Expansion noise - capillary tube into evaporator

[desA]

Hi Des,

I am changing on a weekly base blocked driers on small cup tube systems such as air driers, beverage cabinets, water coolers, domestic fridges and freezers.

This refrigerant is too good when it comes to cleaning.

By using regular 1\4 drier, the problem disappear.

Thanks very much, chemi-cool. Your experience is very valuable.

Do you perhaps have a link, or picture of the 'regular 1/4" drier' you're referring to?

[desA]

has anyone used that 1234 whatever refrigerant it is yet.
maybe you could step us into the future des by taking that path.
i know more r&d = more work

Interesting point. I've been following that refrigerant for some time & would love to test it as a drop-in on some of these test machines. I'm not sure that it is commercially available yet.

[nike123]

Thanks very much, chemi-cool. Your experience is very valuable.

Do you perhaps have a link, or picture of the 'regular 1/4" drier' you're referring to?

I think that he mean something like Danfoss DCL/DML 032S

http://rc.danfoss.com/TechnicalInfo/.../PDE00B322.pdf

[chemi-cool]

Thanks Nik,

Thats the one, DCL 052's

[desA]

I think that he mean something like Danfoss DCL/DML 032S

http://rc.danfoss.com/TechnicalInfo/.../PDE00B322.pdf

Thanks so much, Nike.

[desA]

To extend the discussion a little further.

I calculated what I estimated to be 'critical charge' for this system, (evacuated, N2)x3, charged in steps until this 'critical charge', observing system response along this charge-addition process. I've set up LP & HP gauges, & a temp sensor (still rudimentary at this point).

Well, very interesting so far. General qualitative observations. There is none of the pinging, popping, machine-gun noise - only a rushing sound as would be expected from a high-velocity flow, in a thin tube. The compressor is also running slightly warmer to the open hand (don't laugh), than previously. Te,sat has dropped a few degrees.

I'm going to make a judgement call at this stage, subject to additional feedback from the experimental stage. It looks like the system could very well have been over-charged, as received. Let's see as testing develops.

[chemi-cool]

Des, Can you set a video camera and let us see in real time how it goes?

After all, by now we are part of your experiment...

[desA]

Des, Can you set a video camera and let us see in real time how it goes?

After all, by now we are part of your experiment...

Hahaha... That's an incredibly interesting concept. Let me think a bit more on that. I have colleagues who've done this in their factories & labs. Could be very interesting...

[desA]

I'd like to ask for views on determining the 'critical charge' for a cap-tube system. I've listed my logic below. This is very much open to discussion.

What I have done so far, is as follows (based on all liquid residing in evap in off-duty condition):
1. Evaporator internal tube volume = Ve [L];
2. Suction pipe max height - 90% tubes liquid-full => Vc=0.9*Ve [L]
3. R134a liquid density at 30'C ~ 1187.51 kg/m3 (data program);
4. Critical mass of cold refrigerant in evap - xxx g

I'm now up to this charge limit, having previously charged in steps of 20g, from a lower starting point.

What other items can/should be included in this critical charge calculation? In other words, are there a few more grams I could legitimately squeeze into the system, without endangering the compressor - suction line - portion of, compressor ?sump?.

I have kept an eye on the evap superheat, throughout. It is currently in the region of 7-7.5K.

Obviously, I'll put up test results as the experimental side progresses. I also have a reliable CapTube program on hand.

[desA]

Maybe this article could have some interesting information.

Thanks Nike. My curiosity overcame my miserly tendencies. I downloaded the paper a few minutes ago. I'll study it over the next few days.

There is always something useful in these papers.

[Chef]

What I have done so far, is as follows (based on all liquid residing in evap in off-duty condition):
1. Evaporator internal tube volume = Ve [L];
2. Suction pipe max height - 90% tubes liquid-full => Vc=0.9*Ve [L]
3. R134a liquid density at 30'C ~ 1187.51 kg/m3 (data program);
4. Critical mass of cold refrigerant in evap - xxx g

I have kept an eye on the evap superheat, throughout. It is currently in the region of 7-7.5K.

Obviously, I'll put up test results as the experimental side progresses. I also have a reliable CapTube program on hand.

What you will be interested in is the correct charge for optimum COP and best operating conditions over the full range. If you calc the charge in the OFF condition it may be way off, but it will serve as a maximum value.

When you calc the charge in operation you will need to know in what phase the cap tube is operating in. For instance if it will always have SC then there will be more liquid in the condenser but if it runs mainly with X then there will be far less liquid in the condenser.

Choose your operating cycle on a PH diagram, get the value of X (or SC) for h3 and X for h4.
Lets assume X=0.0 at h3,
then using a simple linear condensation through the condenser Vol liquid = 0.5*Vc
It may be more accurate to use a logarithmic formulae for the condensate and most certainly in the case for SC.

