I am open to suggestions. Reliabilty is a key issue. I have to plan for longevity at 300 degrees. Maybe that answers that.
But always open to suggestions
But always open to suggestions
Cooler line pressure drop
- Thread starter Killerbees
- Start date
- Status
- Joined
- Sep 28, 2002
- Messages
- 39,793
Let me see here.
You need hose with JSAE J1019 on it.
C5D Hose
For petroleum base or phosphate ester fluids. Recommended for diesel engine flexlines.
• Meets or exceeds SAE J1402 Type AII, SAE J1019 and DOT FMVSS 106-74 Type AII.
• 400 to 1,500 psi working pressure.
• One high tensile steel wire braid over polyester braid.
• CPE tube.
• Sizes 3/16" through 7/8" I.D.
• Green braided textile covers for easy identification.
• Temperature range –40°C to +149°C (–40°F to +300°F).
Uses C5 style field attachable couplings.
C5E Hose
For air brake hose, power steering, fuel filter, engine and transmission coolant lines and hot +149°C (+300°F) lube lines. Recommended for diesel engine flexlines.
• Meets DOT FMVSS 106-74 Type AI and SAE J1019 hot oil circulation test.
• 300 to 1,500 psi working pressure.
• One braid high tensile carbon steel wire over polyester braid.
• Textile braid impregnated with rubber on black cover.
• Nitrile tube.
• Sizes 3/16" through 11/8" I.D.
Uses C5E type field attachable or permanent couplings.
Gates
This stuff is pretty cheap in the smaller sizes (5/16, etc) ..I think it would be at a decent discount to teflon ..or so I would reason
You need hose with JSAE J1019 on it.
C5D Hose
For petroleum base or phosphate ester fluids. Recommended for diesel engine flexlines.
• Meets or exceeds SAE J1402 Type AII, SAE J1019 and DOT FMVSS 106-74 Type AII.
• 400 to 1,500 psi working pressure.
• One high tensile steel wire braid over polyester braid.
• CPE tube.
• Sizes 3/16" through 7/8" I.D.
• Green braided textile covers for easy identification.
• Temperature range –40°C to +149°C (–40°F to +300°F).
Uses C5 style field attachable couplings.
C5E Hose
For air brake hose, power steering, fuel filter, engine and transmission coolant lines and hot +149°C (+300°F) lube lines. Recommended for diesel engine flexlines.
• Meets DOT FMVSS 106-74 Type AI and SAE J1019 hot oil circulation test.
• 300 to 1,500 psi working pressure.
• One braid high tensile carbon steel wire over polyester braid.
• Textile braid impregnated with rubber on black cover.
• Nitrile tube.
• Sizes 3/16" through 11/8" I.D.
Uses C5E type field attachable or permanent couplings.
Gates
This stuff is pretty cheap in the smaller sizes (5/16, etc) ..I think it would be at a decent discount to teflon ..or so I would reason
Good thought. Can -12 be had for under $3 lf?
The Teflon may be cheaper than you think. Plus the added temperature protection and SS bling factor, seems popular without practicality though. One other minor point, teflon has a beautiful surface for minimizing pressure drop, since we are on the topic.
Can be used with reuseable fittings (DIY) which makes the hoses easy to replace, cheap for life.
The Teflon may be cheaper than you think. Plus the added temperature protection and SS bling factor, seems popular without practicality though. One other minor point, teflon has a beautiful surface for minimizing pressure drop, since we are on the topic.
Can be used with reuseable fittings (DIY) which makes the hoses easy to replace, cheap for life.
- Joined
- Sep 28, 2002
- Messages
- 39,793
- Joined
- Sep 28, 2002
- Messages
- 39,793
Although Teflon has a low coefficient of friction, I don't know how much lower it is then the aforementioned hoses. I didn't catch any #'s
The reusable fittings will probably reduce long term costs ..but they are proprietary to teflon hose. This will, again, raise the ante of the package a bit and will obligate the user to assuring that teflon hose is available for refitting/refreshing as needed.
Well, you appear to have this situation well in hand here.
It appears that the rad type cooler, with tanks on either end, is the preferred cooler of choice for those who use air/oil coolers. Apparently the tank chamber makes the transition from one laminar flow to several a bit easier.
Somehow I envision a full width version of this is what you're looking for ..
The reusable fittings will probably reduce long term costs ..but they are proprietary to teflon hose. This will, again, raise the ante of the package a bit and will obligate the user to assuring that teflon hose is available for refitting/refreshing as needed.
