I wrote in the past already a VB program which integrates the Planck function over a bandwith range.
Will play with it or adapt it to find n for the Optris band as a function of the temperature
I wrote in the past already a VB program which integrates the Planck function over a bandwith range.
Will play with it or adapt it to find n for the Optris band as a function of the temperature
LDM - the basic physics means that the value of n is highly temperature dependent. MFMP verified this n ~ 3 at just one temperature (where indeed numerically n ~ 3), and if you talk to them they will say they did not validate at other temperatures because they broke the camera. That is why their high temperature results were wrong (though much less wrong than the Lugano authors).
Will investigate this further.
If you have more info how to calculate n, then I would be gratefull to receive it
Recalculating inflated temperatures when wrong emissivites where used
As a first approach to recalculate the Lugano dummy run for the case that the reported temperatures where inflated due to using wrong emissivity settings on the Optris thermal camera, we need to recalculate the temperatures to their real ones.
The procedure which can recalculate the temperatures is based on the following formula published by Optris which can be found on page 9 of their IR-basics document.
U = C · [ε Tobjn +(1 – ε) · Tambn – Tpyrn]
The meaning of the parameters is as follows :
---------U-----------The voltage from the thermal camera sensor
---------ε-----------The in band emissivity set on the Optris
---------Tobj------The temperature of the measured object in degree K
---------Tamb----The ambient temperature
---------Tpyr------The temperature of the camera sensor
---------n-----------A constant depending on the used sensor frequency band
We can use this formula for two situations, the first with the wrongly used emissivity ε1 and the accompanying measured temperature Tobj1, the second case with the correct emissivity ε2 to be used with the Optris and the correct temperature Tobj2.
Since the measured sensor voltage U is only dependent on the amount of radiation coming from the object under observation, this value is the same for both situations.
Thus we can fill in the above formula for both situation and then equate them.
This is written out below
C · [ε1 Tobj1n + (1 – ε1) · Tambn – Tpyrn] = C · [ε2 Tobj2n + (1 – ε2) · Tambn – Tpyrn]
Simplifying this gives :
ε1 Tobj1n + (1 – ε1) · Tambn = ε2 Tobj2n + (1 – ε2) · Tambn
ε2 Tobj2n = ε1 Tobj1n + (ε2 - ε1) · Tambn
Tobj2n = (ε1/ε2) Tobj1n + (1 - ε1/ε2) · Tambn
Tobj2 = [(ε1/ε2) Tobj1n + (1 - ε1/ε2) · Tambn ] 1/n
The last formula above will be used in a spreadsheet to recalculate the assumed inflated temperatures to the assumed real temperatures.
For high temperatures the term (e1/e2)Tobj1n is much larger then the term
(1 - ε1/ε2) · Tambn and the last term can in that case be discarded.
(Note that the term with Tamb can not be discarded for lower temperatues since the errors become quite large)
This leads the to the following formula also used by the MFMP to be used for high temperatures :
(Tobj2/Tobj1) = (ε1/ε2)1/n
The MFMP verified this formula at higher temperatues to be working with a value of n=3.
If you don't believe the electrical data, (such as it is) then the rest of the calculations are a waste of time.
I think that the electrical data is about right, otherwise I could not have compared the convective and radiated power with the electrical power.
If I do a recalculation based on assumed inflated temperatures, then I convert those temperatures to their assumed correct ones.
I then do a new calculation of the convected and radiated thermal power.
If this newly calculated power differs much from the electrical power, then that is another argument that the temperatues where not inflated.
I don't see where in such a calculation I don't accept that the reported electrical power in the Lugano report is about right as you suggest.
And I disagree about the fact that calculations can be a waste of time. It has given me a lot of additional insight in the issues involved. You learn from it !
But let's have a real discussion if we have some data to discuss about
The real input power, in my opinion, is very close to what was reported. I
My feeling (For what it is worth) is otherwise, so let's wait for the recalculation and see who was right
In turn, this increases the MFMP Main Tube temperature relative to input power, skewing the simulation.
