How do you explain this?
Ha! I did not recall any of that. Merciful amnesia perhaps. In any case, none of it made it into the Proceedings.
How do you explain this?
Ha! I did not recall any of that. Merciful amnesia perhaps. In any case, none of it made it into the Proceedings.
[...] In any case, the experimental evidence shows that Tin was asymptotically approaching Tamb. There is no possibility at all that this very specific trend can be induced by a heat flux emanating from the Ecat. It is the clear sign that the temperature of the water inside the tube is going to equalize the temperature of the surrounding ambient, and this can happen only if the water is still. This is the more simple, straightforward and congruent explanation.
I haven't still understood if you exclude it, and, in case, why, and which specific alternative explanation you propose.
I don't know whether there's "no possibility at all", but I'm afraid that there's not enough information in the report to confidently state that it certainly didn't.
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.
See below.
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).
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 ?
Did you ever publish your measurements results of your replica ?
What where your conclusions ?
Thanks for your hard work. I am currently in an area of very poor internet, and soon to be where there is none, so I will be brief.
I was mainly interested in building the IR emissivity tester to a reasonable level of ease of testing. The results are those for my plain cylinder that I have posted some info from previously, and several flat slabs. I will attempt a ribbed version some time later in the summer, and attempt to rent an IR camera with suitable abilities. I do have stored thermocouple data, pages of notes, etc., which I will compile now that I have fewer distractions in the evenings.
For the finned calculations, I don't think that I can help at present. I have been attempting to look up some better information and equations, but it is nearly impossible at present. I can barely read a page of the Forum once an hour, and sometimes barely post. Luckily the Forum software stores my posts so I can re-send until they go through.
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.
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.
To summarize my conclusion: The erroneous use of the IR camera emissivity function resulted in nearly doubling the numerical value of the true temperature, which results in nearly 4 times the calculated power compared to the real power emitted by the Lugano device, due to the T^{4 }relationship of temperature to radiant power. The ratio of false radiant output to real radiant output increases with the overall temperature, independent of the overall physical size or input power of the device measured, and is directly dependent of the emissivity delta between the real (high) and artificial (low) emissivity functions used for the IR camera, for any equivalent isothermal surface area. (IE: The addition of caps, rods, or other areas of cooler temperature compared to the maximum temperature zones reduce the combined apparent ratio of output to input, while for each individual area (segment) the ratio is directly dependent on the temperature and emissivity error delta only.) Additional areas of lower temperature obscure the T^{4 }inflation relationship when the total input and combined radiant output of several areas of significantly different temperature are compared. The inflated "COP" effect can be optimized by simplified architecture (one diameter, reduced non-radiant losses), higher (real) temperature, and increasing the delta between the (low) camera emissivity function and (high) real camera emissivity function. The general idea is to maximize the inflated T^{4 }power relative to the real T^{4 }power, resulting in the T^{4 } of the inflated temperature's numerical values to increase at an accelerated rate compared to those of the T^{4 }of the real temperature, as the real temperature increases, thereby systematically and logarithmicly increasing the numerical value of the apparent and false COP with increasing temperature. A false "COP" of 5 is achievable without much difficulty.
Thanks Alan,
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.
Ha! I did not recall any of that. Merciful amnesia perhaps.
Yes, most likely. It's probably a selective amnesia, triggered by topics connected to the Bologna demo (1). But don't worry, fortunately the Vortex mail-archive helps us to refresh our memory.
Btw, your reportage from ICCF16 is really very interesting. It shows that the results of the Bologna demo have been widely discussed between you, Melich and Storms, and considered "a definitive triumph" even in case "there was a only a tiny bit of steam".
Your reportage also mentions Test 1: "Levi remarked somewhere that he felt confident in the machine after the Dec. 16 test [Test 1] and also when he saw it run with no input, in heat after death." As you can see in this same thread, for some days I'm discussing with *can* on the interpretation of the few experimental evidences that appear in the calorimetric report, including the photo with temperature curves taken from a PC screen, mentioned also in this phrase of yours: "The data acquisition system failed, as noted by Levi in his report, which is why they had to use a photo of the screen."
Until now there are two possible interpretations (2). The one described in the official calorimetric report, which states that the Ecat produced almost 10 kW, initially with an input power of 1120 W, and, later, in self-sustaining mode during the last 15 minutes. My alternative explanation is that in the middle of the test the water flow has stopped. Well, since you have some experience in calorimetry and T-probes, can I ask you which one of these two interpretations appears more realistic to you, or if you have a third possible alternative to suggest?
