Tooth jump = bump?

[Bob4x4]

Picture it like having more mechanical timing in a distributed

 

[ENafziger]

1. Would someone please chime in regarding the validity of this claim?
2. Is the spacer really superior (to twisting the pump)?
3. Does the M&H spacer result in too much advance for a bone stock truck?

Twisting the pump results in advancing timing across the board by a set amount (for example, timing is 3° advanced at every point throughout the engine load/speed range).

The M&H spacer only adds to the range of dynamic timing, accomplished by the KSB. In other words, the base timing at idle, low speed/load stays the same. As the engine ramps up in speed/load, the amount of timing advance that the KSB can give is increased by the M&H spacer. This is probably more important if you also increase the available speed/load the engine can put out (such as 3200 rpm spring, etc). This could be advantangeous, because advancing the base timing too much will have a detrimental effect on low end torque and/or turbo lag time.

--Eric

 

[James1]

Twisting the pump results in advancing timing across the board by a set amount (for example, timing is 3° advanced at every point throughout the engine load/speed range). The M&H spacer only adds to the range of dynamic timing, accomplished by the KSB. In other words, the base timing at idle, low speed/load stays the same. As the engine ramps up in speed/load, the amount of timing advance that the KSB can give is increased by the M&H spacer. This is probably more important if you also increase the available speed/load the engine can put out (such as 3200 rpm spring, etc). This could be advantangeous, because advancing the base timing too much will have a detrimental effect on low end torque and/or turbo lag time.

Thanks Eric, that's helpful. Sorry to be a noodge, but I now have more questions.



I have no plans to turn up my pump fuelwise, but I do tow periodically and would like to tow at highway speeds with OverDrive turned off. SO, I have bought the "366" governor spring (3200 rpm) from the local Bosch shop and plan to get it installed without making other adjustments to the pump.



Based on what you said, the M&H spacer would have essentially no effect on my truck at idle. With the background I described above, would the M&H spacer give me too much timing advance at highways speeds while towing (bone stock)?



And again, thank you.



Jim

 

[Mhannink]

Do you mean the physical hydraulic delay in causing the injector to pop-off?



I know that SOI isn't the "tell all", but it is a (somewhat) useful thing for me to roll around in my head and reference. I see you have a 3rd gen like myself, and are aware that rail pressure, injector size, etc, etc all have an effect on what the "timing" actually is.



I guess my confusion, is trying to reconcile what I do at work, with what I do at play. At work, our engines are highly instrumented, with in-cylinder pressure transducers, high speed data acquisition (one fifth of a degree CA resolution), with complicated and very expensive real-time combustion feedback (heat release plots, pressure traces, 10% mass fraction burned, 50% MFB, 90% MFB, IMEP, BMEP, NMEP, peak pressure, pressure rise rates, etc). The data is at such high resolution and transfer rate, that even with a large buffer, it is nearly impossible to save the data to the outer portion of the hardrive's circumference, as the data can't be written that fast. And then, I jump in my truck to head home, and pick 0, 1, 2, or 3 for timing on Smarty, and have no idea what any of them do, how aggressive Marco went on his "advanced" timing settings, etc, etc. Or I jump in the '91. 5, and wonder how much bsfc improvement I could get by advancing timing, but struggle with the apprehension of sacrificing longevity due to raising peak cylinder pressure too high.



This is probably not productive to talk about such things, but I've been somewhat soft and wordy and rambling today for some reason. Which brings me to say, I think I'm going to sell my Smarty, TST, and TS MP-8 and get EFI Live for the '06. I still won't have a clue what is going on in the cylinder, but at least timing will be what I tell it to be... or something like that



--Eric

enafziger, im wondering if you might be able to help me out. as i mentioned i pulled out all my school books and engine theory and operation, as well as all of the material i have collected over the years pertaining to engines both gas a diesel. my collection is missing a few crucial things pertaining to what i am trying to do. for starters i have a handout from school that grapghicaly displayes the phases of combustion, its a great reference material with some helpful facts but its for visualization only, so far ive been able to grapgh the static cylinder pressure curve with some acuracy. however i cant find any information pertaining to cylinder pressure rise when fuel is delivered. i know its going to vary based on volume and duration, which is the information im looking for. i need things like rate of flame propagation, and rate of exapansion, also any conditions that would change those rates. with that i cant probably figure out what the cylinder pressure curve looks like for any given set of conditions.



the thread i quoted above is exactly the information i desire.