The simulation I did was on a model of the MFMP dogbone. (without rods)
The remark that the simulated temperatures where near the Lugano ones must be interpreted as that they where much closer to the Lugano ones then those of the MFMP, but indeed higher (skewed) then the Lugano ones.
However those simulations did at that time not take into the latest findings which I incorporated in my spreadsheet.
So If I can find time in the near future I can redo those simulation with my latest findings. But those simulations take a lot of my time.
The reported Dummy temperatures were inflated, based on your work. Just not by much.
What is more interesting is the Dummy peak temperature. It is just below the point where emissivity-temperature errors really become apparent. Personally I doubt that is an accident.
What's important is that if they where inflated, what then the real input power must have been.
So the next thing for me to work on is making a spread sheet based on inflated temperatures and calculate what in that case the total convective and radiated thermal power would have been.
I think that the total power in that case is much lower. (But we will find out)
This simulation also shows that the MFMP thermal state report is deeply flawed in its conclusions about the dummy.
I find it a pitty that the MFMP did not cross check their measurements by calculating the convective and radiated thermal power and compare it with their electrical power setting.
If you do that you will see that the convective and thermal power is larger then their electrical power, which is indeed an indication that there is something wrong with their measurements.
Also I did in the past some thermal simulations with my thermal FEM program on my dogbone model and the temperatures the FEM program calculated where much lower then those reported by the MFMP.
For their 500 Watt run, about the power setting of the Lugano run, the simulated temperatures where close to the Lugano ones and thus much lower then those reportes by the MFMP.
What bugs me is what could have been gone wrong with their measurements.
Lugano dummy run recalculated - Final
The close agreement between applied electrical power and the confected and radiated thermal power makes it unlikely that the reported temperatures where inflated
Presented here is the, what I believe final, recalculation of the total convected and radiated power of the Lugano dummy run.
Compared to the first recalculation (post #226) the following changes have been implemented :
1. Stacked rod correction
In my first recalculation I included only on sets of rods while there are two sets.
This has been corrected.
2. Viscosity of air
In the calculation of the convective heat transfer coefficients I had wrongfully used the dynamic viscosity of air instead of the kinematic viscosity of air.
This has been corrected and all convective heat transfer coeffcients in the spreadsheet have been updated.
3. Convective heat transfer correction factor for the ribbed area
Due to having no spacing between the ribs, the convective heat transfer of the ribbed area is less efficient. Thus) a correction factor needs to be applied. This correction factor was, using CFD simulation (see post #389) , found to have a value of 0.752
The spreadsheet now includes this correction for the convective heat of the ribbed area.
The updated spreadsheet is included as an attachment to this post.
For information on additional changes compared to the original calculations of the Lugano report see the comments in the original post #226
The outcome of the spreadsheet calculation is that the difference between applied electrical power and the total convected and radiated thermal power is 1.62 %.
If the reported temperatures for the Lugano dummy run where inflated, then the difference would have been much larger. As such it is unlikely that the temperatures where inflated
Do you mean that with everything else unchanged, a tube would run hotter with the fins (with a 2 cm valley bottom diameter), than a plain bare tube with a 2 cm diameter, notwithstanding the apparent extra surface area caused by the fins using the fin design used for the Lugano device?
For a bare tube the convective power can be calculated as :
Q = h x At x (Ts - Ta)
---------h--------Convective heat transfer coefficient
---------At-------The area of the tube
---------Ts-------Temperature of the surface
For the finned tube the power is calculated as :
Q = C x h x Af x (Ts - Ta)
---------C--------Convective heat transfer correction factor for the finned tube
---------h--------Convective heat transfer coefficient
---------Af-------The area of the fins
---------Ts-------Temperature of the surface
C, the correction factor has a value of .752
h is for both cases 13.28
At, the area of a bare tube of 20 mm diameter and 200 mm length is 0.0125 m^2
Af, the area of a finned tube is 0.0263 m^2
For the surface temperature we take 445 degree C
For the ambient temperature we take 21 degree C
The calculated powers for both cases are then
Qt-------------70.75 Watt for the bare tube
Qf------------111.36 Watt for the finned tube
Clearly with fins the dissipated power by convection is larger and thus the temperature lower.
