Expansion noise - capillary tube into evaporator

[desA]

If you are breaking into the system,
Install pressue point at condensor outlet(cap inlet) and at cap outlet (evap inlet)

Thanks for that, MF. Good idea. Then we can get a good handle on exactly what is going on across the cap tube, as well as across the condenser coil & evap.

[Chef]

If you post the compressor volumetric flowrate we can then begin to see what the cap is doing and if the original length is right or your new numbers are right.

There are of course a number of tests you can do before breaking the system to see how close your original cap is.

Chef

[desA]

If you post the compressor volumetric flowrate we can then begin to see what the cap is doing and if the original length is right or your new numbers are right.

I'll write up the compressor details tonight. Excellent point.

There are of course a number of tests you can do before breaking the system to see how close your original cap is.

I agree - absolutely. I would like to push the current arrangement, as it stands, to its absolute limit. Once we've reached the end of what this can deliver, we can modify the piping. The final cut is so terribly terminal, isn't it?

[desA]

Compressor details

Rated performance - at 220V, 50Hz
Capacity.... : 910 W
Input power.: 386 W
COP.......... : 2.36
Current.......: 1.80 A

Rating condition
Evaporating temp : 7.2'C
Condensing temp.: 54.4'C
Ambient temp.....: 35'C
Return gas temp.: 35'C
Liquid temp........: 46.2'C

Compressor & motor data
Bore, stroke.......: 25 mm / 18mm
Displacement......: 8.83 ml/rev

Nothing is mentioned about the rotational speed.

In Coolpack at rated conditions, a thermodynamc balance is obtained at Vs=1.25 m3/h.

[Chef]

Could either 1450 or 2900 RPM and the latter seems close your 1.25

Chef

[mad fridgie]

Could either 1450 or 2900 RPM and the latter seems close your 1.25

Chef

I also expect high speed, but at high speed there is normally motor slipage, I would work on 2800 RPM
1.48M3H( also checked) i suspect volumetric efficiency is down at the stated compression ratios, hence the lower than expected performance. or even more slipage

[desA]

The thermodynamic balance at manufacturer's rated specs, providing Vs=1.25 m3/h (Coolpack - note, COP's also correspond at this balance point), would imply a slippage of around 18.67% against maximum possible delivery.

Perhaps the manufacturers are just erring on the side of caution?

[Chef]

The thermodynamic balance at manufacturer's rated specs, providing Vs=1.25 m3/h (Coolpack - note, COP's also correspond at this balance point), would imply a slippage of around 18.67% against maximum possible delivery.

It seems quite a lot 18% for slippage - maybe some motor experts can chip in with whats more common?

Still even at 2900RPM which is the basic industry standard as I believe it and some volumetric efficiency and the fact we do not know exactly how Coolpack calculates it I would propose using 1.48m3Hr until further clarification.

At least we can begin to run some models and see where things lay with a common data base but of course its your call.

Chef

[desA]

It seems quite a lot 18% for slippage - maybe some motor experts can chip in with whats more common?

Still even at 2900RPM which is the basic industry standard as I believe it and some volumetric efficiency and the fact we do not know exactly how Coolpack calculates it I would propose using 1.48m3Hr until further clarification.

At least we can begin to run some models and see where things lay with a common data base but of course its your call.

Chef

That's fine. We can work with that & see what comes out of the calcs.

[chemi-cool]

2900 rpm is after slippage calculation, without it it would be 3000.

The same with 1450 rpm, should be 1500.

[Chef]

I have run a cap tube estimation program for this diameter cap tube. At mid range it estimates a required length of 1.2023 m & at hot condition, a length of 1.7517 m.

It would seem that MF & Chef were probably correct in estimating that the current cap tube is too long.

I have back-checked the standard cap tube length for the compressor used. This length (at dia) (2.246m) corresponds to the rated compressor condition of Te,sat = 7.2'C, Tc,sat=54.4'C, Tc,suct=35'C, SC=5K.

