygenated air during the intake stroke and this is like having a 5% increase in engine displacement!
For a given "AIT-BP combination" my "turbo equations" predict a "threshold value of EGT" which I call EGT* and if the turbine inlet EGT is greater than or equal to EGT* then the compressor will produce the given "AIT-BP combination" with a turbine inlet EBP that's equal to or less than the given compressor output BP! Don't think in terms of hot gas trapped in a closed container where increasing the temperature increases the pressure because that's not how a turbine works! Upstream from the spinning turbine blades a nozzle performs a "lossless conversion" of some exhaust gas "heat energy" into exhaust flow "velocity energy" and it's this high velocity flow of exhaust gas that spins the turbine blades to generate the turbine driveshaft "work energy" needed to drive the compressor hard enough to produce a given "AIT-BP combination"!
Downstream from the turbine this exhaust flow "velocity energy" has been "consumed" as turbine driveshaft "work energy" and this means the upstream amount of "heat energy" that underwent the "lossless conversion" into the exhaust flow "velocity energy" has also been "consumed" and this means the post-turbine EGT is lower than the pre-turbine EGT and the amount of "heat energy" that was "consumed" to generate the turbine driveshaft "work energy" is equal to the pre-turbine EGT minus the post-turbine EGT multiplied by the "capacity" of the exhaust gas for releasing its "heat energy" for a given decrease in gas temperature. Therefore having a higher EGT means more "heat energy" is available for "lossless conversion" into exhaust flow "velocity energy" and having more exhaust flow "velocity energy" means more driveshaft "work energy" can be generated with less EBP!
In other words from the "turbo's perspective" having a higher EGT means a lower EBP is need for the turbine to produce the output power necessary to drive the compressor hard enough to generate a given "AIT-BP combination" and if the EGT is greater than or equal to EGT* then the EBP needed will be less than or equal to the BP generated and this gives the "potential" for the "engine" to operate with no net pumping loss! However from the "engine's perspective" in order to actually operate with an "EBP that's less than or equal to the BP" the exhaust mass that's ejected from the cylinder during the "exhaust blow-down phase" of the "power stroke" must be greater (by a given amount) than the exhaust mass that's ejected from the cylinder during the BDC to TDC portion of crankshaft rotation that defines the "actual exhaust stroke". I've got "engine equations" which describe this process but first I want to concentrate on understanding how a turbo works from a Thermodynamic perspective!
Turbo discussions usually focus on "internal features" like compressor wheels and turbine A/R options. However once these "internal features" have been defined then a few basic "Thermodynamic principles" can be employed to quantify the performance of the "actual turbo package" as seen by an "external observer", namely the "driver" looking at his "gauges" as the "diesel engine" receives a boosted airflow from the "actual compressor" and returns a flow of hot exhaust gas back to the "actual turbine".
The "actual compressor" can be analyzed as a "black box compressor" that uses "ideal compression" to produce boost and the "actual turbine" can be analyzed as a "black box expander" that uses "ideal expansion" to generate the turbine output power necessary to drive the compressor and the specific "internal features" are taken into account by using their measured "efficiency parameters" which specify how closely their "actual" performance approaches "ideal black box" performance.
In the following analogy I use (.....) to give the analogous turbo parameter. As a crude analogy for illustrating a "black box" approach consider a hydroelectric plant turbine. An "external observer" can quantify this turbine's "ideal" power output by just knowing the height of the water level above the turbine's location because this height difference (temperature difference) determines the maximum possible "potential energy" ("heat energy") input to the turbine. Of course an "internal observer" is aware that this "potential energy" is converted to "velocity energy" and that it's a high velocity flow of water (exhaust gas) that spins the turbine's blades but the "external observer" can take into account any "inefficiencies" caused by these "internal features" by simply "correcting" the "actual height" of the water level to an "effective height" and then use the "potential energy" input of this "effective height" difference to calculate the turbine's "actual" power output using an "ideal" hydrodynamic "black box" model for the "actual" turbine.
In the case of a turbo an "external observer" makes corrections to the "ideal temperature differences" across the "black box compressor" and the "black box expander" using measured "efficiency data" for the specific "internal features" known to be inside the "actual turbo" and this corrects for "actual" versus "ideal" performance! The exhaust gas is like a "heat energy sponge" and mechanical devices like "turbine blades" and "piston power strokes" operate as "Thermodynamic devices" which "squeeze" some "heat energy" from the "sponge" and convert it into mechanical "work energy" output!