As an additional comment : C x h = 9.99, the convective heat transfer coefficient I derived for the finned area.
Hope this explains it for you.
Lugano ECAT convection correction
Due to the close spacing of the ribs of the Lugano ECAT, the convective heat transfer of the ribs is less effective. Thus there needs to be applied a correction to the calculated convective heat transfer of the ribs.
Without a correction the convective heat transfer is calculated based on the calculated convective heat transfer coefficient for the base tube with a diameter of 20 mm.
The convected power is then calculated by multiplying the convective heat transfer coefficient of the tube by the area of the fins and the difference temperature of the surface and the ambient temperature. The formula :
Q = h x Af x (Ts - Ta)
---------Af-------The area of the fins
---------Ts-------Temperature of the surface
However due to the fins being close to each other and their respective convective heat flows interacting, the convective heat transfer will be less effective and will result in a lower convected power.
Evaluating the above formula it means that the convective heat transfer coefficient has a lower value then one would expect when calculating it based on a bare tube. (which is the basis for the calculation in the Lugano report). The question now is how to determine the correct value of the convective heat transfer coefficient for a finned area as used on the Lugano ECAT.
The method followed was to calculate with the CFD software the convected power for two different section lengths of a finned area, the second section having halve the length of the first section.
The sections where simulated with a temperature of 445 degree C and an environmental temperature of 21 degree C.
Subtracting from the power of the longer section the power of the halve section yields the power of a halve section.
Since for both simulated sections the convected power of the sides at the ends is equal, this power of the sides is then automaticcaly removed by the subtraction and only the power of the finned area is calculated in this way.
Deviding the calculated convected power of the halve section by the fin area of the halve section and by the difference between section temperature and ambient temperature gives the convective heat transfer coefficient.
This convective heat transfer coefficient obtained in this way was found to have a value of 9.99
Since the value for a bare tube of 20mm diameter, tube temperature of 445 degree C and an ambient temperature of 21 degree C is 13.28 the convective heat correction factor becomes :
-------------------------Correction factor = 9.99/13.28 = .752
The conclusion is that due to having the fins next to each other without additional spacing in between, the convective heat transfer of the finned area is lower then one would expect based on the calculation in the Lugano report.
Simulating convection with CFD - update 2
In my previous post I wrote :
Next step now is to do some additional simulations to gain more insight on the effect of using higher levels of meshes (smaller mesh sizes) instead of using a boundery layer. This since if the smallest mesh size is less then the thicknes of the first (smallest) boundery layer, the simulation results are expected to be to be as good as with a boundery layer. (But the calculation overhead much larger due to the greater amount of meshes)
Thus I started simulating the bare tube again using meshing levels 1, 2, 3 and 4 and without a boundery layer. A higher messing level then 4 was not possible due to exceeding the maximum number of meshes allowed for my version of the program.
I did not do a rerun for level 1, since that was already simulated and the result presented in my previous post. For all 4 meshing levels a cross sectional view was made. These are presented in the following figure.
Each case was again simulated with a tube temperature of 445 degree C and an environmental temperature of 21 C.
From the reported heat flux for each simulated case the heat tranfer coefficient was calculated.
The results can be found in the following table :
Level-----------heat transfer coefficient-------------deviation (%)
---3------------------------13.03--------------- ----------------- -1.9
The reported deviation is the deviation from the value calculated with the method used in the Lugano report, the value being 13.28
For the smallest grid simulated, the deviation from the value derived from the calculation method used in the Lugano report is only -0.3%
Level 4 simulation with one layer per mesh size
Meshing a ribbed area requires many more meshes then for a bare tube as used in the simulations above.
This is because the area between the ribs must be meshed with a small mesh size for proper simulation of the convective heat between the ribs.
In order to reduce the number of meshes in order to stay within the limits of my program I deceided for the bare tube to do an experiment with one layer of meshes per mesh size instead of the 4 layers as used in the examples shown in the figure above. This reduces the number of total meshes.