You also wrote at the end of your reportage: "People such as Melich and Levi, who know the most about this machine, seem to have the highest confidence that it is real." These are very important opinions, considering that they both have a PhD in physics, and have long been teachers at high-level universities. In addition, the first is one of the maximum expert in LENR, while the second was a member of the American Skeptic Society, and was trained in LENR by one of the fathers of the Ni-H technology. They were probably the most talented and informed members of the two groups A and B (3) that in January 2011 collaborated in the drafting of the UniBo calorimetric report. So, I would ask you, could they have been cheated for so long by someone like Rossi on the calorimetric performances of an alleged LENR device?
(1) https://www.lenr-forum.com/for…D/?postID=30161#post30161
I don't know whether there's "no possibility at all", but I'm afraid that there's not enough information in the report to confidently state that it certainly didn't.
I wonder which kind of information would you have expected from a document where it is reported that "… the original data has been lost …".
Sorry, I have no more argument for you. If the temperature graph of the Decembre 16 test is not convincing enough for you, I'm afraid that I will not be able to find more convincing arguments for showing you the flaws present in the January 14, and February 10 tests.
Thank you for the all the nice graphs you have prepared following my indications.
To clarify the wording above: I did acknowledge that the trend in the graph could indeed be interpreted as the water flow getting interrupted and that there's not enough information in the report allowing to tell that it [the flow] didn't [get interrupted].
No need to shut the door angrily.
To clarify the wording above: I did acknowledge that the trend in the graph could indeed be interpreted as the water flow getting interrupted and that there's not enough information in the report allowing to tell that it [the flow] didn't [get interrupted].
No need to shut the door angrily.
Sorry, your wording was cryptic. I thought that "it didn't" meant "the report had not enough information for concluding that the flow stopped". Consequently, I was really discouraged, not angry at all.
However, if you acknowledge that in the December test the water was stopped, it also means that at least one of the testers was aware that the Ecat was unable to operate as claimed. Now, if you wish, we could go on with the subsequent tests.
QuoteTo summarize my conclusion: The erroneous use of the IR camera emissivity function resulted in nearly doubling the numerical value of the true temperature, which results in nearly 4 times the calculated power compared to the real power emitted by the Lugano device, due to the T^{4 }relationship of temperature to radiant power. The ratio of false radiant output to real radiant output increases with the overall temperature, independent of the overall physical size or input power of the device measured, and is directly dependent of the emissivity delta between the real (high) and artificial (low) emissivity functions used for the IR camera, for any equivalent isothermal surface area. (IE: The addition of caps, rods, or other areas of cooler temperature compared to the maximum temperature zones reduce the combined apparent ratio of output to input, while for each individual area (segment) the ratio is directly dependent on the temperature and emissivity error delta only.) Additional areas of lower temperature obscure the T^{4 }inflation relationship when the total input and combined radiant output of several areas of significantly different temperature are compared. The inflated "COP" effect can be optimized by simplified architecture (one diameter, reduced non-radiant losses), higher (real) temperature, and increasing the delta between the (low) camera emissivity function and (high) real camera emissivity function. The general idea is to maximize the inflated T^{4 }power relative to the real T^{4 }power, resulting in the T^{4 }of the inflated temperature's numerical values to increase at an accelerated rate compared to those of the T^{4 }of the real temperature, as the real temperature increases, thereby systematically and logarithmicly increasing the numerical value of the apparent and false COP with increasing temperature. A false "COP" of 5 is achievable without much difficulty.
Yes. Very very nice job.
But all of it would have been moot if only correct calibration with a reliably known power input to a blank reactor would have been performed at the time, over the entire relevant temperature range. Somehow, of all the renown scientists involved with Rossi, none seemed to know or care about the concept of calibration and blank runs. Truly strange.
Here's a Plot of the last test I reported on: Thermocouple recording and Lugano Iterative Method emissivity T line (yellow).
The second Plot has additional Lugano Iterative Method values added, retrieved from the prior test with better low T control.
I want to remind everyone about my tests with the Prominent Gamma L Diaphram pump
Prominent Gamma/L 0232 Flow Rate Test (final results)
a) The specification is the MINIMUM flow at 2 bar -- NOT the maximum. (Sorry, Jed)
I don't agree with this. I believe that the information (32 l/h) shown on the back label of the Doral pumps give the maximal rate of pumping at which accurate dosing is guaranteed by the company. Higher flow rates are possible, as you found, but the pump's accuracy declines at these flows. I don't think there is any sense in which this shows a minimum flow rate.
d) Combining b and c gives a multiplier of 2.2 over the specification
Yes but late in the Doral test (particularly in November and December of 2015) a flow rate of more than 80 l/h is required for the remaining working pumps to supply the daily pumped volume claimed in the Penon report, and you are still way under that.
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
It will take some time before I start simulating again since I am off to Finland for some weeks
Safe and happy travels, we look forward to more careful analysis upon your return.
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 - 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)