now i dont know if your giving me that information would be an issue or not, and as i dont want you to get in trouble with your employer dont give me any information that they dont want released, i dont want you to get in trouble on my behalf.



on another note, another tid bit of info im having trouble finding is the formula for air flow in a forced induction engine, i am working on ordering a book on turbocharging that seems to go pretty deep into the engineering aspect of it, the formula might be in there but i dont know, also i need the formula for volumetric effiecincey for forced induction engines and the fomula for thermal effiecency. really any information that is relivant to these sorts of aspects of an engine would be great.



like i said i am working on calculating and plotting as much of this as i can and soon i hope to be able to post this information for all of us. i will also post all information, rules of thumb, formulas, and theories that determine or guide all of these things to help it all make sense.

 

[ENafziger]

Thanks Eric, that's helpful. Sorry to be a noodge, but I now have more questions.

I have no plans to turn up my pump fuelwise, but I do tow periodically and would like to tow at highway speeds with OverDrive turned off. SO, I have bought the "366" governor spring (3200 rpm) from the local Bosch shop and plan to get it installed without making other adjustments to the pump.

Based on what you said, the M&H spacer would have essentially no effect on my truck at idle. With the background I described above, would the M&H spacer give me too much timing advance at highways speeds while towing (bone stock)?

I don't mind questions! That's how I learn...

There's not a ton of definitive information on the M&H spacers. However, the KSB is designed to add dynamic timing advance up to the VEs 2500 rpm defueling limit; thus, it only seems logical to me, that for best efficiency it would benefit to allow the KSB to keep advancing timing with rpm up to 3200 rpm... which is exactly what the M&H spacer does (to my understanding). I think the M&H spacer and 3200 rpm spring would go great together on a stock truck. I know you're not wanting to do alot of mods, but have you talked to them about the M1 or M2 pin and how it would benefit you?

--Eric

 

[ENafziger]

enafziger, im wondering if you might be able to help me out.

I'd be glad to, but you may have to be patient.

I have a few plots I think I can share, but I can't figure out how to get them to show up here... copying pasting doesn't work. I can't think of anything right now other than trying to make a jpeg image out of the plots, uploading to a photo imaging site, and then posting here as a picture... which I can do, but I don't have time for right now.

Garrett's site has some good basic info on it, particularly on turbo charging/engine efficiency. Some general diesel stuff at: TurboByGarrett.com - Turbo Tech101

And here, for volumetric efficiency stuff... TurboTech 103 Expert really has the most useful info, particularly toward the bottom of the page. TurboByGarrett.com - Tech Center

Some common numbers that come to mind for a DI compression ignition engine are: cylinder pressure at start of injection ~ 400 psi, maximum allowable cylinder pressure for sustained load ~ 160 bar or 2,300 psi, peak pressure rise rate 10 bar/CA degree, maximum piston speed 12 m/s, etc.

The concept of flame propagation speed doesn't really apply to diesel combustion. In the SI (spark ignition) world, it's useful because you can often assume that the "flame" will start at the spark plug and propagate out (although sometimes the exhaust valve can be hot enough to ignite the fuel/air mixture slightly before the spark). In diesel combustion, the mixture is always lean, and mostly very lean. Since the combustion mode is via compression ignition, you end up with lots of localized ignition spots within the cylinder that start the combustion process. These spots are dependant upon so many things that it's hard to do much without advanced combustion modeling software. So, the diesel equivalent, so to speak, to the SI flame propagation speed, is Mass Fraction Burned, or MFB. We use DCat some from Drivven for real-time heat release and combustion feedback. There are others as well, AVL etc, but mostly they will indicate 10%, 50%, and 90% MFB. Some people use 5% as the start of combustion indicator, and some 1% (although that's so noisy that results are not consistant). Thus, I suppose you could compare CA degrees between MFB10 and MFB90 as "burn time". There are so many things that affect burn time, that it can vary drastically... fuel properties, engine design, piston design, injection pressure, cam design, egr rate, swirl and tumble, etc, etc. As a general rule of thumb, injection timing needs to be advanced 8 - 10° for low cetane fuels to have the same combustion phasing (MFB50) as a high cetane fuel (low cetane meaning 40 and high meaning 55 or so).

There are so many variables in the operation of an engine, that even with advanced simulation, it is difficult to predict things... and with everything, there is a limit of what is "good". For example, for optimum bsfc (brake specific fuel consumption), you would ideally want all heat release to occur instantaneously right after TDC; however, this is not achievable do to chemical kinetics and physical nature of things... but even if it were possible, allowable peak pressure rise rate (~10 bar/° as stated above) would cause you to have to spread out the burn. The 10 bar/° number has to do with head gasket integrity, connecting rod strength, rod bearings, etc.