Once X is 0.1 or 0.2 then the volume of liquid falls dramatically and can be found using
vol liquid =(Vc-X*Vgas*Vc)/(1-X) where Vgas is specific gas volume and Vc is condenser volume. Again doing it logarithmically and in finite elements down the condenser would be better. You can see it falls to below 10% of the value at X=0 very quickly.

The evap can similarly be treated and also for the piping but only you can determine this as you have the geometry.

As for the SH - is this SH out of the evap or into the compressor, it is important as cap systems can have SH as low as 1C at design conditions at evap exit.

Any cap dimensions, pressures and temps yet?

Chef

[desA]

Thanks very much, Chef.

What you will be interested in is the correct charge for optimum COP and best operating conditions over the full range. If you calc the charge in the OFF condition it may be way off, but it will serve as a maximum value.

The issue here, is the definition of critical charge, as I see it. If we are saying that the charge may never exceed a certain holding volume in the evaporator, in the cold condition, then this may very well occur long before the system optimum COP, or maximum power (different points), can occur.

In other words, the logic may be something like this:
1. Where system refrigerant volume is small, evap internal volume dominates;
2. Where system refrigerant volume is large, evap internal volume not dominant;

Cold condition (off-cycle):
1. Charge migrates to evap:
1.1 Storage critical charge - liquid flood-back control critical.
2. Charge migrates to evap:
2.1 Evap has ample bufer storage - liquid flood-back not critical.

Operating condition:
1. Cold-condition charge dominates - operating conditons just have to fit in.
2. Operating charge important - operational critical charge, to prevent liquid floodback.

As I see it, there are two very different scenarios here, which seem to be loosely lumped under the concept of 'critical charge'.

[desA]

Further to the determination of operational optimum charge

For a larger system volume, where the evap internal volume exceeds the total system charge, the system optimum for either COP, or max condenser power (two different scenarios), can really only come by a huge amount of testing, where the charge level is gradually increased & cycle performance figures closely recorded.

Typically, this can take some 20-30 rounds of tests, to hone in on the optimum charge range. The final balance of COP, to Qc,max can turn in a fairly short difference in charge.

[desA]

When you calc the charge in operation you will need to know in what phase the cap tube is operating in. For instance if it will always have SC then there will be more liquid in the condenser but if it runs mainly with X then there will be far less liquid in the condenser.

With this kind of coil condenser, the coil winds its way from top, to bottom, with a vertical up-leg for the outlet fluid. Judging from the sight-glass, with its clouds of bubbles, the SC would appear to be in the range 0-4K. Since the temp gradient over such a tank is some 10K different from bottom-to-top, an amount of liquid sub-cooling, could be expected - subject to adequate coil surface area (never a sure thing).

Choose your operating cycle on a PH diagram, get the value of X (or SC) for h3 and X for h4.
Lets assume X=0.0 at h3,
then using a simple linear condensation through the condenser Vol liquid = 0.5*Vc

Ok, that's an elegant way to look at it. Decent rule-of-thumb. Thanks for that.

It may be more accurate to use a logarithmic formulae for the condensate and most certainly in the case for SC.

Can I ask you to perhaps expand on this point a little further?

Once X is 0.1 or 0.2 then the volume of liquid falls dramatically and can be found using
vol liquid =(Vc-X*Vgas*Vc)/(1-X) where Vgas is specific gas volume and Vc is condenser volume.

Ok. Understood. I had to re-read a few times. Condenser.

Again doing it logarithmically and in finite elements down the condenser would be better. You can see it falls to below 10% of the value at X=0 very quickly.

If you explain the logarithmic part? dTlm? Which temperatures? The finite element part I understand.

The evap can similarly be treated and also for the piping but only you can determine this as you have the geometry.

Good point.

As for the SH - is this SH out of the evap or into the compressor, it is important as cap systems can have SH as low as 1C at design conditions at evap exit.

SH at evap discharge & SH' at compressor inlet. Surely with an SH ~ 1K at evap discharge, the evap would be very close to flooding - with some droplet carryover? This could be a little tricky. In my system, the operating SH today, at mid cycle is ~ 6.5K at evap discharge. Observing the evap tube hairpins also gives a reasonable 'feel' for the amount of the core involved in super-heating. Currently this is in the range of around 24-26.5% of core tube count.

Any cap dimensions, pressures and temps yet?

I've got the pressure & temps parts moving & should have a set of results tomorrow. I've been running rough trials to check on the evap behaviour in regards to flooding. I'll have to go off & get a decent micrometer to measure the cap tube diameter.

[desA]

Does all the refrigerant migrate to the evaporator during the off-cycle - with a heated condenser?

How do we know this to be true?

This is the principle I'm trying to establish in term of the definition of the term 'storage critical charge' as distinct from 'operational critical charge'.