What is normal procedure for reducing cap tube length?

I'll also reset the liquid piping arrangement to flow downwards into the cap strainer.

You have back checked the tube length is 2.246m with discharge pressure (Pdis) of about 14.7bar.

Then you calculate for your high temp condition which we assume is 62C condensing and a Pdis of about 17.7bar. In this instance you arrive at 1.75m length.

This is very strange as the tube should be longer with the higher pressure Pdis=17.7 - maybe in the region of 3.7m

Something seems to be seriously wrong here as you cant increase the pressure ratio across the cap and also make it shorter.

It would be a nice idea to braze your new trial cap tube into stub ends of pipe and fit it to the system with flare fittings, certainly make it easier and cleaner to change the cap easily as the R&D progresses.

Also the sight glass is going to be the most valuable piece of kit you have as its the only way to know if you start getting into X territory.

But before you cut the tube why not do some parametric studies, ie get it up to 60C and place some ice on the evap, the suction pressure will fall, the mass flow will fall and the sight glass might start showing more bubbles. The more ice, the more bubbles, plot several evap temps against the bubble count (you will have to invent something here) and do it all again at 50C and 40C.

Do similar sets of measurements without ice and even some heating but working on SC.

Now plot it all out as mass flow versus SC and X for the various delta pressures and you will have the ultimate design graph.

Chef

[desA]

You have back checked the tube length is 2.246m with discharge pressure (Pdis) of about 14.7bar.

Then you calculate for your high temp condition which we assume is 62C condensing and a Pdis of about 17.7bar. In this instance you arrive at 1.75m length.

This is very strange as the tube should be longer with the higher pressure Pdis=17.7 - maybe in the region of 3.7m

From my estimates, the original cap tube was sized at the compressor's rated condition & is, in fact, too long for the in-service application. The cap estimator I use, requests:
Te,sat; SH; Tc,sat; SC; Q'e or m'g, as well as lp piping & evap dP.

Something seems to be seriously wrong here as you cant increase the pressure ratio across the cap and also make it shorter.

The original cap tube length appears to be incorrect. I calculate the cap tube estimate at each operating point across the heating range, then compare with the original tube - in all cases, so far, these estimates are much lower.

It would be a nice idea to braze your new trial cap tube into stub ends of pipe and fit it to the system with flare fittings, certainly make it easier and cleaner to change the cap easily as the R&D progresses.

A good tip. Thanks for this - I will do this.

Also the sight glass is going to be the most valuable piece of kit you have as its the only way to know if you start getting into X territory.

Judging from the current way the system is running, I sincerely doubt that we will get into X territory.

But before you cut the tube why not do some parametric studies, ie get it up to 60C and place some ice on the evap, the suction pressure will fall, the mass flow will fall and the sight glass might start showing more bubbles. The more ice, the more bubbles, plot several evap temps against the bubble count (you will have to invent something here) and do it all again at 50C and 40C.

I've been doing a range of parametric studies, to determine system response to mass charge, to internal heat-transfer augmentation etc. Today, funnily-enough, is the ice-day, but in a different part of the heat-pump anatomy

I'll give the evap-ice a whirl as well - why not? Good idea.

Do similar sets of measurements without ice and even some heating but working on SC.

Now plot it all out as mass flow versus SC and X for the various delta pressures and you will have the ultimate design graph.

Good stuff... I'll press onwards & hopefully have some useful results to contribute to the thread.

Oddly-enough, I've been able to damp out the cap tube noise almost completely. There is no more machine-gun sound.

[Chef]

I make my point again. You checked the compressor design conditions and came up with a tube legnth of 2.2m but then your program calls for 1.7m at a higher pressure - up to 17.7 from 14.7bar. This is not right and something is wrong.

All my results show tubes around 3+ metres for the high temp range to get the SC you want. ACC cubigel and Ashrae all have similar results.