On the compressor side the incoming airflow is like a "heat energy sponge" and mechanical devices like "compressor wheels" and "piston compression strokes" operate as "Thermodynamic devices" which force the sponge to "soak up" some "heat energy" by converting the mechanical "work energy" input into "heat energy"! The amount of "heat energy" that's "released" or "absorbed" by the "sponge" is directly proportional to the "change" in the sponge's "temperature" as it passes through the "Thermodynamic device" and to the sponge's "capacity" for "storing" and "releasing" heat energy.
The above viewpoint is based on the "first law of Thermodynamics" which says that "heat energy" can be converted into "work energy" which is what happens in the "black box expander" and the "first law" also says that "work energy" can be converted into "heat energy" which is what happens in the "black box compressor"! Well knowing full well what low opinions some diesel forum members have for "mathematical theories" in general and for "my theories" in particular I'm compelled to point out that this "first law" isn't someone's "mathematical theory" but rather it was "discovered" during an experiment that should be of great interest to all those who participate in dyno events!
In 1843 James Joule performed a classic experiment using an insulated container of water with a paddlewheel inside that was turned by a source of calibrated mechanical "work energy" and by measuring changes in the water's temperature produced by the stirring action of the paddles he was able to demonstrate the equivalence of "heat energy" and mechanical "work energy"! So 166 years ago by using what I'd call the worlds first "water brake dynamometer" Joule determined that 1 Btu of "heat energy" was equal to 772 ft-lbf of "work energy" and today the accepted number is 1 Btu=778 ft-lbf so Joule's water brake measurements were within 1% of the correct value and that speaks well for the accuracy of water brake dynamometers in general. It's also interesting that the term hp was originated by James Watt decades before Joule's experiment so people were using hp to describe the time rate of mechanical "work energy" (as in 1 hp=33,000 ft-lbf/min) long before they even knew about the relationship between "heat energy" and mechanical "work energy"!
So it turns out that one can in fact build an engine that generates hp without knowing anything whatsoever about "heat energy" being exchangeable for a like amount of mechanical "work energy" and vice versa or about the underlying "Thermodynamic principles" involved in this "energy exchange" and conversely one is not necessarily an "expert" in "Thermodynamics" just because they've managed to build a 500 hp 7.3L power stroke! So keep this in mind when comparing explanations for "how something works" that are given by "one who owns of a 500 hp engine" versus explanations that are given by an "engineer"!
I'm not sure when my travel schedule will let me get back to this thread and start giving some "intuitive" derivations of the turbo and engine equations and since I know everyone wants to see results even without understanding how they were obtained or exactly what they mean I'll go ahead and give some but it's against my better judgment! Below are 9 operating conditions labeled #1 through #9 and they all have the same AIT=80*F, AAP=14.7 psia, TCE=Turbo Compressor Efficiency=0.7, TTE=Turbo Turbine Efficiency=0.7, ICEE=Intercooler Exchanger Efficiency=0.75, ICEAT=Intercooler Exchanger Air Temperature=90*F, well there're too many details to list them all right now so let me just point out that...
...in the table below AFR=Air Fuel Ratio, EGT* is the threshold EGT for having the "psig gauge value" of EBP=BP, dTt=delta Temp turbine is the "actual" temperature difference across the turbine, TSHP=Turbine Shaft HP and this is also the exact driveshaft hp needed by the compressor to produce the indicated BP at the given AIT=80*F, dTc=delta Temp compressor is the "actual" temperature difference across the compressor, EGTa is the "available EGT" from the engine based on a "residual heat energy" model which assumes that combustion initiates at a fixed relationship relative to TDC of the power stroke independent of RPM and BP, AFPD=Air Filter Pressure Drop psi, ICPD=Intercooler Pressure Drop psi, and the PTPD =Post Turbine Pressure Drop to the tailpipe outlet is such that PTPD=ICPD.