The result of that simulation was that the calculation of the convective heat transfer coefficient had for this case a value of 13.28, the same as the value calculated with the method in the Lugano report.
My conclusion is that a level 4 simulation without a boundery layer will give results very close to the values obtained with the method used in the Lugano report.
Simulating convection with CFD - update 1
As a first approach to simulating natural convection for a ribbed tube I started with simulating convection of a bare tube.
I used the bare tube case since calculating natural convection for a bare tube is a standard case which is well researched and as such the calculation of the convective heat transfer coeefficients are near accurate using established formula's such as the ones used in the Lugano report. This makes it possible to compare the value of the calculated heat transfer coeffcient with those of the simulations.
The case simulated was for a bare tube with a diameter of 20 mm. Tube temperature was set to 445 degree C (About the central temperature of the Lugano ECAT during the dummy run) and an environmental temperature of 21 degree C.
For the meshing of the fluid (Air) a level 1 grid was used (2 seizes of meshes). For the meshing of the tube a level 1 was used (2 seizes of meshes).
Then the heat flux from the tube by natural convection was simulated for two cases, the first being with the original derived mesh, the second with the original mesh but with an added boundery layer mesh at the surface of the tube for improved simulation accuracy of the convective heat transfer.
A cross sectional view of both cases is shown in the following figure.
The properties of the fluid (Air) where simulated with the janaf formula's for the thermodynamic properties of air and with the sutherland approach for the transport coefficients.
(I had found somewhere that using these gives the best results when calculating convective heat transfer)
First I calculated, following the method outlined in the Lugano report, the convective heat transfer coeffcient for the tube. For the mentioned temperatures the coefficient had a value of 13.28
For the case without the boundery layer mesh I calculated back from the total heat flux from the tube as reported by the CFD program a convective heat transfer coefficient of 10.82
For the simulated case with the added boundery layer mesh the obtained value was 13.74, the difference with the Lugano calculation method being 3.7 %.
The conclusion is that even with a coarse mesh grid for the fluid, but with the addition of a boundery layer mesh, the calculated value of the convective heat transfer is close to the value obtained by existing formula's such as the ones used in the Lugano report.
Maybe the accuracy can be improved somewhat more by adapting the janef and sutherland material coeffcients such that they give a closer match in the temperature range of the simulation.
Currently the error is in my opinion quite acceptable, also taking into account that a convective heat transfer correction factor for the ribbed area of the ECAT will be based on a ratio calculation.
Since for both a bare tube and a ribbed tube for the same temperatures, the deviations of thermal coefficients used by the CFD program will be equal for both cases, the ratio calculation will limit the error somewhat.
Next step now is to do some additional simulations to gain more insight on the effect of using higher levels of meshes (smaller mesh sizes) instead of using a boundery layer. This since if the smallest mesh size is less then the thicknes of the first (smallest) boundery layer I used in the example above, the simulation results are expected to be to be as good as with a boundery layer. (But the calculation overhead much larger due to the greater amount of meshes)
That would indeed be a possible counter measure
However radar antenna's and/or related waveguides tend to be tuned to the radar frequency, so that limits the frequency span which you can use.
Also current processor power likely (But I don't know about recent developments) probably allows to follow the frequency hopping in real time.
But as alway, there are counter measures, counter-counter measures, ....., etc.
Developping these keeps some people of the streets
What are you babbling on about now Adrian? Thinking an aeroplanes manoeuvrability has any bearing its capabilities, shows that you are at least 25 years behind the curve.
And where did you hear about the specifications of the SU57s sensor and stealth technologies?* Off Putin? Or one of those strange low-brow right-wing near-conspiracy websites you keep posting links to?
And look at the big exhausts sticking out the back of the SU. Ever heard of radar cross-section? What a joke. It’s probably one stinger missile away from an embarrassment.
I’d stick to furnace design (and Rossi worshipping) if I were you pal.
* Or the F-22’s, for that matter.