If you put in a 1.2m tube you probably will be in X but you will no doubt find out when you bolt it on. You have already had bubbles in the sightglass so you may already be closer than you think

You say you need these input parameters:-
Te,sat; SH; Tc,sat; SC; Q'e or m'g, as well as lp piping & evap dP.
Not really sure why SH would do anything to affect the cap tube size, its based solely on Tcsat, SC(or X) and Tesat. Only if your running a simulation of the whole system will SH show up but then you will need condenser and evap volumes, surface areas, U, charge, Ta, load, etc.

And for the piping and evap Dp - for very detailed analysis it has some merit but its influence is very small indeed. I think you have bigger problems than +/- 0.5%

Chef

[mad fridgie]

There does seem to be some sort of error,
When did your calcs on the original cap, did you have a very high level of sub-cooling, as your earlier tests indicated, When you have revistited at the higher discharge pressure, did you have only small amount of sub-cooling, this could indicate the need for a shorter cap?

[desA]

Simulation of thermodynamic equilibrium at compressor rated conditions. This balances completely with compressor rated specifications, right down to COP.

I will, below, establish the credibility of the cap-tube estimation tools.

[desA]

Capillary tube estimate at rated conditions. Measured/estimated captube length 2.2462 m (some internal projections not estimated. In other words - bang on).

Please note, maximum allowable superheat in cap-tube estimator is 20K. Actual superheat at compressor suction, is 27.8K.

[desA]

DanCap simulation. Suggested length even shorter.

[Gary]

Any conclusions based upon the original data are invalid, because the strainer is upside down.

Since there is no liquid seal at the cap tube inlet, given a normal charge the superheat would be very high. Extra refrigerant has been added to bring down the superheat, which causes the subcooling to be high, which in turn drives up the Tc,sat, which is masked by the fact that you are measuring the water temp at the top of a static tank with substantial stratification.

On a cap tube system, strainer/drier positioning is not some minor detail to be skimmed over. It changes everything.

[desA]

Thanks, Gary.

I'm not sure why the machine builders decided to install the cap strainer the way they have. It goes contrary to every local refrigerator I've seen. I have a feeling that the unit this one may have been modelled off would have had the exact-same orientiation. It is a famous brand name, of long-standing.

Now, the odd thing with this type of heat-pump is that, due to the poor nature of the refrigerant-to-water heat-transfer, Tc,sat ends up rising very slowly, over the duration of the heating cycle. It almost seems to 'fall into itself' as it were. This is despite very large subcooling values.

It does not respond as you would expect a high-performance condenser to respond. It is very different, I've found.

This 'falling in on itself' is an oddity that, ironically, in the long run, contributes to a very flat heating curve & fairly even COP curve.

[Gary]

Results of yesterday's run.

Tc,exit = exit from tank coil.

If we assume that Tw is in error and that the TD is in fact constant, then the TD remains at 18K and the actual Tw at the end of the run is 44C (18K below Tc,sat), what does that do for your 'falling in on itself' theory?

[desA]

TD for a condenser does not remain constant through any heat-pump cycle. It tends to decrease to a greater, or lesser degree, depending on the type of condenser used.

For these tests, they are quasi-static, in the sense that a small amount of water is passed through the system, in order to slow down the heatup rate - in order to allow data capture (manual), without temps moving off too far.

So, in this sense, the tank is not a static mass of water - it is being adjusted as the incoming water feeds in from the bottom. So, at worst, you could estimate the 'average' water temp in the tank to be (Tw,in + T,top)/2.

So, given that scenario, at the hottest reading Tw,av ~ (28+60)/2 = 44'C... So, yes this is remotely possible.

Practically, though, I would expect the average water temp to be a fair bit higher than this. as the coil does move all the way to the bottom. I'd at a rough guess actually expect something closer to 50'C, or a little higher, judging from the time it takes to get the tank contents to drop off the highs, when new fresh water is pushed through after a run.

Basically, the main reason I see for the 'falling in on itself' observation, is that the condenser effectiveness drops off dramatically over the cycle & the condenser just does not push back as expected.