.......RPM.....BP....MAT....MAF...FWHP....AFR....E GT*...dTt..TSHP...dTc...EGTa....AFPD...ICPD
#1..2,000...14.3...131....28.9.....185.....26.0... 0,971...152...28.7...174...0,988....0.28....0.60
#2..2,400...14.0...131....34.0.....217.....25.5... 1,007...152...33.8...174...1,002....0.38....0.70
#3..2,800...13.7...131....38.2.....236.....25.0... 1,044...152...38.0...174...1,017....0.48....0.80
#4..2,800...17.8...140....43.1.....271.....24.5... 1,066...184...51.9...211...1,044....0.49....0.90
#5..2,800...21.0...148....46.5.....297.....24.0... 1,114...209...63.7...240...1,069....0.58....1.07
#6..2,800...24.1...154....49.6.....343.....22.0... 1,152...231...75.4...266...1,148....0.67....1.23
#7..2,800...25.9...158....51.3.....387.....20.0... 1,171...243...82.3...281...1,240....0.72....1.33
#8..2,800...28.3...163....53.5.....420.....19.0... 1,201...258...91.7...300...1,296....0.79....1.47
#9..3,000...30.1...166....57.7.....454.....18.0... 1,224...270...103....315...1,355....0.86....1.58
Conditions #1 through #3 illustrate operating at nearly a constant BP where the MAF increases in response to increasing the RPM and the FWHP increases in response to both increasing the MAF and decreasing the AFR because the MFF=MAF/AFR! The dTc values determine the "normalized (per lbm/min unit of MAF) power" needed to drive the compressor and this "normalized" compressor input power requirement only depends on the AIT and the BP which both remain constant so the dTc is also a constant 174*F independent of the MAF. The dTt values determine the "normalized (per lbm/min unit of MAF) power" generated by the turbine and since the "normalized" turbine output power requirement is equal to the "normalized" compressor input power requirement which only depends on the AIT and the BP which both remain constant the dTt is also a constant 152*F independent of the MAF! However even for a constant "normalized" power requirement the TSHP values increase from 28.7 hp to 38.0 hp in direct proportion to the increases in the MAF from 28.9 lbm/min to 38.2 lbm/min.
In general the values of dTt in the table for each of the 9 conditions are those "turbine temperature differences" that are required for the turbine to produce a "normalized" turbine output power that's equal to the "normalized" compressor input power requirement and these "turbine temperature differences" aren't a function of EGT. However if the EGT is equal to EGT* then the required "turbine temperature differences" can be achieved with an EBP that's equal to the BP. If the EGT is less than EGT* then a larger value of EBP which is larger than the BP is needed in order to generate the needed "temperature differences" across the turbine! Likewise if the EGT is larger than EGT* then a lower value of EBP which is less than the BP can generate the needed "temperature differences" across the turbine!
Conditions #3 through #8 illustrate operating at a constant RPM where the MAF increases from 38.2 lbm/min to 53.5 lbm/min in response to increasing the BP and the FWHP increases in response to both increasing the MAF and decreasing the AFR. Now the "normalized (per lbm/min unit of MAF) power" needed to drive the compressor increases as the BP increases and that means the temperature difference across the compressor dTc increases from 174*F to 300*F as the BP increases from 13.7 psig to 28.3 psig and the temperature difference across the turbine dTt must follow suite and increase from 152*F to 258*F in order for the turbine to generate the needed "normalized (per lbm/min unit of MAF) power" to drive the compressor and the TSHP increases from 38.0 hp to 91.7 hp due to both increasing the BP and due to increasing the MAF! Condition #9 at an AFR=18 is beginning to blow some black smoke and the TSHP increases to 103 hp due to increases in both the BP and the MAF!
At condition #6 the EGTa is approximately equal to the EGT* and for conditions #7 through #9 the EGTa is progressively larger than the EGT* and this "crossover EGT effect" for having the "EBP less than the BP" has been documented on test engines! So it appears to me that a good tuning strategy is to adjust the injection timing and pulse duration to get the EGTa higher than the EGT* at the lowest possible BP value and that "tweaking" the tune to "lower" the EGTa for a given BP value isn't the correct approach because that winds up increasing the EBP! When dealing with "Thermodynamic devices" such as "diesels" and "turbos" which derive their output powers from a supply of "heat energy" then as long as no design limits are being exceeded "temperature" isn't an "enemy" to be defeated but rather an "ally" to be embraced!
The reason the temperature difference across the turbine dTt is less than the corresponding temperature difference across the compressor dTc is that for a given temperature difference the exhaust gas releases more "heat energy" than is absorbed by the corresponding airflow passing through the compressor! Also as I've said many times the reason why installing an "open element" air filter directly onto the compressor inlet is a bad idea is because under the hood air temperature is higher than ambient and for a given BP the power required to drive the compressor is directly proportional to the "absolute value" of the AIT. The compressor needs to "work harder" to "compress hotter air" and since this "harder work" happens upstream from the intercooler having the best intercooler or not even having an intercooler makes no difference concerning this "penalty aspect" of "open element" air filters!