Well, you appear to have this situation well in hand here.
It appears that the rad type cooler, with tanks on either end, is the preferred cooler of choice for those who use air/oil coolers. Apparently the tank chamber makes the transition from one laminar flow to several a bit easier.
Somehow I envision a full width version of this is what you're looking for ..
How funny, I have one of those in my hands. 3.5" thick Fluidyne, it is a nascar blank. Needs 140 mph to get rated rejection of 60K BTU/hr.
And restrictive!!! Both oil and air flow. A 5 psi drop at 7 gpm, goes exponentially off the chart from there. But I am prototyping it anyway. I need one of those with about 3 more rows. Better, rows oriented vertically. But you nailed it, those dimensions are perfect.
And restrictive!!! Both oil and air flow. A 5 psi drop at 7 gpm, goes exponentially off the chart from there. But I am prototyping it anyway. I need one of those with about 3 more rows. Better, rows oriented vertically. But you nailed it, those dimensions are perfect.
- Joined
- Sep 28, 2002
- Messages
- 39,793
I'm ignorant on the btu rating of various coolers. As you indicate the one you're looking at needs 140 mph to reach 60kbtu ..but I don't see how they can figure that without temp differentials.
When we ordered a heat exchanger, you had a "Q" factor. That indexed the medium that you were transfering the btu's through. Then you factored in the square footage ..and it went from there. We naturally had condensors that were ultimately rejecting the exchanged heat ..and they too were rated under the same criteria. You had intermediate losses in terms of condensate in subsequent effects ..so your condensor(s) didn't have to match your primary heat exchanger(s) Ultimately, it came down to your differential across the exchanger, indexed for the medium (Q) that determined your btu exchange.
So, do you have the standards that were used to require 140 mph for your desired 60k btu? We had an btu/hr figure that was in there somehow, since we were dealing with lbs/hr of saturated live steam ..so we could merely read our steam draw and our btu/hr was easily figured.
When we ordered a heat exchanger, you had a "Q" factor. That indexed the medium that you were transfering the btu's through. Then you factored in the square footage ..and it went from there. We naturally had condensors that were ultimately rejecting the exchanged heat ..and they too were rated under the same criteria. You had intermediate losses in terms of condensate in subsequent effects ..so your condensor(s) didn't have to match your primary heat exchanger(s) Ultimately, it came down to your differential across the exchanger, indexed for the medium (Q) that determined your btu exchange.
So, do you have the standards that were used to require 140 mph for your desired 60k btu? We had an btu/hr figure that was in there somehow, since we were dealing with lbs/hr of saturated live steam ..so we could merely read our steam draw and our btu/hr was easily figured.
- Joined
- Sep 28, 2002
- Messages
- 39,793
Back the pressure "toll" with the afore imaged exchanger and it's lack of size...
Why couldn't you tandem two of those side by side ..and run two in parallel? The toll should be the same as one ..assuming that you get your bends and such smooth enough. Velocity would be cut in half and your dwell time would double.
Wouldn't that bring your figure down to 35 mph? I reasoned this by figuring a single unit with half the flow should bring your 140 figure down to 70 - times two - should bring it down to 35. OTOH, a single unit of double the length (width) would only bring it down to 70 ..or so I figure since the original flow is at the higher velocity.
Why couldn't you tandem two of those side by side ..and run two in parallel? The toll should be the same as one ..assuming that you get your bends and such smooth enough. Velocity would be cut in half and your dwell time would double.
Wouldn't that bring your figure down to 35 mph? I reasoned this by figuring a single unit with half the flow should bring your 140 figure down to 70 - times two - should bring it down to 35. OTOH, a single unit of double the length (width) would only bring it down to 70 ..or so I figure since the original flow is at the higher velocity.
I use a similar looking cooler for my ATX cooler. It is 4" high, 24" long and 1.5" thick. It does a great job and I wondered why it was so much better.
Then I realized that when I put it in I got rid of the factory steel cooler lines and used 3/8" hose all the way. The OEM lines were 5/16" ID and had crush 90 degree bends further resticting the flow.
I feel the larger hose really allows the cooler to do a good job. I really like the end tank style over the stacked plate and tube and fin styles. Previously I had a stacked plate and tube and fin plumbed in series while still using the OEM steel lines. They could not come close to doing the job the single end tank style does with the larger hose.