I'm still trying to figure out how to get a decent represenative average tank temperature. It would take a fair bit of lower bleed-off to get to internal water during a test & this would modify the test results.

[Gary]

The condenser may coil all the way to the bottom of the tank, but the bottom of the coil is filled with liquid because the strainer is upside down and the system is overcharged as a result. The condenser is far less effective than it should be.

[Gary]

For these tests, they are quasi-static, in the sense that a small amount of water is passed through the system, in order to slow down the heatup rate - in order to allow data capture (manual), without temps moving off too far.

So, in this sense, the tank is not a static mass of water - it is being adjusted as the incoming water feeds in from the bottom. So, at worst, you could estimate the 'average' water temp in the tank to be (Tw,in + T,top)/2.

So, given that scenario, at the hottest reading Tw,av ~ (28+60)/2 = 44'C... So, yes this is remotely possible.

Replacing warm water at the top with cold water at the bottom increases the dT between top and bottom, lowering the average temp. Average temp of 44C is not a remote possibility. It is a strong probability.

[desA]

The condenser may coil all the way to the bottom of the tank, but the bottom of the coil is filled with liquid because the strainer is upside down and the system is overcharged as a result.

I'd agree with the system being grossly-overcharged. Can you perhaps explain further why this can be attributed to the strainer orientation?

The condenser is far less effective than it should be.

To be sure - it is absolutely awful.

How would you improve the coil in this case?

[desA]

Replacing warm water at the top with cold water at the bottom increases the dT between top and bottom, lowering the average temp. Average temp of 44C is not a remote possibility. It is a strong probability.

I'm not a betting man...

I'll have to work out a way to measure the internal contents.

[Gary]

I'd agree with the system being grossly-overcharged. Can you perhaps explain further why this can be attributed to the strainer orientation?

The overcharge is there to compensate for the lack of liquid seal at the cap tube inlet. Without the overcharge the evap would be starved and the superheat would be high.

If the strainer were oriented properly, there would be a liquid seal at the cap tube inlet, the evap would be fed properly and the overcharge would not be needed.

[desA]

Originally Posted by desA
I'd agree with the system being grossly-overcharged. Can you perhaps explain further why this can be attributed to the strainer orientation?

desA:
The overcharge is there to compensate for the lack of liquid seal at the cap tube inlet. Without the overcharge the evap would be starved and the superheat would be high.

If the strainer were oriented properly, there would be a liquid seal at the cap tube inlet, the evap would be fed properly and the overcharge would not be needed.

Thanks for the detailed explanation. This makes sense, but I don't think that it the full solution.

Now, what if the system were actually manufactured in imbalance. In other words, if the evap, compressor & condenser are mis-matched? How would this affect the system charge? How would the condenser system shape affect required system charge?

[Gary]

Thanks for the detailed explanation. This makes sense, but I don't think that it the full solution.

Now, what if the system were actually manufactured in imbalance. In other words, if the evap, compressor & condenser are mis-matched? How would this affect the system charge? How would the condenser system shape affect required system charge?

Some matches are better than others, but there are no mis-matches... and the system will seek and find a point of balance, however disadvantageous that balance point may be.

Let's fix the obvious problem and then see what we have.

[desA]

Let's fix the obvious problem and then see what we have.

Are we 100% certain that this 'obvious problem' is the real underlying reason for the system operating the way it currently does?

Are there examples of any other cap-tube systems operating successfully with strainers in the inverted position? What of the systems which do not have strainers (other than blockages)?

I'm asking this in order to make sure that we fully evaluate all possibilities, before moving down one path.

[Chef]

If you have a lot of SC then it wont really matter which way the filter is oriented as the condenser has quite a bit of liquid in it, but once the SC gets small then having the filter run downhill as Gary is suggesting is probably essential. You can just cut it and change it - but once you cut you may as well change all things that need to be done and without the parametrics you cant decide what else needs to be changed.