While I agree often with your technically insights, I think this time your comment are made too rash
There are a lot western defense technology journals where you can find information on the SU57
If you had read them you would know that the SU57 relies less on its reduced radar cross section but more on active changing the radar signal aimed at the aircraft. This active manupulation of the radar signal makes the SU57 appear on the the radar of the F22 at a total other location then where the aircraft is.
You can look at it in this way : the radar signal of the F22 is reflected in circles, like the circular waves in the water when you drop a stone in it.
At any poin on a receiving wave (And thus also at the radar reciver) the direction of the target is the normal to the wavefront.
By actively giving properly timed/phased signals on the same radar frequency back you can distort the wavefront at the reciver of the targeting radar, thereby modifying the position/direction where the receiver thinks where the target is.
And you can't do anything about this. ( I have done military research in this area)
Much more important then my simulations is that You and Russ continue to make progress.
Hope to see that reported on the atom-ecology thread when I am away !
Simulating convection with CFD
Since my last post I have been spending time on how to simulate with my CFD software natural convection. There are two major parts which determine a correct outcome of a CFD simulation.
These are :
1. Quality of the meshing grid
As is stated in several articles about CFD simulation, the quality of the simulation result is largely dependent on the quality of the simulation mesh.
A higher messing level will give a better results then a more coarse meshing grid.
A level 0 grid is the base meshing grid you define. Each next level reduces the x, y and z dimensions of a grid cell by a factor two. Thus a level 3 cell will have x, y and z dimensions which are 2^3 = 8 times smaller then the cell dimensions of your original grid.
However a small grid increases the amount of meshing cells beyond the amount allowed for my version of the CFD software. So I am currently limited in what I can do.
2. Boundery layer
For convective heat transfer, the interaction between the solid part you are simulating and the fluid (Air) is largely occuring in the so called boundery layer near the surface of the solid.
If the convection is upward then at the lower point of the solid, the boundery layer is near zero.
It becomes wider upwards till at the highes point it has it's maximum width.
A rough estimate is that the maximum width of the boundery layer is the characteristic length raised to the power 1/4.
For the characteristic length (without fins) of the central part of the Lugano ECAT we can take the diamer, being 20 mm. This gives an estimated maximum boundery layer thickness of 20^(1/4) = 2.1 mm.
The CFD software can create a special boundery layer consisting of several meshing layers parallel to the surface of the solid for improved convection simulation accuracy.
Only very recently I found out how to properly create such a mesh boundery layer with my CFD software.
The idea is now to proceed with several simulations of a bare tube , each with other meshing settings and different boundery layers.
The hope is to arive at a mesh quality which gives about the same convective heat transfer coefficient as determined by existing formula's for a bare tube.
If we succeed we then can proceed with the simulation of a ribbed area and determine a correction factor for the convective heat transfer coefficient of the finned area of the Lugano ECAT.
However I have currently some doubts if there will be a result which I have enough confidence in to be presented. (And which will pass some sanity checks).
It will take some time before I start simulating again since I am off to Finland for some weeks
Jed, while I agree with you about the lack of solid data from Jiang et al, this thermocouple stuff is a canard I thought had flown away. I have never known a faulty thermocouple read high, they either read low or don't read at all. Resistance thermometers may read high under some circumstances, but while I have tortured many hundreds of thermocouples to death, they never read high even in extremis.
May I disagree ?
Some thermocouple circuits have a high ohmic pull-up resistor, which when the thermocouple gets broken, cause a high voltage at the input. (We used 10 MegaOhm resistors)
This will result in a high temperature displayed when a thermocouple is broken and will be an alert that there is something wrong.
The high value displayed depends on the electronics and software, but will normally so much different from the expected value, that you will not interpret it aa a valid measurment.
Something like this perhaps?
The link is for longitudinal fins, not for anular fins.
Also the shape of the fins in the link is rectangular, not triangular as in the Lugano case.
So I don't think we can apply the numbers of the calculator to the Lugano case.
It seems to me that the valley of the fins must convect much more poorly than the outer parts of the fins, due to the rarified air in that area, due to the increased heat and lesser air flow possible in that valley bottom region. So to some degree that surface area is not as efficient as the increase in physical area due to the ribs might suggest.