The above description is a "simplified" one because for a given BP the "temperature difference" across the compressor actually depends on the AIT and the CPR=Compressor Pressure Ratio and the corresponding "temperature difference" across the turbine actually depends on the EGT and the TPR=Turbine Pressure Ratio and the EGT* depends on both the CPR and TPR as well as on the AFR, TCE, TTE, and the AIT, but those details will be covered in future installments.
I've developed equations that describe how a turbo works from a "Thermodynamic perspective" and for a given MAF=Mass Air Flow and AIT=Air Inlet Temperature at the compressor inlet and a given BP=Boost Pressure in the intake manifold, the EGT=Exhaust Gas Temperature and EBP=Exhaust Back Pressure at the turbine inlet must have a particular "EGT-EBP combination value" in order for the turbine to generate the "TSHP=Turbine Shaft HP" needed to drive the compressor hard enough to produce the observed "AIT-BP combination value"!
Unlike equations found in Thermodynamic text books which use complicated sounding parameters like "Enthalpy" my intuitive approach is easier to follow and understand. Since my equations use parameters like "FEHP=Flow Energy HP" and "IEHP=Internal Energy HP" they look different than the text book versions but my equations give the same results! How do I know my equations give the same results as those in Thermodynamic text books? Well I actually broke down and bought a new and expensive Thermodynamic text book just to make sure they do!
Why make an effort to understand from a "Thermodynamic perspective" how a turbo really works? Because turbo-diesel performance depends on the "particular EGT-EBP combination" that is producing a "given AIT-BP combination" and some "EGT-EBP combinations" are better than others and most posters seem to be looking for ways to get their EGT as low as possible for a given operating condition and this isn't the correct approach! Thermodynamics quantifies the "EGT-EBP combinations" that can produce a "given AIT-BP combination" and some of these "EGT-EBP combinations" are "undesirable" because they result in the "EBP being larger than the BP" while other "EGT-EBP combinations" are "highly desirable" because they result in the "EBP being less than or equal to the BP".
There're two reasons why having the "EBP less than or equal to the BP" is highly desirable. First during the exhaust stroke the piston pushes up against the EBP and this requires a mechanical "work energy" input from the crankshaft for each exhaust stroke and during the intake stroke the piston is pushed down by the BP and this provides a "work energy" input back to the crankshaft during each intake stroke so as long as the "EBP is less than or equal to the BP" there's no net pumping loss! When this condition for no net pumping loss is met it should then be clear that the turbo isn't receiving a "net input" of "work energy" from the crankshaft during a combined intake-exhaust stroke and that the turbo isn't being powered by the "exhaust strokes" pushing a flow of exhaust gas through the turbine like an "air compressor" powers an "air tool"!
The second reason for wanting the "EBP less than or equal to the BP" is this condition avoids "unintended EGR" by not allowing exhaust gas to backup into the intake manifold during the valve overlap interval and produce an "effective decrease" in engine displacement which is what happens if the EBP is greater than the BP! The reason I traded my trusty Ford 7.3L in for a "used" CAT 7.2L engine was to avoid having "designed in EGR"! On the other hand if the "EBP is less than the BP" then the "clearance volume" and the "swept volume" are both completely filled with oxygenated air during the intake stroke and this is like having a 5% increase in engine displacement!
For a given "AIT-BP combination" my "turbo equations" predict a "threshold value of EGT" which I call EGT* and if the turbine inlet EGT is greater than or equal to EGT* then the compressor will produce the given "AIT-BP combination" with a turbine inlet EBP that's equal to or less than the given compressor output BP! Don't think in terms of hot gas trapped in a closed container where increasing the temperature increases the pressure because that's not how a turbine works! Upstream from the spinning turbine blades a nozzle performs a "lossless conversion" of some exhaust gas "heat energy" into exhaust flow "velocity energy" and it's this high velocity flow of exhaust gas that spins the turbine blades to generate the turbine driveshaft "work energy" needed to drive the compressor hard enough to produce a given "AIT-BP combination"!
Downstream from the turbine this exhaust flow "velocity energy" has been "consumed" as turbine driveshaft "work energy" and this means the upstream amount of "heat energy" that underwent the "lossless conversion" into the exhaust flow "velocity energy" has also been "consumed" and this means the post-turbine EGT is lower than the pre-turbine EGT and the amount of "heat energy" that was "consumed" to generate the turbine driveshaft "work energy" is equal to the pre-turbine EGT minus the post-turbine EGT multiplied by the "capacity" of the exhaust gas for releasing its "heat energy" for a given decrease in gas temperature. Therefore having a higher EGT means more "heat energy" is available for "lossless conversion" into exhaust flow "velocity energy" and having more exhaust flow "velocity energy" means more driveshaft "work energy" can be generated with less EBP!