Then I realized that when I put it in I got rid of the factory steel cooler lines and used 3/8" hose all the way. The OEM lines were 5/16" ID and had crush 90 degree bends further resticting the flow.
I feel the larger hose really allows the cooler to do a good job. I really like the end tank style over the stacked plate and tube and fin styles. Previously I had a stacked plate and tube and fin plumbed in series while still using the OEM steel lines. They could not come close to doing the job the single end tank style does with the larger hose.
"So, do you have the standards that were used to require 140 mph for your desired 60k btu?"
Yes, in the form of real test data from the manufacturer. Actual scaling of realistic highway speeds, in a simulation program, indicate the cooler you pictured, as having 47K BTU/hr. Not bad really for such a diminutive profile. That is 90 F ambient, 200 F oil, and 20 gpm. The bad news is the cooler itself drops 20 psi. No good. So realistically, flow will scale back some, and it will reject 30 K BTU/hr.
Yes, in the form of real test data from the manufacturer. Actual scaling of realistic highway speeds, in a simulation program, indicate the cooler you pictured, as having 47K BTU/hr. Not bad really for such a diminutive profile. That is 90 F ambient, 200 F oil, and 20 gpm. The bad news is the cooler itself drops 20 psi. No good. So realistically, flow will scale back some, and it will reject 30 K BTU/hr.
"Why couldn't you tandem two of those side by side ..and run two in parallel? The toll should be the same as one ..assuming that you get your bends and such smooth enough. Velocity would be cut in half and your dwell time would double."
This question leads to the virtues of stacked plate designs, and the heat rejection superiority they offer. The configuration you mention would not be affordable, the coolers run $600 each. Yes parallel would be the only way to do it. Residence time is not why, however. Optimizing heat rejection (over time in a SS closed cycle stream operation) means minimizing residence/dwell, in a given design. Oddly, the presence of some significant pressure drop is important. Without it, we know that flow will be an inferior laminar. It needs to be turbulent, or heat exchange will be reduced dramatically, 30% or more.
slow flow in long tubes-bad
fast flow in short tubes-good
The temp of the oil at the exit and the mass flow rate, define rejection. M:dot X Cp X delta T (T2-T1) If not familiar just ignore the equation.
Suffice to say that you can't move the oil too fast. Heat rejection approaches optimal maximum, as flow rate approaches infinity. In this example of 2 parallel coolers, as long as each has it's own stream, and turbulent oil flow, heat rejection will be approx double, possibly a little more due to lower pump head.
If the same cooler pictured had vertical tubes (many) instead, the scenario would be different. Heat rejection, BTU/hr, would go up 20% roughly. But expensive to manufacture. Existing stack plate designs do a fair job. A 60-80 plate design, vertical tubes, that will fit in the same spot, a little wider, will reject 50K BTU/hr on paper. 2" thick instead of 3", means more airflow at low speeds, more optimal for real roads. Pressure drop at 20 GPM is less than 1 psi, and laminar/transition flow exists. But that's ok, they planned for it, and add turbulators/mixers. Turns out, this method kicks but, and let's the pump live a long life.
This question leads to the virtues of stacked plate designs, and the heat rejection superiority they offer. The configuration you mention would not be affordable, the coolers run $600 each. Yes parallel would be the only way to do it. Residence time is not why, however. Optimizing heat rejection (over time in a SS closed cycle stream operation) means minimizing residence/dwell, in a given design. Oddly, the presence of some significant pressure drop is important. Without it, we know that flow will be an inferior laminar. It needs to be turbulent, or heat exchange will be reduced dramatically, 30% or more.
slow flow in long tubes-bad
fast flow in short tubes-good
The temp of the oil at the exit and the mass flow rate, define rejection. M:dot X Cp X delta T (T2-T1) If not familiar just ignore the equation.
Suffice to say that you can't move the oil too fast. Heat rejection approaches optimal maximum, as flow rate approaches infinity. In this example of 2 parallel coolers, as long as each has it's own stream, and turbulent oil flow, heat rejection will be approx double, possibly a little more due to lower pump head.
If the same cooler pictured had vertical tubes (many) instead, the scenario would be different. Heat rejection, BTU/hr, would go up 20% roughly. But expensive to manufacture. Existing stack plate designs do a fair job. A 60-80 plate design, vertical tubes, that will fit in the same spot, a little wider, will reject 50K BTU/hr on paper. 2" thick instead of 3", means more airflow at low speeds, more optimal for real roads. Pressure drop at 20 GPM is less than 1 psi, and laminar/transition flow exists. But that's ok, they planned for it, and add turbulators/mixers. Turns out, this method kicks but, and let's the pump live a long life.