That reason for a lower convection rate was also mentioned in literature as the reason for the total convective heat transfer dropping off below the optimum fin spacing.
If I remember well, you mentioned this before and you where right.
My plan now is to make a short ribbed section ( 5 cm) in CAD and import that in the CFD software I am using.
I can then estimate how much finite element sections I run short off.
Maybe I can then negotiate a deal with the CFD software vendor.
I had contact with them in the past and they where at that time willing to negotiate customized deals (eg lower price)
In the meantime I am open to any suggestions where to find literature on the effect of fin spacing which can give the answers we need.
I built a simple Lugano-esque dummy device in 5 days, most of which were curing time for the mold and cement, and tested it in a couple of hours.
Did you ever publish your measurements results of your replica ?
What where your conclusions ?
Convection of the Lugano ECAT rib area
I updated the convective heat transfer coefficients on my Excel sheet for the Lugano dummy run recalculation
The Error is somewhat lower now, about 9.1 %
With that error I expect that if we consider that the temperatures where inflated, the error will be much lower and close to the zero percent.
Whatever the outcome will be, it seems that the correct calculation of the convective heat transfer of the ribbed area will be the factor that determines the correctness of the calculation.
That because even after correcting my heat transfer coefficients, there is still a large difference between the convective heat transfer of the ribbed area in my calculation and that of the Lugano report.
During the recalculation, looking at the contents of the Lugano report and consulting literature, new questions about the convective heat calculation of the rib area arose.
1. Area used for the calculation of the convective heat transfer
The Lugano report mentions 61 ribs for the finned area. (You realy can count 69 from the picture !)
Since we have 10 sections, each section then contains 6.1 ribs
The Lugano team calculated the area of the ribs as Af = 2x Pi x(Ra^2 - Rb^2), Ra being the diamer of tube + fins ( 12.3 mm) and Rb the diameter of the tube (10 mm).
This calculation gives an area of the fin of 3.22E-4 meter
For 6.1 fin per section the total fin area of a section becomes 6.1 x 3.22 E -4 = 1.96E-3 m^2
This is a large difference compared to the section fin area I calculated as being 2.63E-3 m^2 (34% more)
To make things worse, the Lugano team rounded the value of 6.1 ribs per section to 6 in their calculation of the convective heat transfer of a rib section which lowers the calulated heat transfer even more.
The Lugano team referred to the "Heat Transfer Handbook " For using this approximate formula for the rib area.
Indeed does that book use that formula and refers to another work for a more detailed explanation, the book "Extended surface heat transfer".
However I could not find that explanation in the last reference.
2. Fin spacing
Convective heat transfer from fins are normally calculated for one fin, For multiple fins you can multiply the convective heat transfer of one fin with the number of fins if the fins are widely spaced.
If the spacing gets closer the total convective heat transfer rises due to the effect of having more fins. That is until we arrive at the optimal spacing , where the total convective heat transfer will be at it's maximum. If from the optimal spacing you lower the distance between the fins even more, then the total convective heat transfer drops off from it's maximum, and quite rapidly.
Thus the convective heat transfer coefficient of multiple fins is not linear dependent on the number of fins, but also depends on the spacing and is quite non linear near the optimal spacing value.
For the Lugano ECAT I don't know what the value of the optimal spacing is and if the spacing of the fins is less, equal or grater then the optimal spacing.
Even if we know the value of the optimal spacing it will not give us the needed heat transfer coefficients for a recalculation.
The conclusion is that for an accurate Lugano dummy run recalculation we have to find out how to calculate the correct heat transfer coefficient of the finned area.
This means determining which area to use and also to find out if we need a correction for the used fin spacing.
It looks that that there is currently at least one option to calculate the proper heat transfer coefficients for the ribbed area.
That is to model a finned section and to simulate that section with CFD software.
My CFD software is too limited (restricted maximum number of finite elements I can use) to undertake that task.
Maybe I have to consider upgrading my license (Quite expensive).