Moreover, it is generally assumed that better atomization, resulting from higher injection pressure, increases combustion efficiency. This is generally true. However, you will find a point that the parasitic loss of pumping the fuel to higher pressures is greater than the increase in efficiency from better atomization... and as such, bsfc is actually lower.

--Eric

 

[Mhannink]

I'd be glad to, but you may have to be patient.

i can be paitent, i have many other things going on as well and can come back to this when you have the data for me. whenever its conveinent for you is fine.



I have a few plots I think I can share, but I can't figure out how to get them to show up here... copying pasting doesn't work. I can't think of anything right now other than trying to make a jpeg image out of the plots, uploading to a photo imaging site, and then posting here as a picture... which I can do, but I don't have time for right now.



Garrett's site has some good basic info on it, particularly on turbo charging/engine efficiency. Some general diesel stuff at: TurboByGarrett.com - Turbo Tech101



And here, for volumetric efficiency stuff... TurboTech 103 Expert really has the most useful info, particularly toward the bottom of the page. TurboByGarrett.com - Tech Center

i will read into these and see what i garner out of them and apply it to my work.



Some common numbers that come to mind for a DI compression ignition engine are: cylinder pressure at start of injection ~ 400 psi, maximum allowable cylinder pressure for sustained load ~ 160 bar or 2,300 psi, peak pressure rise rate 10 bar/CA degree, maximum piston speed 12 m/s, etc.



400psi thats at full load? based on compression ratio, boost pressure, and ambeint air pressure? the pressure rise rate, thats max allowable/best case? is there a rule of thumb rate? or a formula that could factor in fuel volume, cylinder pressure, temps, etc. to determine the rate of expansion/heat release, for a given set or paramater i. e. half load, low rpm, cruising, etc. piston speed, 12 meters? per second also is that average max speed or peak max speed?

also is there a ratio for the expansion, like that for black powder? say an oz. of fuel will expand to a certain amount, and at a rate in time?



The concept of flame propagation speed doesn't really apply to diesel combustion. In the SI (spark ignition) world, it's useful because you can often assume that the "flame" will start at the spark plug and propagate out (although sometimes the exhaust valve can be hot enough to ignite the fuel/air mixture slightly before the spark). In diesel combustion, the mixture is always lean, and mostly very lean. Since the combustion mode is via compression ignition, you end up with lots of localized ignition spots within the cylinder that start the combustion process. These spots are dependant upon so many things that it's hard to do much without advanced combustion modeling software. So, the diesel equivalent, so to speak, to the SI flame propagation speed, is Mass Fraction Burned, or MFB. We use DCat some from Drivven for real-time heat release and combustion feedback. There are others as well, AVL etc, but mostly they will indicate 10%, 50%, and 90% MFB. Some people use 5% as the start of combustion indicator, and some 1% (although that's so noisy that results are not consistant). Thus, I suppose you could compare CA degrees between MFB10 and MFB90 as "burn time". There are so many things that affect burn time, that it can vary drastically... fuel properties, engine design, piston design, injection pressure, cam design, egr rate, swirl and tumble, etc, etc. As a general rule of thumb, injection timing needs to be advanced 8 - 10° for low cetane fuels to have the same combustion phasing (MFB50) as a high cetane fuel (low cetane meaning 40 and high meaning 55 or so).



the reason i asked about flame propogation is i came across a company that is doing work on the physical properties of fuel using an additive, from what i garnished the additive has an affect on flame propagation, heres the link Fuel Additives: Fuel Treatment - Viscon if you have time i would be interested to hear your feed back on this, i rarely buy into a companies advertising.



There are so many variables in the operation of an engine, that even with advanced simulation, it is difficult to predict things... and with everything, there is a limit of what is "good". For example, for optimum bsfc (brake specific fuel consumption), you would ideally want all heat release to occur instantaneously right after TDC; however, this is not achievable do to chemical kinetics and physical nature of things... but even if it were possible, allowable peak pressure rise rate (~10 bar/° as stated above) would cause you to have to spread out the burn. The 10 bar/° number has to do with head gasket integrity, connecting rod strength, rod bearings, etc.