[Chef]

DesA

This is a better equation to calc the charge.

Charge (Kg)=Vc*g[1+(1-X)/X] / [1+(g/L)(1-X)/X]
where Vc is vol of condenser m3
g is gas density Kg/m3
L is liquid density Kg/m3

I do it every 1% along the condenser and get X from heat transfer equations for the input conditions using an equation solver. You wont need that complexity so it would be easy to put in excel and set the value of X from .01 to 1 in 100 steps and then sum up the results. Easier this way.

If you want to take into account SC and de-superheating then you could set a few segments at each end to 0 and 1 and see the changes easily.

You can even 'shape' the curve of X along the condenser from linear to exponential or whatever you like by changing X to alter slower the entrance and faster at the exit.

Hope this works better for you.

Chef

[Gary]

Does all the refrigerant migrate to the evaporator during the off-cycle - with a heated condenser?

How do we know this to be true?

This is the principle I'm trying to establish in term of the definition of the term 'storage critical charge' as distinct from 'operational critical charge'.

If we take two half full refrigerant containers and connect them with a hose, then warm one of the containers, the vapor will condense in the cooler container and evaporate in the warmer container.

This will continue until the cooler container is full of liquid and/or the warmer container has only vapor.

At this point the pressures will equalize and will correspond to the temperature of the cooler container.

[desA]

If we take two half full refrigerant containers and connect them with a hose, then warm one of the containers, the vapor will condense in the cooler container and evaporate in the warmer container.

This will continue until the cooler container is full of liquid and/or the warmer container has only vapor.

At this point the pressures will equalize and will correspond to the temperature of the cooler container.

An excellent analogy.

Now lets bring the interconnecting pipework into the picture. Logically, any piping exposed to ambient air should be at similar temperature to the idle evaporator. It is logical then, that condensation must also take place in the liquid line, filter-drier (150~ 200g), suction line, compressor sump.

In other words, the 'storage critical charge' may need to logically be extended beyond that of the evaporator itself.

[desA]

DesA

This is a better equation to calc the charge.

Charge (Kg)=Vc*g[1+(1-X)/X] / [1+(g/L)(1-X)/X]
where Vc is vol of condenser m3
g is gas density Kg/m3
L is liquid density Kg/m3

I do it every 1% along the condenser and get X from heat transfer equations for the input conditions using an equation solver. You wont need that complexity so it would be easy to put in excel and set the value of X from .01 to 1 in 100 steps and then sum up the results. Easier this way.

If you want to take into account SC and de-superheating then you could set a few segments at each end to 0 and 1 and see the changes easily.

You can even 'shape' the curve of X along the condenser from linear to exponential or whatever you like by changing X to alter slower the entrance and faster at the exit.

Hope this works better for you.

Chef

Thanks very much, Chef. Much obliged.

I'll program that up & we can use it for discussion purposes.

[Chef]

SH at evap discharge & SH' at compressor inlet. Surely with an SH ~ 1K at evap discharge, the evap would be very close to flooding - with some droplet carryover? This could be a little tricky. In my system, the operating SH today, at mid cycle is ~ 6.5K at evap discharge. Observing the evap tube hairpins also gives a reasonable 'feel' for the amount of the core involved in super-heating. Currently this is in the range of around 24-26.5% of core tube count.

Usually a cap system has a suction accumulator and depending on its size you can push the SH down to gain extra evap area for cooling. This is especially true if the tube is sized to run in X at the most common operating point as very little extra liquid can be pushed into the evap as the condenser is already starting to be very low. But if you have SC and a SH of 1C then any small change in condenser conditions could push more liquid into the evap and possibly cause floodback. In the case of running in SC then some SH would be safer.

Maximum COP is obtained with an SC of about 1C or 2C and SH of just a few C.

Chef

[mad fridgie]

Usually a cap system has a suction accumulator and depending on its size you can push the SH down to gain extra evap area for cooling. This is especially true if the tube is sized to run in X at the most common operating point as very little extra liquid can be pushed into the evap as the condenser is already starting to be very low. But if you have SC and a SH of 1C then any small change in condenser conditions could push more liquid into the evap and possibly cause floodback. In the case of running in SC then some SH would be safer.

Maximum COP is obtained with an SC of about 1C or 2C and SH of just a few C.

Chef

I am not quite sure if the bottom statement is quite true, (refrigerant dependent) if a liquid/suction heat exchanger is included there will be an increase in net cooling effect. thus increasing the COP. But considering this a heat pump, cooling is not required. "Shut up mad!"
The goal is to keep the suction pressure high, i am not sure that you can achieve the desired results using a cap, with out haveing other methods of controls, which would likely cost more than a TXV

[mad fridgie]

If you are trying to compare against the cheaper end air cond/heat pumps that use caps, be aware that they are designed to meet a very specific standard. Acheive an end results becomes easy.
The reality is with these units, as soon as you deviate from design performance and efficiency dive out of the window, compared to systems that have a modulating expansion device.
With ASHP, there does not seem to be such a standard to work to.