Convective heat transfer correction
As stated in my previuos post I saw a large difference between the convective heat transfer coeeficient presented on page 16 of the Lugano report for zone 5 of the rib area on page 16 and the one I calculated for zone 5.
The Lugano report gave a value of 12.75 for zone 5 while i calculated (for the slightly lower corrected zone temperature) a value of 14.52.
Since page 16 the Lugano report gives the values of all the parameters from which the convective heat transfer coefficient of zone 5 was calulated I could compare those with the values I used.
Both are given in the following table :
If we compare the values then it can directly be seen that there is a large difference in the values of the viscosity v.
It turns out that I wrongfully used the tabulated values of the dynamic viscosity instead of the kinematic viscosity .
From the values it can also be seen that the Lugano team used for the expansion coefficient the value 1/T (T being the film temperature) instead of the real tabulated values for air.
As a third remark it can be seen that they used only two digits accuracy as can be seen by the trailing zeros of their values. As alreade mentioned in an earlier post these roundings can add up to inceased errors in the calculations.
I now have to update, based on the kinematic viscosity, all convective heat transfer coefficients used in the Excel sheet of the Lugano dummy run recalculation.
If we, for sake of discussion, suppose that the recursive method did exaggerate the dummy reported rib area temperature, and so reduce it to the temperature it might have been without the exaggeration, might then re-calculated output and reported input become much closer together? I propose about 380 C, rather than about 450 C, simply by looking at table 2, and comparing to the value demonstrated there when using an emissivity of 1.0 . (The temperatures of the other parts can be left the same for now).
As I already stated when I presented the results, this is indeed something I planned on doing.
However such a recalculation does not make sense if we don't have a good explanation for the large increase in convective energy of the finned area.
I need to know if there is an error in that calculation or not.
What I already see is that there is a large difference between my convective heat transfer coefficient and that in the report.
So I have to investigate this first before doing that next step.
But I have to go slowly since my concussion is playing up.
Lugano dummy run recalculation compare
This post gives in a table the difference between the published Lugano dummy run data and the recalculation I did.
(I had planned to make this comparison sooner, but fell and now have a concussion.)
The difference can be seen in the following table
I have these comments concerning the differences :
Radiation for the rods is almost the same, however the convection is lower due to the lower value of the stacking correction factor for the convection.
The radiation of the ribs is much higher in the recalculation then in the Lugano report.
This is due to the increased emissivity because of the reflection berween the fins (infinite refelction method) an dthe larger area (real area times the view factor to the background)
The convection is also significant higher an this is largely due to the difference in area's used in the recalculation and the Lugano report.
In the Lugano report a fin area of 2 x Pi x (Rf^2-Rt^2) is used, Rf being the radius of the top of the fin (12.3 mm) and Rt the tube radius (10 mm) giving a fin area of 3.22E-4 squared meter.
In the recalculation a fin area of 3.81E-4 squared meter was used, based on the real fin area calculated by substracting the area of two cones.
This gives an increase of 19% in area but thsi difference is not enough to explain the increase of almost 50 % in convection for the rib area.
For the end caps the differences are minor.
Most of the differences between the original Lugano dummy run data and the recalculation can be explained.
However the reason for the large difference in convective heat transfer of the rib area is not known. This may need some further investigation.
Dummy run recalculation correction
I have to admit that there was an error in the spreadsheet I published.
The power of the rods was not multiplied by 2, since we have two sets of rods.
The new calculation now shows an error of 12 % (see adapted spreadsheet)
It will now be very interesting to see what the calculation with the inflated temperatures is going to bring us.
In addition to the above calculation I can say that some preliminary FEM simulations of the dogbone i did in the past showed that the calculated temperatures where already in close agreement with those reported in the Lugano report.
If the temperatures in the report where inflated then the FEM simulations should have shown lower temperatures then those in the Lugano report.
That was not the case.
Lugano dummy run recalculated
The dummy run recalculation of the Lugano report presented here is based on the assumption that correct emssivities settings where used for the Optris temperature measurements and thus that the reported temperatures for the dummy run in the report where correct.