BTW, you will find the 911 track guys using the long stacked plate with vertical tubes. They also move 20 gpm. It is the closest no-compromise approach I am aware of thus far.
As the previous poster points out, line size can destroy the effectiveness of even the best design. Hence the need for this thread.
If a -8 line was used in the long stacked plate 20 gpm application, the cooler would drop 1 psi, the lines and connections would drop 100 psi.
The one disadvantage that may exist to stacked plate, is they may not have the reliability of the tube/fin counterparts.
As the previous poster points out, line size can destroy the effectiveness of even the best design. Hence the need for this thread.
If a -8 line was used in the long stacked plate 20 gpm application, the cooler would drop 1 psi, the lines and connections would drop 100 psi.
The one disadvantage that may exist to stacked plate, is they may not have the reliability of the tube/fin counterparts.
- Joined
- Sep 28, 2002
- Messages
- 39,793
I need a little more edumication here on this point. Now sure, you've got to cycle your oil as fast as possible for collecting the heat ..but dumping it isn't the same thing unless you've got a heat sink of equal potential.
In our cooling towers ...we had rotating distribution arms. They basically sprinkled the hot btu rich water over diffusion material (a psuedo turbulant flow if you will - effectively mulitplying the cooling surface by multiples of 1000s) as the high speed fans drew air upwards from the louvered bottom vents of the towers. If you slowed the flow, the draft being the same, you got a broader delta T ..that's just common sense. Now the system itself rejected less heat then opening the flow up and having a lower delta ...but that was only a situation of having 100% output at all times (that it mattered in terms of cooling). If we had 2 more towers ..the flow could crawl through them and it would be the best of both worlds. Max flow and max delta T.
So I don't quite get how having less oil pass over the same cooling surface cannot result in lower outlet temperatures. Otherwise you could say that the smaller the cooler ..the better it would cool since the velocity through it would be closer to infinity when compared to a cooler with more paths that slowed the velocity through it.
Help the rabble here. You've created a contradiction in my mind's eye.
Your example of evaporative cooling is quite a different process, using latent heat of H20 vaporization to extract heat. Water is the bestt coolant in the universe in it's natural state. If only we could carry a depletable pond with us when we drive.
Closed loop heat exchange, with fixed volumes of fluids, has a simple energy balance. The heat added to the air=the heat removed from the oil. The variables are:
1. M.dot oil-mass flow rate of the oil, a function of pump RPM, system design pressure drop, oil viscosity, etc
2. M.dot air-mass flow rate of air, a function ambient temp, and vehicle speed (altitude also, Headwind, as well as coolers construction and placement
3. inlet T of air
4. inlet T of oil
The outlet T's fall where they may. The energy removed/gained for oil is
(m.dot X delta T X Cp)oil
Cp is the constant, heat capacity of the fluid. If you slow M.dot of one fluid, the outlet T will drop, but the BTU's rejected in that hour will be less.
"So I don't quite get how having less oil pass over the same cooling surface cannot result in lower outlet temperatures"
It does, but this is where most people go wrong evaluating a cooler. Drop is only HALF of the energy balance.
Slow it to a crawling pace, you could get oil down to 1 deg F over ambient, right? But what good is it at 3 cc/s per second?
Our example of very slow oil flow, overly simplified: 1 gpm Oil in at 200 F. Air in at 100. Let's say halfway down the tube, oil is at 120. Now the oil has to progress the remaining half with only a 20 deg temp differential, 100 degree air trying to cool 120 degree oil. No significant heat exchange occurs in the second half of the cooler. It is a practical waste of space. The highest heat exchange occurs, integrally, where the fluid temp differences are the highest, in this example, near the cooler inlet. Differential is 100 degrees (and 5x the heat exchange rate, over a 20 degree differential)
speed up flow now (or arrange the tube vertically) to 10 gpm (10x) and we may only produce a 10 degree drop in the oil. But the energy removed is 10gpm X 10deg=100 units.