Moreover, it is generally assumed that better atomization, resulting from higher injection pressure, increases combustion efficiency. This is generally true. However, you will find a point that the parasitic loss of pumping the fuel to higher pressures is greater than the increase in efficiency from better atomization... and as such, bsfc is actually lower.



--Eric

We use DCat some from Drivven for real-time heat release and combustion feedback. There are others as well, AVL etc, but mostly they will indicate 10%, 50%, and 90% MFB. Some people use 5% as the start of combustion indicator, and some 1% (although that's so noisy that results are not consistant). Thus, I suppose you could compare CA degrees between MFB10 and MFB90 as "burn time". There are so many things that affect burn time, that it can vary drastically... fuel properties, engine design, piston design, injection pressure, cam design, egr rate, swirl and tumble, etc, etc. As a general rule of thumb, injection timing needs to be advanced 8 - 10° for low cetane fuels to have the same combustion phasing (MFB50) as a high cetane fuel (low cetane meaning 40 and high meaning 55 or so).

the later half i have a good understanding of fuel properties and so on, but could you explain more the the earlier part that i have copied here, i would like as much knowledge as i can gain from what your telling me, but some of the terms being used are new to me and i dont fully grasp their meaning.



another question i thought of today was is there anyway to calculate the coeffiecent of friction inside the engine? im trying to find how much work is being done to over come just those losses inside the engine. also along with thermodynamics and the associated formulas i want as much info as you can give me on this. i know ill have more questions as i have more information so be ready for it.

 

[Mhannink]

another question, full load? if i remember right that is measured at max hp rpm and full fuel correct? so if im calculating anything related to full load then rpm is based off adv. hp. at whatever rpms its rated at correct?

 

[ENafziger]

Wow... well, I'm definately not the best one to be giving a course on engine and combustion, but I'll answer as I have time.

Full load is somewhat subjective, primarily to the constraints I gave earlier. The 400 psi cylinder pressure I referenced is simply a rough value for the pressure in the cylinder at combustion TDC at the time fuel is injected... regardless of speed/load. The 160 bar or ~2300 psi value I gave is a generally accepted upper limit of allowable maximum cylinder pressure for a typical light duty modern diesel engine. Yes, 12 meters/sec is generally accepted maximum piston velocity. Obviously you can push these limits further, but at the detriment of longevity. OEMs are rarely interested in pushing things to the ragged edge for model year release because they warranty parts. For the research world, the OEMs push things to failure.

When an OEM publishes "full load", there is usually a rated torque and a rated horsepower, obviously occuring at two different engine speeds. When we are running an engine for research purposes, there is no such thing as maximum load; rather, you always are limited by a known boundary condition of probably engine failure... for example, the maximum pressure rise rate, cylinder peak pressure, exhaust temperature, piston velocity, etc. Within those boundaries, you can change things however you would like.

Most of the work going on now isn't just inject fuel and burn it... it has to do with different combustion strategies such as PCCI, RCCI, HCCI, etc, etc, where combustion is pre-mixed charged, reactivity controlled, homogeneously charged, etc. Variable compression ratio, camless engines, dual-fuel (gasoline/diesel) engines, novel 6 stroke cycles, waste heat recovery, turbo compounding, "super" turbos, emulsifications, etc are being looked at. The trend is toward smaller displacement, more highly turbocharged engines.

There is no good formula for friction. It is much better measured than calculated. With a motoring dyno, you get the engine up to temperature and kill fueling. There are other ways to do it if you don't have a motoring dyno, but it's not a paper and pencil exercise.

By definition,

fmep = imep - bmep

or friction mean effective pressure is the difference between indicated mean effective pressure and brake mean effective pressure. Indicated work is the area enclosed on a p-V diagram from an engine and represents the work done by the gas on the piston. This indicated work per unit swept volume is imep.

The negative work that occurs during the intake and exhaust strokes is termed pumping loss, and per unit swept volume, is termed pmep.

There is sometimes confusion with the term imep, such that some people include pumping work in imep and some don't. To be more clear, gross imep = net imep + pmep.

The work output of an engine as measured by a brake or dynomometer is termed bmep, or p sub b (pb). Brake power = (pb)(piston stroke)(piston area)(number of mechanical cycles of operation per second)

Mechanical efficiency is defined as bmep/imep.