In other words from the "turbo's perspective" having a higher EGT means a lower EBP is need for the turbine to produce the output power necessary to drive the compressor hard enough to generate a given "AIT-BP combination" and if the EGT is greater than or equal to EGT* then the EBP needed will be less than or equal to the BP generated and this gives the "potential" for the "engine" to operate with no net pumping loss! However from the "engine's perspective" in order to actually operate with an "EBP that's less than or equal to the BP" the exhaust mass that's ejected from the cylinder during the "exhaust blow-down phase" of the "power stroke" must be greater (by a given amount) than the exhaust mass that's ejected from the cylinder during the BDC to TDC portion of crankshaft rotation that defines the "actual exhaust stroke". I've got "engine equations" which describe this process but first I want to concentrate on understanding how a turbo works from a Thermodynamic perspective!
Turbo discussions usually focus on "internal features" like compressor wheels and turbine A/R options. However once these "internal features" have been defined then a few basic "Thermodynamic principles" can be employed to quantify the performance of the "actual turbo package" as seen by an "external observer", namely the "driver" looking at his "gauges" as the "diesel engine" receives a boosted airflow from the "actual compressor" and returns a flow of hot exhaust gas back to the "actual turbine".
The "actual compressor" can be analyzed as a "black box compressor" that uses "ideal compression" to produce boost and the "actual turbine" can be analyzed as a "black box expander" that uses "ideal expansion" to generate the turbine output power necessary to drive the compressor and the specific "internal features" are taken into account by using their measured "efficiency parameters" which specify how closely their "actual" performance approaches "ideal black box" performance.
In the following analogy I use (.....) to give the analogous turbo parameter. As a crude analogy for illustrating a "black box" approach consider a hydroelectric plant turbine. An "external observer" can quantify this turbine's "ideal" power output by just knowing the height of the water level above the turbine's location because this height difference (temperature difference) determines the maximum possible "potential energy" ("heat energy") input to the turbine. Of course an "internal observer" is aware that this "potential energy" is converted to "velocity energy" and that it's a high velocity flow of water (exhaust gas) that spins the turbine's blades but the "external observer" can take into account any "inefficiencies" caused by these "internal features" by simply "correcting" the "actual height" of the water level to an "effective height" and then use the "potential energy" input of this "effective height" difference to calculate the turbine's "actual" power output using an "ideal" hydrodynamic "black box" model for the "actual" turbine.
In the case of a turbo an "external observer" makes corrections to the "ideal temperature differences" across the "black box compressor" and the "black box expander" using measured "efficiency data" for the specific "internal features" known to be inside the "actual turbo" and this corrects for "actual" versus "ideal" performance! The exhaust gas is like a "heat energy sponge" and mechanical devices like "turbine blades" and "piston power strokes" operate as "Thermodynamic devices" which "squeeze" some "heat energy" from the "sponge" and convert it into mechanical "work energy" output!
On the compressor side the incoming airflow is like a "heat energy sponge" and mechanical devices like "compressor wheels" and "piston compression strokes" operate as "Thermodynamic devices" which force the sponge to "soak up" some "heat energy" by converting the mechanical "work energy" input into "heat energy"! The amount of "heat energy" that's "released" or "absorbed" by the "sponge" is directly proportional to the "change" in the sponge's "temperature" as it passes through the "Thermodynamic device" and to the sponge's "capacity" for "storing" and "releasing" heat energy.
The above viewpoint is based on the "first law of Thermodynamics" which says that "heat energy" can be converted into "work energy" which is what happens in the "black box expander" and the "first law" also says that "work energy" can be converted into "heat energy" which is what happens in the "black box compressor"! Well knowing full well what low opinions some diesel forum members have for "mathematical theories" in general and for "my theories" in particular I'm compelled to point out that this "first law" isn't someone's "mathematical theory" but rather it was "discovered" during an experiment that should be of great interest to all those who participate in dyno events!