Previously it was 1gpm X 80deg=80 units
All coolers share this problem, the closer the the oil outlet you measure, the less heat exchange is occuring. So long tubes in narrow coolers suffer the most, especially since they contain more P drop forcing lower fluid velocities. Short vertical arrangements tower over in rejection rate, because the temp of the oil at the end of the tube is still high enough to contribute significantly to rejection rate (BTU/hr) and also because more M.dot can be established (lower P drop).
It is analogous to voltage. The higher the voltage (potential) the more electricity you can move. The higher the 2 fluid temp differential (potential) the higher BTU transfer.
The best we could do for heat exchange RATE, is to up mass flows of each fluid (air is a fluid) so that oil cools to 199 (1 degree), and air heats to 101 (1 degree). That represents an unobtainable fantasy in the real world, the mass flows would be supersonic. But the point is there.
Sorry this was long winded. There is no easy way for me to relate some things.
Closed loop heat exchange, with fixed volumes of fluids, has a simple energy balance. The heat added to the air=the heat removed from the oil. The variables are:
1. M.dot oil-mass flow rate of the oil, a function of pump RPM, system design pressure drop, oil viscosity, etc
2. M.dot air-mass flow rate of air, a function ambient temp, and vehicle speed (altitude also, Headwind, as well as coolers construction and placement
3. inlet T of air
4. inlet T of oil
The outlet T's fall where they may. The energy removed/gained for oil is
(m.dot X delta T X Cp)oil
Cp is the constant, heat capacity of the fluid. If you slow M.dot of one fluid, the outlet T will drop, but the BTU's rejected in that hour will be less.
"So I don't quite get how having less oil pass over the same cooling surface cannot result in lower outlet temperatures"
It does, but this is where most people go wrong evaluating a cooler. Drop is only HALF of the energy balance.
Slow it to a crawling pace, you could get oil down to 1 deg F over ambient, right? But what good is it at 3 cc/s per second?
Our example of very slow oil flow, overly simplified: 1 gpm Oil in at 200 F. Air in at 100. Let's say halfway down the tube, oil is at 120. Now the oil has to progress the remaining half with only a 20 deg temp differential, 100 degree air trying to cool 120 degree oil. No significant heat exchange occurs in the second half of the cooler. It is a practical waste of space. The highest heat exchange occurs, integrally, where the fluid temp differences are the highest, in this example, near the cooler inlet. Differential is 100 degrees (and 5x the heat exchange rate, over a 20 degree differential)
speed up flow now (or arrange the tube vertically) to 10 gpm (10x) and we may only produce a 10 degree drop in the oil. But the energy removed is 10gpm X 10deg=100 units.
Previously it was 1gpm X 80deg=80 units
All coolers share this problem, the closer the the oil outlet you measure, the less heat exchange is occuring. So long tubes in narrow coolers suffer the most, especially since they contain more P drop forcing lower fluid velocities. Short vertical arrangements tower over in rejection rate, because the temp of the oil at the end of the tube is still high enough to contribute significantly to rejection rate (BTU/hr) and also because more M.dot can be established (lower P drop).
It is analogous to voltage. The higher the voltage (potential) the more electricity you can move. The higher the 2 fluid temp differential (potential) the higher BTU transfer.
The best we could do for heat exchange RATE, is to up mass flows of each fluid (air is a fluid) so that oil cools to 199 (1 degree), and air heats to 101 (1 degree). That represents an unobtainable fantasy in the real world, the mass flows would be supersonic. But the point is there.
Sorry this was long winded. There is no easy way for me to relate some things.
I forgot to mention, and I previously alluded to it. Turbulence is key in these tubes or plates. So assuming a vertical array is always better, is not always true, unless Reynolds number over 4000 is assured, or turbulators are used, as in the case of many stacked plate mfrs.
Laminar flow is very bad in these applications.
Laminar flow is very bad in these applications.
Design of ACHEs for Viscous Liquids
For process fluids with outlet viscosities up to 20 centipoises, it is possible by using large diameter tubes and high velocities (up to 10 ft/sec) to achieve a Reynolds number at the outlet above the 2,000 critical Reynolds number, and to keep the flow in the transition region. However, this usually results in pressure drops of 30 to 100 psi. In view of the disadvantages of designing for laminar flow, this increased pressure drop is normally economically justifiable because the increase in the operating and capital cost of the pump is small compared with the decrease in the cost of the turbulent exchanger.