Thermal efficiency is work out/heat energy in. In the case of a diesel engine, I guess you could simplify to:

TE = (0. 01929*HP) / Fuel flow

TE being thermal efficiency
HP being horsepower
Fuel flow in gallons per hour diesel

I don't mind answering specific questions, but in reality, for general information, your probably better off reading some SAE journals, Wikipedia, or Internal Combustion Engines handbook. I'm not the best at explaining things, and quite frankly am too time limited right now to try. :/

--Eric

 

[Mhannink]

Wow... well, I'm definately not the best one to be giving a course on engine and combustion, but I'll answer as I have time.



Full load is somewhat subjective, primarily to the constraints I gave earlier. The 400 psi cylinder pressure I referenced is simply a rough value for the pressure in the cylinder at combustion TDC at the time fuel is injected... regardless of speed/load. The 160 bar or ~2300 psi value I gave is a generally accepted upper limit of allowable maximum cylinder pressure for a typical light duty modern diesel engine. Yes, 12 meters/sec is generally accepted maximum piston velocity. Obviously you can push this limit farther, but at the detriment of longevity. OEMs are rarely interested in pushing things to the ragged edge for model year release because they warranty parts. For the research world, the OEMs push things to failure.



When an OEM publishes "full load", there is usually a rated torque and a rated horsepower, obviously occuring at two different engine speeds. When we are running an engine for research purposes, there is no such thing as maximum load; rather, you always are limited by a known boundary condition of probably engine failure... for example, the maximum pressure rise rate, cylinder peak pressure, exhaust temperature, piston velocity, etc. Within those boundaries, you can change things however you would like.



Most of the work going on now isn't just inject fuel and burn it... it has to do with different combustion strategies such as PCCI, RCCI, HCCI, etc, etc, where combustion is pre-mixed charged, reactivity controlled, homogeneously charged, etc. Variable compression ratio, camless engines, dual-fuel (gasoline/diesel) engines, novel 6 stroke cycles, waste heat recovery, turbo compounding, "super" turbos, emulsifications, etc are being looked at. The trend is toward smaller displacement, more highly turbocharged engines.



There is no good formula for friction. It is much better measured than calculated. With a motoring dyno, you get the engine up to temperature and kill fueling. There are other ways to do it if you don't have a motoring dyno, but it's not a paper and pencil exercise.



By definition,



fmep = imep - bmep



or friction mean effective pressure is the difference between indicated mean effective pressure and brake mean effective pressure. Indicated work is the area enclosed on a p-V diagram from an engine and represents the work done by the gas on the piston. This indicated work per unit swept volume is imep.



The negative work that occurs during the intake and exhaust strokes is termed pumping loss, and per unit swept volume, is termed pmep.



There is sometimes confusion with the term imep, such that some people include pumping work in imep and some don't. To be more clear, gross imep = net imep + pmep.



The work output of an engine as measured by a brake or dynomometer is termed bmep, or p sub b (pb). Brake power = (pb)(piston stroke)(piston area)(number of mechanical cycles of operation per second)



Mechanical efficiency is defined as bmep/imep.



--Eric

and this is why i want to get into research, this is absolutely fascinating, anyways first off dont consider this a course on engine and combustion, just a more in depth look, as i indicated before i am a mechanic, trained in depth and to great detail on all aspects of mechanics, i. e. engine theory, powertrains, electrical, electronics, hydraulics, etc. , as good as those classes where, they stop short of this level, mostly cause in the feild repairing these things the theories are needed for understanding, the math and fine points are never used, as long as you understand how it is supposed to work you can pinpoint the problem. as for myself im in between the two, i am/was a mechanic but i am studing and working towards a mechanical engineering degree, im advancing past the diag. and repair to the desig and build side of it. this is where my questions are coming from, but i have a good base for understanding all of this, not to mention physics have always been my strong suit.



anyways, 400 psi cyl. pressure is the pressure @tdc compression just prior to fuel being injected correct? so that is a mean pressure factoring in all loads, across all diesels engines? i ask because the cylinder pressure is going to change a lot with load. as i know it, and i could be wrong, cylinder pressure = comp. ratio(atomosphieric air pressure + boost pressure), that would change a lot with changes in load.



2,300 psi, that is more or less absolute? under no conditions, excepting moded engines, should cylinder pressure be higher than that value, if so then increased wear and possible catistrauphic(sp?) engine failure could result.