In 1843 James Joule performed a classic experiment using an insulated container of water with a paddlewheel inside that was turned by a source of calibrated mechanical "work energy" and by measuring changes in the water's temperature produced by the stirring action of the paddles he was able to demonstrate the equivalence of "heat energy" and mechanical "work energy"! So 166 years ago by using what I'd call the worlds first "water brake dynamometer" Joule determined that 1 Btu of "heat energy" was equal to 772 ft-lbf of "work energy" and today the accepted number is 1 Btu=778 ft-lbf so Joule's water brake measurements were within 1% of the correct value and that speaks well for the accuracy of water brake dynamometers in general. It's also interesting that the term hp was originated by James Watt decades before Joule's experiment so people were using hp to describe the time rate of mechanical "work energy" (as in 1 hp=33,000 ft-lbf/min) long before they even knew about the relationship between "heat energy" and mechanical "work energy"!
So it turns out that one can in fact build an engine that generates hp without knowing anything whatsoever about "heat energy" being exchangeable for a like amount of mechanical "work energy" and vice versa or about the underlying "Thermodynamic principles" involved in this "energy exchange" and conversely one is not necessarily an "expert" in "Thermodynamics" just because they've managed to build a 500 hp 7.3L power stroke! So keep this in mind when comparing explanations for "how something works" that are given by "one who owns of a 500 hp engine" versus explanations that are given by an "engineer"!
I'm not sure when my travel schedule will let me get back to this thread and start giving some "intuitive" derivations of the turbo and engine equations and since I know everyone wants to see results even without understanding how they were obtained or exactly what they mean I'll go ahead and give some but it's against my better judgment! Below are 9 operating conditions labeled #1 through #9 and they all have the same AIT=80*F, AAP=14.7 psia, TCE=Turbo Compressor Efficiency=0.7, TTE=Turbo Turbine Efficiency=0.7, ICEE=Intercooler Exchanger Efficiency=0.75, ICEAT=Intercooler Exchanger Air Temperature=90*F, well there're too many details to list them all right now so let me just point out that...
...in the table below AFR=Air Fuel Ratio, EGT* is the threshold EGT for having the "psig gauge value" of EBP=BP, dTt=delta Temp turbine is the "actual" temperature difference across the turbine, TSHP=Turbine Shaft HP and this is also the exact driveshaft hp needed by the compressor to produce the indicated BP at the given AIT=80*F, dTc=delta Temp compressor is the "actual" temperature difference across the compressor, EGTa is the "available EGT" from the engine based on a "residual heat energy" model which assumes that combustion initiates at a fixed relationship relative to TDC of the power stroke independent of RPM and BP, AFPD=Air Filter Pressure Drop psi, ICPD=Intercooler Pressure Drop psi, and the PTPD =Post Turbine Pressure Drop to the tailpipe outlet is such that PTPD=ICPD.
.......RPM.....BP....MAT....MAF...FWHP....AFR....E GT*...dTt..TSHP...dTc...EGTa....AFPD...ICPD
#1..2,000...14.3...131....28.9.....185.....26.0... 0,971...152...28.7...174...0,988....0.28....0.60
#2..2,400...14.0...131....34.0.....217.....25.5... 1,007...152...33.8...174...1,002....0.38....0.70
#3..2,800...13.7...131....38.2.....236.....25.0... 1,044...152...38.0...174...1,017....0.48....0.80
#4..2,800...17.8...140....43.1.....271.....24.5... 1,066...184...51.9...211...1,044....0.49....0.90
#5..2,800...21.0...148....46.5.....297.....24.0... 1,114...209...63.7...240...1,069....0.58....1.07
#6..2,800...24.1...154....49.6.....343.....22.0... 1,152...231...75.4...266...1,148....0.67....1.23
#7..2,800...25.9...158....51.3.....387.....20.0... 1,171...243...82.3...281...1,240....0.72....1.33
#8..2,800...28.3...163....53.5.....420.....19.0... 1,201...258...91.7...300...1,296....0.79....1.47
#9..3,000...30.1...166....57.7.....454.....18.0... 1,224...270...103....315...1,355....0.86....1.58
Conditions #1 through #3 illustrate operating at nearly a constant BP where the MAF increases in response to increasing the RPM and the FWHP increases in response to both increasing the MAF and decreasing the AFR because the MFF=MAF/AFR! The dTc values determine the "normalized (per lbm/min unit of MAF) power" needed to drive the compressor and this "normalized" compressor input power requirement only depends on the AIT and the BP which both remain constant so the dTc is also a constant 174*F independent of the MAF. The dTt values determine the "normalized (per lbm/min unit of MAF) power" generated by the turbine and since the "normalized" turbine output power requirement is equal to the "normalized" compressor input power requirement which only depends on the AIT and the BP which both remain constant the dTt is also a constant 152*F independent of the MAF! However even for a constant "normalized" power requirement the TSHP values increase from 28.7 hp to 38.0 hp in direct proportion to the increases in the MAF from 28.9 lbm/min to 38.2 lbm/min.