The biggest problem with laminar flow in tubes is that the flow in inherently unstable. The reasons for this can be demonstrated by a comparison of pressure drop and heat transfer coefficient for turbulent versus laminar flow, as functions of viscosity (u) and mass velocity (G):
Flow Type....Delta P......Heat Transfer
----------------------------Function
Turbulent..u^0.2,G^1.8...u^-0.47,G^0.8
Laminar....u^1.0,G^1.0...u^0.0,G^0.33
In an air-cooled heat exchanger, because of imperfect air-side flow distribution due to wind, or because of multiple tube rows per pass, it is likely that the flow through some of the tubes in a given pass is cooled more than that through other tubes.
With turbulent flow, pressure drop is such a weak function of viscosity (0.2 power) and such a strong function of mass velocity (1.8 power), that the flow in the colder tubes must decrease only slightly in order for the pressure drop to be the same as that in the hotter tubes. Also, as the flow slows and the viscosity increases, the heat transfer coefficient drops significantly, (-0.47 power of viscosity, 0.8 power of G), so the over-cooling is self-correcting.
With laminar flow, pressure drop is a much stronger function of viscosity (1.0 power) and a much weaker function of mass velocity (1.0 power), so the flow in the colder tubes must decrease much more to compensate for the higher viscosity. Viscosity of heavy hydrocarbons is usually a very strong function of temperature, but with laminar flow, the heat transfer coefficient is independent of viscosity, and only a weak function of mass velocity (0.33 power), so the selfcorrection of turbulent flow is absent.
The result is that many of the tubes become virtually plugged, and a few tubes carry most of the flow. Stability is ultimately achieved in the high flow tubes as a result of high mass velocity and increased turbulence, but because so many tubes carry little flow and contribute little cooling, a concurrent result is high pressure drop and low performance. The point at which stability is reached depends on the steepness of the viscosity versus temperature curve. Fluids with high pour points may completely plug most of an exchanger.
This problem can sometimes be avoided by designing deep bundles to improve air flow distribution. Bundles should have no more than one row per pass, and should preferably have at least two passes per row, so that the fluid will be mixed between passes.
When a fluid has both a high viscosity and a high pour point, long cooling ranges should be separated into stages. The first exchanger should be designed for turbulent flow, with the outlet temperature high enough to ensure an outlet Reynolds number above 2,000 even with reduced flow. The lower cooling range can be accomplished in a serpentine coil (a coil consisting of tubes or pipes connected by 180' return bends, with a single tube per pass). The low temperature serpentine coil should, of course, be protected from freezing by external warm air recirculation ducts.
Closed loop tempered water systems are often more economical, and are just as effective as a serpentine coil. A shell and tube heat exchanger cools the viscous liquid over its low temperature range on the shell side. Inhibited water is recirculated between the tube side of the shell and tube and an ACHE, where the heat is exhausted to the atmosphere.
For viscous fluids which are reasonably clean, such as lube oil, it is possible to increase the tube side coefficient between four- and tenfold, with no increase in pressure drop, by inserting turbulence promoters, and designing for a lower velocity. It is then advantageous to use external fins to increase the airside coefficient also. In addition to the increase in heat transfer coefficient, turbulence promoters have the great advantage that the pressure drop is proportional to the 1.3 power of mass velocity, and only to the 0.5 power of viscosity, so that non-isothermal flows are much more stable. The simplest and probably the most cost-effective promoters are the swirl strips, a flat strip twisted into a helix.
For process fluids with outlet viscosities up to 20 centipoises, it is possible by using large diameter tubes and high velocities (up to 10 ft/sec) to achieve a Reynolds number at the outlet above the 2,000 critical Reynolds number, and to keep the flow in the transition region. However, this usually results in pressure drops of 30 to 100 psi. In view of the disadvantages of designing for laminar flow, this increased pressure drop is normally economically justifiable because the increase in the operating and capital cost of the pump is small compared with the decrease in the cost of the turbulent exchanger.
The biggest problem with laminar flow in tubes is that the flow in inherently unstable. The reasons for this can be demonstrated by a comparison of pressure drop and heat transfer coefficient for turbulent versus laminar flow, as functions of viscosity (u) and mass velocity (G):
Flow Type....Delta P......Heat Transfer
----------------------------Function
Turbulent..u^0.2,G^1.8...u^-0.47,G^0.8
Laminar....u^1.0,G^1.0...u^0.0,G^0.33
In an air-cooled heat exchanger, because of imperfect air-side flow distribution due to wind, or because of multiple tube rows per pass, it is likely that the flow through some of the tubes in a given pass is cooled more than that through other tubes.