12 m/s im still not sure, is that maximum mean speed or maximum peak speed? can the piston peak higher than that and be ok long term or does the average speed of the piston have to be less than that for longevity?



as for full load, im going to have to look up in my books or email one of my old instructors and get the definition we use in industry, i think it varies from your definition, i want to find the disconnect. yes, the manufacturer does give peak torque and horsepower and related rpm, but thats usually in the service manual lit or on the comp. on the engine information plates only advertised hp is displayed among other specs, torque is often not included. full load is either are peak torque and full fuel or peak horsepower and full fuel, i cant remeber which. i know full fuel is key, i think full load is where you are at fuel fuel and pull the engine down to peak torque, generating torque rise, this is the full load point. there is mroe to it but i will have to go back and look.



in my research for my little project i have come across many articles talking about what you seem to be researching it all looks very interesting, the 6 stroke theory is a new one to me though, can you explain the principle?



my last question for the moment is with the friction information, the way you said its measured, does that seem really acurate? wouldnt the compression affect the measurement, or did i just miss something in your explanation? i will have to read over a few more times before i will really start to get it.



i hope i dont sound argumenitive in any of this, something just dont compute right now and i want to find out why, i do not doubt your information, but there are things missing between me and you, im just trying to build a bridge across that gap.

 

[Mhannink]

Thermal efficiency is work out/heat energy in. In the case of a diesel engine, I guess you could simplify to:



TE = (0. 01929*HP) / Fuel flow



TE being thermal efficiency

HP being horsepower

Fuel flow in gallons per hour diesel



I don't mind answering specific questions, but in reality, for general information, your probably better off reading some SAE journals, Wikipedia, or Internal Combustion Engines handbook. I'm not the best at explaining things, and quite frankly am too time limited right now to try. :/



first i thank you for your efforts, i try to take as much information in as i can from all sources good bad and indifferent, i understand you are time limited and please dont go out of your way for my benifit, any explanations/information is good, whenever it is convientint for you. i always have lots of questions and like to discuss things, and if someone is busy and cant get involved thats ok we all have lives to live and those come first.

 

[HHhuntitall]

Thanks for the time you have given... ... I'm understanding a little more as I follow along..... I'm somewhat like MHannick, I've been involved in mechanics all my life, and often have to design a lot of improvised equiment and tools for field use..... we do quite a bit of farming..... so bigger Cummins and Cats are nothing new... . but reading your engineering formulae is exciting... . how wierd is that?!?!

 

[ENafziger]

anyways, 400 psi cyl. pressure is the pressure @tdc compression just prior to fuel being injected correct? so that is a mean pressure factoring in all loads, across all diesels engines? i ask because the cylinder pressure is going to change a lot with load. as i know it, and i could be wrong, cylinder pressure = comp. ratio(atomosphieric air pressure + boost pressure), that would change a lot with changes in load.

Almost right, but calculating cylinder pressure is a little trickier than that. Fluids heat up as they are compressed, and as such, the temperature rise causes an associated pressure rise... for an ideal gas, I think the pressure rise due to temperature could be related by:

PV = nRT

or solving for pressure, P = nRT / V

where P is absolute pressure, n is number of moles of the fluid, R is the gas constant, T is temperature, and V is volume. The main thing here is that temperature and pressure are directly proportional, such that doubling the temperature will double the pressure (assuming the fluid stays the same, the amount of fluid is constant, and the volume doesn't change).

Nonetheless, by using the specific heat ratio of the working fluid, you can account for the pressure rise due to temperature. (Specific heat ratio is designated by lowercase "gamma", but I don't know how to make that symbol here. ) Thus, the formula becomes:

Pcylinder = Po x CR^gamma

The specific heat ratio for air is 1. 4, such that the formula becomes:

Pcylinder = Po x CR^1. 4

Po is cylinder pressure at bottom dead center
CR is compression ratio

So, for an example, if Po is atmospheric pressure of 1 bar, and compression ratio is 10, we have:

Pcylinder = 1 bar x 12 ^ 1. 4
= 32. 4 bar
= 470 psi

So, part of the resultant cylinder pressure is from the compression of the air, and part is from expansion of the air due to temperature rise.

At least I think that's what I remember...

--Eric

 

[ENafziger]

my last question for the moment is with the friction information, the way you said its measured, does that seem really acurate? wouldnt the compression affect the measurement, or did i just miss something in your explanation? i will have to read over a few more times before i will really start to get it.

You're exactly right... motoring trace gives you pmep + fmep, or pumping loss plus friction loss. With pressure trace information and the p-V diagram, you know imep(gross) and imep(net). Then, imep(gross) - imep(net) = pmep. So, then when you get the motoring trace (pmep + fmep), subtract pmep and you're left with friction

Frictional loss is primarily dependant on rpm, followed by temperature.