In general the values of dTt in the table for each of the 9 conditions are those "turbine temperature differences" that are required for the turbine to produce a "normalized" turbine output power that's equal to the "normalized" compressor input power requirement and these "turbine temperature differences" aren't a function of EGT. However if the EGT is equal to EGT* then the required "turbine temperature differences" can be achieved with an EBP that's equal to the BP. If the EGT is less than EGT* then a larger value of EBP which is larger than the BP is needed in order to generate the needed "temperature differences" across the turbine! Likewise if the EGT is larger than EGT* then a lower value of EBP which is less than the BP can generate the needed "temperature differences" across the turbine!
Conditions #3 through #8 illustrate operating at a constant RPM where the MAF increases from 38.2 lbm/min to 53.5 lbm/min in response to increasing the BP and the FWHP increases in response to both increasing the MAF and decreasing the AFR. Now the "normalized (per lbm/min unit of MAF) power" needed to drive the compressor increases as the BP increases and that means the temperature difference across the compressor dTc increases from 174*F to 300*F as the BP increases from 13.7 psig to 28.3 psig and the temperature difference across the turbine dTt must follow suite and increase from 152*F to 258*F in order for the turbine to generate the needed "normalized (per lbm/min unit of MAF) power" to drive the compressor and the TSHP increases from 38.0 hp to 91.7 hp due to both increasing the BP and due to increasing the MAF! Condition #9 at an AFR=18 is beginning to blow some black smoke and the TSHP increases to 103 hp due to increases in both the BP and the MAF!
At condition #6 the EGTa is approximately equal to the EGT* and for conditions #7 through #9 the EGTa is progressively larger than the EGT* and this "crossover EGT effect" for having the "EBP less than the BP" has been documented on test engines! So it appears to me that a good tuning strategy is to adjust the injection timing and pulse duration to get the EGTa higher than the EGT* at the lowest possible BP value and that "tweaking" the tune to "lower" the EGTa for a given BP value isn't the correct approach because that winds up increasing the EBP! When dealing with "Thermodynamic devices" such as "diesels" and "turbos" which derive their output powers from a supply of "heat energy" then as long as no design limits are being exceeded "temperature" isn't an "enemy" to be defeated but rather an "ally" to be embraced!
The reason the temperature difference across the turbine dTt is less than the corresponding temperature difference across the compressor dTc is that for a given temperature difference the exhaust gas releases more "heat energy" than is absorbed by the corresponding airflow passing through the compressor! Also as I've said many times the reason why installing an "open element" air filter directly onto the compressor inlet is a bad idea is because under the hood air temperature is higher than ambient and for a given BP the power required to drive the compressor is directly proportional to the "absolute value" of the AIT. The compressor needs to "work harder" to "compress hotter air" and since this "harder work" happens upstream from the intercooler having the best intercooler or not even having an intercooler makes no difference concerning this "penalty aspect" of "open element" air filters!
The above description is a "simplified" one because for a given BP the "temperature difference" across the compressor actually depends on the AIT and the CPR=Compressor Pressure Ratio and the corresponding "temperature difference" across the turbine actually depends on the EGT and the TPR=Turbine Pressure Ratio and the EGT* depends on both the CPR and TPR as well as on the AFR, TCE, TTE, and the AIT, but those details will be covered in future installments.
#2
10-07-2009, 09:56 PM
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...Thermodynamics quantifies the "EGT-EBP combinations" that can produce a "given AIT-BP combination" and some of these "EGT-EBP combinations" are "undesirable" because they result in the "EBP being larger than the BP" while other "EGT-EBP combinations" are "highly desirable" because they result in the "EBP being less than or equal to the BP"...