With turbulent flow, pressure drop is such a weak function of viscosity (0.2 power) and such a strong function of mass velocity (1.8 power), that the flow in the colder tubes must decrease only slightly in order for the pressure drop to be the same as that in the hotter tubes. Also, as the flow slows and the viscosity increases, the heat transfer coefficient drops significantly, (-0.47 power of viscosity, 0.8 power of G), so the over-cooling is self-correcting.
With laminar flow, pressure drop is a much stronger function of viscosity (1.0 power) and a much weaker function of mass velocity (1.0 power), so the flow in the colder tubes must decrease much more to compensate for the higher viscosity. Viscosity of heavy hydrocarbons is usually a very strong function of temperature, but with laminar flow, the heat transfer coefficient is independent of viscosity, and only a weak function of mass velocity (0.33 power), so the selfcorrection of turbulent flow is absent.
The result is that many of the tubes become virtually plugged, and a few tubes carry most of the flow. Stability is ultimately achieved in the high flow tubes as a result of high mass velocity and increased turbulence, but because so many tubes carry little flow and contribute little cooling, a concurrent result is high pressure drop and low performance. The point at which stability is reached depends on the steepness of the viscosity versus temperature curve. Fluids with high pour points may completely plug most of an exchanger.
This problem can sometimes be avoided by designing deep bundles to improve air flow distribution. Bundles should have no more than one row per pass, and should preferably have at least two passes per row, so that the fluid will be mixed between passes.
When a fluid has both a high viscosity and a high pour point, long cooling ranges should be separated into stages. The first exchanger should be designed for turbulent flow, with the outlet temperature high enough to ensure an outlet Reynolds number above 2,000 even with reduced flow. The lower cooling range can be accomplished in a serpentine coil (a coil consisting of tubes or pipes connected by 180' return bends, with a single tube per pass). The low temperature serpentine coil should, of course, be protected from freezing by external warm air recirculation ducts.
Closed loop tempered water systems are often more economical, and are just as effective as a serpentine coil. A shell and tube heat exchanger cools the viscous liquid over its low temperature range on the shell side. Inhibited water is recirculated between the tube side of the shell and tube and an ACHE, where the heat is exhausted to the atmosphere.
For viscous fluids which are reasonably clean, such as lube oil, it is possible to increase the tube side coefficient between four- and tenfold, with no increase in pressure drop, by inserting turbulence promoters, and designing for a lower velocity. It is then advantageous to use external fins to increase the airside coefficient also. In addition to the increase in heat transfer coefficient, turbulence promoters have the great advantage that the pressure drop is proportional to the 1.3 power of mass velocity, and only to the 0.5 power of viscosity, so that non-isothermal flows are much more stable. The simplest and probably the most cost-effective promoters are the swirl strips, a flat strip twisted into a helix.
- Joined
- Sep 28, 2002
- Messages
- 39,793
So? Your point. You merely expand the cooler to compensate.
Sure evaporative cooling has unique characteristics that can't be duplicated in a closed system. Even in that system your priciple essentially prevails. Without the fancy equations ..I'll readily admit that, for whatever ancillary (or fundamental for that matter)reason the higher velocity flow will reject more btu's/hr. That totally assumes that it's through the same medium/cooler/exchanger. It can't alter the fact, that some things being variable, you, under all circumstances will want the lowest outlet temp.
In our forced circulation heat exchangers, we couldn't lower tube velocity ..even though it WOULD increase btu uptake..UNTIL.... The same evaporative properties that you cited were the only reason. At some point, the counter current flow would become saturated and vaoporize ..losing heat transfer ability. Effectively boiling the liquour in the tube. With the three pass (baffled) you would get more GAIN over less liquour. We would cool less steam for the sake of a higher differential.
I could just as easily choke you to 70 mph with the aforeviewed cooler and say that you've now have got to increase your flow to 40 gpm to get your 60kbtu/hr. There has to be recipricol equivalencies.
You've spanked me with all the cipherin' there. That is, you're going to have to do the "Cliff Notes" version if I'm to follow you. I couldn't get past the first couple of lines.
"It can't alter the fact, that some things being variable, you, under all circumstances will want the lowest outlet temp."
well, yes, I agree, if flow rate is held constant. But lowest outlet temp does not necessarily mean highest rejection rate, I assume/think we are on the same page there. The point is and was, I can flow 20 gpm through a vertical array of tubes. I CAN"T flow it through a horizontal tube cooler of the same dimensions, because the P drop is 100 psi with the screaming velocity.