Chef

[Gary]

Are we 100% certain that this 'obvious problem' is the real underlying reason for the system operating the way it currently does?

Are there examples of any other cap-tube systems operating successfully with strainers in the inverted position? What of the systems which do not have strainers (other than blockages)?

I'm asking this in order to make sure that we fully evaluate all possibilities, before moving down one path.

Is it a problem? Yes.

Is it the only problem? Probably not.

[desA]

If you have a lot of SC then it wont really matter which way the filter is oriented as the condenser has quite a bit of liquid in it, but once the SC gets small then having the filter run downhill as Gary is suggesting is probably essential.

These were my thoughts. Practically, for the current tank/coil setup, with its long vertical dicharge leg, I sincerely doubt that we will see X territory. Intermitant clouds of small bubbles seem to be the norm.

My reasoning is that at lower mass charge, the system response is very obviously under-charge. The compressor also runs far hotter, to the touch, at low charges. As the charge increases, the compressor cools off to a decent running condition.

So, under the current set of system components, a large mass charge is required to make the system settle into a reasonable operating regime.

You can just cut it and change it - but once you cut you may as well change all things that need to be done and without the parametrics you cant decide what else needs to be changed.

Agreed. The parameteric studies are incredibly useful in fully characterising the system response. When finally, system changes are introduced, they will be in a careful manner. We can then learn from the response of each change.

[Gary]

If you have a lot of SC then it wont really matter which way the filter is oriented as the condenser has quite a bit of liquid in it, but once the SC gets small then having the filter run downhill as Gary is suggesting is probably essential. You can just cut it and change it - but once you cut you may as well change all things that need to be done and without the parametrics you cant decide what else needs to be changed.

Chef

Exactly so. The high SC compensates for the incorrect strainer orientation.

Be sure to change the drier orientation, too.

[desA]

Be sure to change the drier orientation, too.

I've been mulling this one over for some time, having installed this orientation in another test heat-pump. I have to say that it gave me some very, very odd results, that were not overly simple to explain away, other than the vertical drier orientation. I am now not too convinced of its effectiveness, I have to say. Sporlan advocate this, but I've not seen it in other literature (probably is there, & I've missed it).

Care has to be taken in some instances, I expect, with the drier position relative to the condenser discharge & type. It can present issues within the condenser itself in terms of liquid-clearing.

For a vertical heat-pump, the clearing issue should not affect the condenser as it already has to push the liquid upwards a fair amount.

I remember Chef's concern regarding the possibility of this filter-drier accumulating too much refrigerant & that this could create a problem under critical charge conditions. Perhaps we could discuss this further?

[mad fridgie]

How do you know that there is sub cooling.
How do you know the pressure drop across the cap.
Presumption is OK for design we need to be taking actual measurements at all point within the systerm.
Also interesting how 2 different software programs give 2 different answers.

[Chef]

The dryer on Danfoss units is top in and bottom out (on the component breakdown I have for this model anyway) so I see no reason to not set it downwards.

The main issue seems to be with the riser on the exit of the condenser and also the temp drop down the condenser will always want to generate quite a lot of SC and hence hold up some liquid. Not sure if trying to get rid of this effect with charge manipulation is a good idea rather than looking at the condenser itself?

Changing the cap can obviously change the SC but if your condenser is not right then may your just shifting the problem elsewhere.

I think some more details on the condenser and the temps you plan to get will help.

Chef

[mad fridgie]

I have had a closer look at the 2 programs, the danfoss has no allowance for sub cooling, so a shorter cap is required, why not re enter detail in first program with no sub cooling

[desA]

The dryer on Danfoss units is top in and bottom out (on the component breakdown I have for this model anyway) so I see no reason to not set it downwards.

If you could perhaps put up a picture of the Danfoss standard arrangement for this dryer - relative to the condenser outlet, I'd be most grateful.