Table#1 "AIT=80*F & BP=17.8 psig combination"
#4,1..EGT=0,917*F & EBP=20.8 psig & EBP=BP+3
#4,2..EGT=0,962*F & EBP=19.8 psig & EBP=BP+2
#4,3..EGT=1,011*F & EBP=18.8 psig & EBP=BP+1
#4,4..EGT=1,066*F & EBP=17.8 psig & EBP=BP & EGT=EGT*
#4,5..EGT=1,128*F & EBP=16.8 psig & EBP=BP-1
#4,6..EGT=1,198*F & EBP=15.8 psig & EBP=BP-2
#4,7..EGT=1,277*F & EBP=14.8 psig & EBP=BP-3
Table#2 "AIT=80*F & BP=21.0 psig combination"
#5,1..EGT=0,987*F & EBP=24.0 psig & EBP=BP+3
#5,2..EGT=1,025*F & EBP=23.0 psig & EBP=BP+2
#5,3..EGT=1,067*F & EBP=22.0 psig & EBP=BP+1
#5,4..EGT=1,114*F & EBP=21.0 psig & EBP=BP & EGT=EGT*
#5,5..EGT=1,165*F & EBP=20.0 psig & EBP=BP-1
#5,6..EGT=1,221*F & EBP=19.0 psig & EBP=BP-2
#5,7..EGT=1,284*F & EBP=18.0 psig & EBP=BP-3
When I had a 70 hp chip installed in my old 7.3L I took measurements of BP when the wastegate was forced to remain fully open and I saw a maximum BP of 12 psig to 14 psig and when the wastegate was forced to remain fully closed I saw a maximum BP of 28+ psig. Based on those measurements I think any of the 14 "EGT-EBP combinations" in the above two tables can be accommodated by a "wastegate position" somewhere between "fully open" and "fully closed"!
I think the wastegate position that accommodates the #4,5 "EGT-EBP combination" in Table#1 would be the best choice for towing up a long steady grade at part throttle with a BP=17.8 psig and if I got on the throttle a little harder and wanted to tow the grade with a BP=21.0 psig I think the wastegate position that accommodates the #5,5 "EGT-EBP combination" in Table#2 would be the best choice. Both choices give a "completely safe EGT" and both choices give a "EBP=BP-1" which implies no EGR and no net pumping loss and in fact a slight net "pumping gain"! However if you're primarily interested in winning drag races these probably aren't the best wastegate positions to employ when the BP momentarily passes through 17.8 psig and 21.0 psig as the BP is increasing to higher values!
Of course in the above examples I'm only talking about "Thermodynamic possibilities" and to actually "realize" a particular "possibility" the engine needs to produce the right amount of "waste heat energy" and the "wastegate controller" needs to position the wastegate at the correct degree of opening and hold it there! To achieve my particular choices of "EGT-EBP combinations" where the "EBP=BP-1" the engine needs to produce more "waste heat energy" than the bare minimum need for the desired values of BP when the wastegate is fully closed so that the "excess" exhaust flow "heat energy" can be "bypassed" through the "wastegate" and the desired values of BP can still be achieved.
The above consideration is where a VGT turbo "shines" because it has in effect a "variable area nozzle" upstream from the turbine so that all of the exhaust flow "heat energy" passes through the turbine and none of it is ever "wasted". This allows a "VGT controller" to use all the available "waste heat energy" from the engine and employ the largest effective nozzle area to get the best "EGT-EBP combination" for producing a given BP with a minimum EBP!
I'm not sure what options exist if any for reprogramming the stock wastegate control strategy but the ultimate would be an independent controller similar to the one that's available for controlling the automatic transmission. Such a controller could be programmed for specific user desires and a controller that used inputs from both the MAP sensor and the EGT probe would allow for some interesting control strategies to be considered!
#5
10-07-2009, 10:39 PM
#6
10-07-2009, 11:22 PM
#7
10-07-2009, 11:37 PM
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#8
10-07-2009, 11:47 PM
#9
10-07-2009, 11:57 PM
I say, keep the info coming! Timmy
#10
10-08-2009, 06:09 AM
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#12
10-08-2009, 07:18 AM
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#13
10-08-2009, 07:36 AM
#14
10-08-2009, 08:05 AM
So I guess if one has a turbo with no waste gate (like the van turbo or 38r with 1.15 housing), we're just hosed? I need to look at my EBP again. I was thinking I was close to 1:1 but now with the 235cc hybrids, I'm sure I'm getting some EGR action happening.
If I understand this, you'll still get more power even with a positive EBP-BP number, just at a cost of efficiency and possible turbo bearing failure if it gets too high. Am I tracking correctly here?
If I understand this, you'll still get more power even with a positive EBP-BP number, just at a cost of efficiency and possible turbo bearing failure if it gets too high. Am I tracking correctly here?
#15
10-08-2009, 10:05 AM