Wasn't trying to be fancy. I just answered your question. I wish I had an easier time with it.
I went back and looked where I may have confused, I said:
"Suffice to say that you can't move the oil too fast."
I mean to convey, "the faster the better". But we have plumbing limitations to worry about, pump head limitations, bypass pressure values curtailing flow...yada yada. So there is a practical limit to the tube oil velocity we can expect. More tubes (shorter also) and longer residence time in the cooler is the direction to optomise. That will, very likely, result in HIGHER outlet temps, yet higher rejection ratings, due to obtaining higher mass flow. M.dot X delta T =
I think you already get what I am saying, I just speak Greek.
well, yes, I agree, if flow rate is held constant. But lowest outlet temp does not necessarily mean highest rejection rate, I assume/think we are on the same page there. The point is and was, I can flow 20 gpm through a vertical array of tubes. I CAN"T flow it through a horizontal tube cooler of the same dimensions, because the P drop is 100 psi with the screaming velocity.
Wasn't trying to be fancy. I just answered your question. I wish I had an easier time with it.
I went back and looked where I may have confused, I said:
"Suffice to say that you can't move the oil too fast."
I mean to convey, "the faster the better". But we have plumbing limitations to worry about, pump head limitations, bypass pressure values curtailing flow...yada yada. So there is a practical limit to the tube oil velocity we can expect. More tubes (shorter also) and longer residence time in the cooler is the direction to optomise. That will, very likely, result in HIGHER outlet temps, yet higher rejection ratings, due to obtaining higher mass flow. M.dot X delta T =
I think you already get what I am saying, I just speak Greek.
- Joined
- Sep 28, 2002
- Messages
- 39,793
Hey, pal ..it don't take much with me when you get much beyond practical discriptions or analogies. I guess I really should have taken those physics courses instead of creative writing and woodshop. If I can't visualize it ...it's a very long road to enlightenment.
Yes, now that you've qualified it in a manner that I can understand it. The only gap between us is the congurencies that are, more or less, taken for granted ..and not necessarily outright expressed.
So, essentially you need a vertical flow rad cooler that's 60+" wide and 4-6" tall and about X" deep ..plumbed with #16 lines between the pump and the cooler.
Now I see why I gravitate to the hands on R&D approach. Knowing what you're dealing with ... sure puts a few obstacles
Somehow, I just feel that this will be easier in practice then it is in concept. I would keep treating this with "more" as symptoms arose.
- Joined
- Sep 28, 2002
- Messages
- 39,793
Okay ..here's one offering that may yield some possible solution.
Use commonly available liquid/oil heat exchangers and use an electric water pump to circulate PG coolant to relatively cheap (properly compact) radiators. The viscous nature of the PG won't be an issue, will require no service, and can be operated under zero pressure. The electri water pump, as used on street rods, would have to be actuated by some method ..but should last several years under intermitent service. The only item not "off the shelf" is the flange fittings to adapt to rad hose.
So ..two Laminova heat exchangers (about $400) run in parallel to reduce pressure alterations, a Moroso (or whomever really makes it) electric water pump, a properly sized (for installation along the rail(s)) cheap radiator, some machined fittings (casting would be cheaper if volume permits) ...and you should be able to hit your 60k btu figure for under your $1000 sales price for profit requirements.
Ah ..here's a 55 gpm electric with standard fittings jegs I'm sure that price could be worked down a bit there.
Use commonly available liquid/oil heat exchangers and use an electric water pump to circulate PG coolant to relatively cheap (properly compact) radiators. The viscous nature of the PG won't be an issue, will require no service, and can be operated under zero pressure. The electri water pump, as used on street rods, would have to be actuated by some method ..but should last several years under intermitent service. The only item not "off the shelf" is the flange fittings to adapt to rad hose.
So ..two Laminova heat exchangers (about $400) run in parallel to reduce pressure alterations, a Moroso (or whomever really makes it) electric water pump, a properly sized (for installation along the rail(s)) cheap radiator, some machined fittings (casting would be cheaper if volume permits) ...and you should be able to hit your 60k btu figure for under your $1000 sales price for profit requirements.
Ah ..here's a 55 gpm electric with standard fittings jegs I'm sure that price could be worked down a bit there.
- Status
Similar threads