--Eric

 

[Mhannink]

Almost right, but calculating cylinder pressure is a little trickier than that. Fluids heat up as they are compressed, and as such, the temperature rise causes an associated pressure rise... for an ideal gas, I think the pressure rise due to temperature could be related by:



PV = nRT



or solving for pressure, P = nRT / V



where P is absolute pressure, n is number of moles of the fluid, R is the gas constant, T is temperature, and V is volume. The main thing here is that temperature and pressure are directly proportional, such that doubling the temperature will double the pressure (assuming the fluid stays the same, the amount of fluid is constant, and the volume doesn't change).



Nonetheless, by using the specific heat ratio of the working fluid, you can account for the pressure rise due to temperature. (Specific heat ratio is designated by lowercase "gamma", but I don't know how to make that symbol here. ) Thus, the formula becomes:



Pcylinder = Po x CR^gamma



The specific heat ratio for air is 1. 4, such that the formula becomes:



Pcylinder = Po x CR^1. 4



Po is cylinder pressure at bottom dead center

CR is compression ratio



So, for an example, if Po is atmospheric pressure of 1 bar, and compression ratio is 10, we have:



Pcylinder = 1 bar x 12 ^ 1. 4

= 32. 4 bar

= 470 psi



So, part of the resultant cylinder pressure is from the compression of the air, and part is from expansion of the air due to temperature rise.



At least I think that's what I remember...



--Eric

this is all for static cylinder pressure right? once boost starts then the pressure rises above the 400 mark, without factoring in heat expansion.

 

[Mhannink]

another question, when im cruising my pyro is right around 600 degrees f, using the formula you gave te is 33% which is about right rule of thumb has always been 1/3 to water 1/3 to exhaust and 1/3 to flywheel. so if i take that 600 and multiply by three thats roughly the average cylinder temp? could it be peak? what do cylinder temps run usually.

 

[ENafziger]

this is all for static cylinder pressure right? once boost starts then the pressure rises above the 400 mark, without factoring in heat expansion. <!-- google_ad_section_end -->

For boosted condition, you would simply use the boost pressure for Po in the above equation, for pressure at bottom dead center. So, if you're running 50 psi boost, that's 3. 45 bar. Playing off the above example:

Pcylinder = 3. 45 bar x 12 ^ 1. 4
= 111. 9 bar
= 1623 psi

another question, when im cruising my pyro is right around 600 degrees f, using the formula you gave te is 33% which is about right rule of thumb has always been 1/3 to water 1/3 to exhaust and 1/3 to flywheel. so if i take that 600 and multiply by three thats roughly the average cylinder temp? could it be peak? what do cylinder temps run usually. <!-- google_ad_section_end --> <!-- / message --><!-- sig -->

No, things are not that simple... you're thinking in terms of degrees instead of energy, and mixing units doesn't work. When they say 1/3 of the energy of the fuel goes to water, 1/3 to exhaust, and 1/3 to the flywheel, they aren't talking degrees F. It's more of an energy transfer, or energy flux. The water and exhaust happen to be energy in the form of heat, whereas the flywheel is mechanical energy output (indirectly anyways... ). As such, the temperature you read on your pyro is indicative of the combined average exhaust gas temp at the point of measurement... but it speaks nothing to gas velocity, flowrate, etc. I don't remember enough from heat transfer to bring out equations here, but Q with a dot over it is heat transfer, and includes both temperate and mass flow.

roughly the average cylinder temp? could it be peak? what do cylinder temps run usually.

No, it's not peak. There are not just simple equations to use for this stuff. I mean, you can kindof get within 40% of the real value sometimes with making a LOT of assumptions and simplifications, but then you have to ask what the motivation is behind even calculating it. You can't run your truck on settings that have a 40% possible error band. Diesel combustion is commonly divided into two parts: the pre-mixed burning phase and the diffusion burning phase. Pre-mixed is a consequence of the mixture prepared during the ignition delay period burning rapidly; the diffusion burning phase accounts for the remainder of combustion. There are different modeling equations to generalize some of that. If you're interested, look up the Whitehouse and Way model, and the Arrhenius equation as it relates to combustion. Following this stuff starts to get above my head quickly, as it's been too long since I've had differential equations, advanced calculus, and upper level thermodynamics. Suffice it to say, that what you read on your pyro is indicative of the bulk gas temperature at that point... while the actual hottest part of the flame front during combustion can be over 2300 K, or north of 3500° F.