The spreadsheet with the calulations is included to this post as an attachment.
Where possible I have made the calculations dependent on the values of variables in the spreadsheet so that changes in those values can be evaluated by the reader.
I intend to make the same calculations also for the case that the Lugano testers used wrong emissivities on their Optris thermal camera. If those calculations become available will publish them also in this thread.
The Lugano dummy run recalculation did include the following calculation changes when compared to the original Lugano report. Additional information on these calculation changes can be found in earlier posts in this thread.
1. Stacked tubes convective heat correction factor
For the stacked tubes convective heat transfer a correction factor of .561 was used instead of the factor .667 (2/3) in the Lugano report.
This new correction factor was obtained by analysing information obtained from a photograph of the Lugano set up. based on the photograph it was calculated that the rods where equilateral stacked with a distance of 1 mm between the tubes.
Simulating a single tube and the three stacked tubes with CFD (Computional Fluid Dynamic) software yielded that the correction factor for the tube stacking is approximate .561
2. Thermal expansion coefficient
For the thermal expansion coefficient of air tabulated values where used instead of the ideal gas law 1/T formula. The thermal expansion coefficient is used for calculating the convective heat transfer exchange coefficient. (Note that on line calculators most of the time also use the ideal gas law formula instead of the tabulated values with the result that these values can be substantial lower then the calculated ones in this document)
3. Fin area radiated heat transfer
The radiated heat transfer of the finned area was calculated based on the real fin area multiplied by the view factor to the background
The view factor between the fins was calculated using the NIST program View3D as being .428 Thus the view factor to the background is 1 - .428 = .572
4. Finned area emissivity correction
The emissivity of the finned area was corrected by a factor calculated from the formula as derived from the infinite reflection method.
5. Fin area temperature correction
The Lugano testers used for the their radiated heat calculation of the finned area of the ECAT the area of the bare tube. They seem not to have been aware that they should have used the view factor in combination with the total fin area. As a consequence they must not have know the effect of the view factor on their calculations, both for the area and the emissivity.
Not taking into account the effect of the view factor on the emissivity, the measured temperatures given by the Optris of the finned area are somewhat higher then the real temperatures.
Therefore the temperatures measured by the Optris of the finned area need to be corrected to their somewhat lower real values.
6. Finned area power
The power of the finned area is calculated with the adjusted temperatures (see point 5) , corrected emissivities and effective area.
7. Fin efficiency for convective heat transfer
The fin efficiency correction was omitted since the average Optris temperature of the measured areas already took care for this correction.
8. Joule heating
The Lugano test team calculated the Joule heating as being 6.7 Watt, but rounded this to 7 Watt. They also calculated that since the interconnections cables extended for a short distance into the rods, that the Joule power generated in the rods was .4 watt.
Thus the external Joule heating was 6.7 -.4 = 6.3 watt. This is the value used in the recalculation.
The outcome of the recalculation is that for the calculated input power of 479.7 Watt of the dummy run, the calculated output power from convective and radiated heat is 538.17 Watt , the difference being
Additional remarks :
A. Broad band emissivities
Broad band emissivities where taken from the same curve as presented in plot 1 of the Lugano report.
B. Heat transfer coefficients
Heat transfer coefficients where calculated using the same procedure as outlined in the Lugano report. However tabulated values for thermal expansion coefficient where used. (point 2 above)
Update juli 5th 2018
New revision since the original version calculated the power of the rods wrongly because it considered only one set of rods instead of two.
Attached Excel file also updated
All this for that therefore thanks to your work. My own estimation is nickel particles couln't be melted and AR uses a trick to do a large amount of H monoatomic ( as said Focardi) as Japanese researchers done by adding a few Pd atoms onto Ni particles ( spillover effect).
I am totally an amateur in the LENR mechanisms.
However if Palladium was added, it should, having mass number 106 shown up in the fuel analyzis of the Lugano ash. And I do not see a clear signal above the noise.
My guess would be that Germanium (mass number 73) was added to create the monoatomic H.