The main issue seems to be with the riser on the exit of the condenser and also the temp drop down the condenser will always want to generate quite a lot of SC and hence hold up some liquid. Not sure if trying to get rid of this effect with charge manipulation is a good idea rather than looking at the condenser itself?

It will not help to dramatically reduce charge on this machine, under its current condenser configuration, in order to try & reduce SC. The geometry & condenser design constraints are the main design bottleneck, in my view.

Changing the cap can obviously change the SC but if your condenser is not right then may your just shifting the problem elsewhere.

Yes, agreed, but...

This style of condenser is widely used by a few 'famous name' brands in the industry. Tuning the cap tube to suit 'what is', would be an early target. Some manfacturers seem to very happy with this type of design & they've been pretty successful with it & I'm sure that they won't welcome change.

I think some more details on the condenser and the temps you plan to get will help.

This will be the next step. I've already begun experimenting on how to improve this.

[desA]

I have had a closer look at the 2 programs, the danfoss has no allowance for sub cooling, so a shorter cap is required, why not re enter detail in first program with no sub cooling

There we are.

[mad fridgie]

There we are.

I am I being a pain in the a***. Duty is differnt

[desA]

Yes, it is - by design.

At the same mass flow, with zero SC, the heat-transfer will automatically balance in this software. It's smart...

[mad fridgie]

so which program can you trust?

[desA]

so which program can you trust?

That is the question. The one that you guys use?

[Chef]

Yes, it is - by design.

At the same mass flow, with zero SC, the heat-transfer will automatically balance in this software. It's smart...

So with 8.3C of SC you get 2.2581m
with 0C of SC you get 2.326m

Just 68mm to reduce all that SC to two phase flow.

That seems 'an interesting result'.

Chef

[Chef]

Here is the picture you asked for about the top to bottom drier.

Parts in a Danfoss.gif

[desA]

Thanks Chef.

[desA]

So with 8.3C of SC you get 2.2581m
with 0C of SC you get 2.326m

Just 68mm to reduce all that SC to two phase flow.

That seems 'an interesting result'.

Chef

Ok, fair-enough then. I'm always open to learning new things - cap tubes are new to me, at least. If you'd like to post the results you have to hand, then we can compare notes. What software are you using?

As these programs always say - this is a recommended starting point. If we find that the program is substantially in error, then I will contact the developer - a colleague & have him investigate.

[Chef]

Ok, then. If you'd like to post the results you have to hand, then we can compare notes.

As these programs always say - this is a recommended starting point. If we find that the program is substantially in error, then I will contact the developer - a colleague & have him investigate.

I do not mind providing you with 'a recommended starting point' for your system.

However I do not believe that I should be involved in helping the developer of the program you are using. If the code is wrong its got nothing to do with me and I have no intention to provide details on how to fix it.

As you know I have developed a cap tube sizing program that includes system simulation. Asking me to present my results to demonstrate a possible flaw in someone else's program so you can forward them on to another company is, quite honestly, disappointing.

This forum is a public domain and they may even be reading this thread - I have no problem with that.

My comments on both the Danfoss program and the one you are using are normally limited to 'interesting results' or 'use with due diligence'. I will try to stay in this area in future.

Chef

My reply was based on your original post before you edited it.

[desA]

Hi Chef. Wow. My thoughts most certainly did not involve industrial espionage of your cap-tube program. Far from it.

I was not aware that you had a cap tube program so dear to your heart - I knew you'd worked with EES & so forth. I'd say that if it will present a problem, then we shut this whole thread, here & now.

You can find the simulation software at www.coolit.com

It's not worth the hassle, honestly. I'm done. I'll program the cap tube simulator myself - it's not that difficult, actually. I have the logic in my library.

[Gary]

As I see it, there is a time for calculation & prediction and there is a time for measurement & analysis.

The time for calculation & prediction is before the system is built

The time for measurement & analysis is after the system is built.

I think we should re-orient the strainer and drier, adjust the charge and then see what the measurements tell us needs to be done next.