We talked above about friction, and it being primarily an experimental measurement. While that is true, there are correlations to somewhat predicting fmep, such as Chen and Flynn's correlation:

fmep = 0. 137 + (pmax / 200) + 0. 162 vp

pmax is cylinder pressure bar
vp is mean piston speed in m/s

Pretty much for anything accurate, you have to use a good computational fluid dynamics program (CFD).

--Eric

 

[HHhuntitall]

Suffice to say, with a 40% accuracy, you're back to trial and error..... and watching real time inputs from testing on an engine stand or putting all that equipment on a truck and trying to track it while on a dyno... ... and then interpreting the results of various combinations. I'm curious as how your company/department handles engine testing. Is it all on "stands" or do you test in vehicle, too? Do you modify engines/cylinder heads/turbos for instrumentation?



Also, I also see you have flow bench equipment..... ever tested the 5. 9 heads? I'm curious if you've ever noted the specific restricitons... ...

 

[ENafziger]

I'm curious as how your company/department handles engine testing. Is it all on "stands" or do you test in vehicle, too? Do you modify engines/cylinder heads/turbos for instrumentation?

Also, I also see you have flow bench equipment..... ever tested the 5. 9 heads? I'm curious if you've ever noted the specific restricitons... ... <!-- google_ad_section_end --> <!-- / message --><!-- sig -->

We do all of the instrumentation in house, with the exception of drilling/tapping the head or block for the in-cylinder pressure transducers. This is something that can be very difficult for the OEMs, even with full 3-D models, etc, to penetrate the head and keep the integrity of cooling passages, oil, etc. We have 7 engine test cells currently, with a large dynomometer in the center of the cell and an engine on either end. Obviously, only one engine can be run in a cell at a time. There is one chassis dyno lab for doing vehicle testing, but everything else is done on an open platform, with an engine just "sitting" open in the room.

The bench flow stuff has little to do with flow testing of engine components; rather, it involves flowing a mixture of compressed gases through a heated catalyst core for "bench" scale development work of aftertreatment systems.

I wish I could do more on a 5. 9 Cummins platform, but I don't have any say-so over that. A typical engine setup with full instrumentation will take 6 months and $100k - 300k. As such, we have to stick exactly with the project at hand, as the cost per hour of running is astronomically out of the ballpark of anything I or someone could afford out of personal curiousity.

The instrumentation is pretty incredible though. It's amazing to be able to see and know nearly everything that is going on inside of a cylinder, and to be able to log all of that information on a cycle-by-cycle basis. The instrumentation is so accurate that you can see acceleration change in the crankshaft between combustion events! It boggles my mind how far technology has come, but the sampling rate of data, processing algorithms, and controls is so advanced that it's possible to do real-time predictive next cycle combustion control! In other words, for combustion regimes approaching the limits of stability, you can take measurements, put the data through algorithms, make a determination of what needs to be done and what needs changed to prevent a misfire, send the information back to the corresponding engine controls/sensors, and make the change in time to affect the very next combustion event!

The emissions equipment is really incredible as well. When we're running, we can see speciation of the exhaust (chemical makeup in the exhaust), the particle size distribution, number of particles of each size, etc. A guy here developed a sampling probe with a diameter smaller than a human hair, that Cummins is using for non-intrusive gas sampling along the length of a catalyst to get an idea of the exact chemical reactions that are taking place at different reaction sites.

--Eric

 

[HHhuntitall]

Awesome. I guess you're there everyday, but I'd have a hard time leaving for a few weeks. That type of in-depth knowledge of what's going on inside an engine. Wow. That's all I can say. The electronic part of technology has come a long way, I'd agree. I'm dissapointed in the mechanical aspect, though. These trucks are really not getting much better mileage than they were 15 years ago. Albiet, there is much more power readily available to the throttle. Where's the compromise?



I know TenneCo and Cummins worked together to develop their emission system, and granted, it's the best on the market right now, but what's in the future? I'm disgusted at the possibilities. Urea injection? Seriously? How can you be making less emissions if you're burning more fuel? Now, I know they consider the emissions they are making more environmentally friendly, and they're more stable elements, readily dispersed, but what about the effeciency? Am I off in saying the standards the Feds have put up for the industry to be met are realistically out of reach? And I'm not trying to draw you into a political discussion, only mechanical reality..... Our advancements into fuel injection have only gone so far in the last 15 years.



It does sound like a tedious, but rewarding job. The number readouts would get old, but working with the engines and experimentation of parameters sounds like sooooo much fun.