[Chef]

I am not quite sure if the bottom statement is quite true, (refrigerant dependent) if a liquid/suction heat exchanger is included there will be an increase in net cooling effect. thus increasing the COP. But considering this a heat pump, cooling is not required. "Shut up mad!"
The goal is to keep the suction pressure high, i am not sure that you can achieve the desired results using a cap, with out haveing other methods of controls, which would likely cost more than a TXV

Hi Mad

Its just that when you plot it out on a PH diagram it is those parameters which give the best COP - it might not be the best point to design for.

Its desA who is doing the design we are just answering his questions and he may be getting paid for it so we can all expect a large (post dated) cheque in the post soon. ??

Not sure why you think a cap is not good for heat pumps? The range of temperature change on the evap is small and also quite small on the condenser and its always within very clear bounds so a cap system duly designed would operate over the range without much change so would be mostly running at its optimum.
In a fridge system the cap sees very large changes in its operating mode so is more difficult to select the correct tube size and its much more of a compromise.

Chef

[desA]

Its desA who is doing the design we are just answering his questions and he may be getting paid for it so we can all expect a large (post dated) cheque in the post soon. ??

I'll answer the remaining posts above, but this did catch my eye & I thought I'd address it head on.

I will state, for the record, that I am not being paid one bean for this research - it is all purely at my own expense. This is an interest & the reason I've put it up on RE so openly, is that we can all learn from the debate.

There are parts of the heat-pump that will not be shown, purely, because that would not be professional practice, nor fair to the original machine builder.

With this kind of thread being put out in the public domain, it should hopefully teach us all something about how these systems really work & allow us to debate in a healthy way. It is the way of the Open Age.

[desA]

Usually a cap system has a suction accumulator and depending on its size you can push the SH down to gain extra evap area for cooling. This is especially true if the tube is sized to run in X at the most common operating point as very little extra liquid can be pushed into the evap as the condenser is already starting to be very low. But if you have SC and a SH of 1C then any small change in condenser conditions could push more liquid into the evap and possibly cause floodback. In the case of running in SC then some SH would be safer.

Surely, though, most designers would want to see some level of SC in the condenser, to create a liquid seal & ensure full condensation takes place (asuming condenser to be of adequate size in the first place)?

Maximum COP is obtained with an SC of about 1C or 2C and SH of just a few C.

Could I ask you to please refer this to a log(p)-h or T-s chart, so that we can debate why this should be so?

[mad fridgie]

Hi Mad

Its just that when you plot it out on a PH diagram it is those parameters which give the best COP - it might not be the best point to design for.

Its desA who is doing the design we are just answering his questions and he may be getting paid for it so we can all expect a large (post dated) cheque in the post soon. ??

Not sure why you think a cap is not good for heat pumps? The range of temperature change on the evap is small and also quite small on the condenser and its always within very clear bounds so a cap system duly designed would operate over the range without much change so would be mostly running at its optimum.
In a fridge system the cap sees very large changes in its operating mode so is more difficult to select the correct tube size and its much more of a compromise.

Chef

On a Ph diagram yes it presumes a constant on the compressor, so the angle would change slightly (related to comp performance) some where in RE this was described in detail.
With a ASHP the process variables are ambient -10 to 45C and water temps from 10 to 70C.
(unlikely to have ambient at 45C and water at 10) So if we can agree that cap is a fixed pressure drop, then if the water is cold (non controled head pressure) then suction will plumet giving poor COP. these changes are very frequent and are over long periods.
With a fridge you are generally dealing with air load, which is a short term load, normal on and offs are with a small working band, yes ambient change, but ambient is low then infiltration load is low, so long term energy use becomes less important.

[mad fridgie]

I do noy not like caps! Or could it be that I do not know how to design a system truely correct with a Cap? So do not get decent performance!! As TXV does give alittle flexability.
Could be!!!

[desA]

With a ASHP the process variables are ambient -10 to 45C and water temps from 10 to 70C.
(unlikely to have ambient at 45C and water at 10) So if we can agree that cap is a fixed pressure drop, then if the water is cold (non controled head pressure) then suction will plumet giving poor COP. these changes are very frequent and are over long periods.

This is the inherant difficulty in designing an AWHP. Both ends, source & sink, are in a state of flux. It makes designing over an operating range extremely challenging.

[Chef]

So if we can agree that cap is a fixed pressure drop, then if the water is cold (non controled head pressure) then suction will plumet giving poor COP. these changes are very frequent and are over long periods.
.

Thats the point about a cat tube in a heat pump - it is not a fixed pressure drop - quite the opposite.

When the water is 10C there will be a lot of SC available and so even if the pressure is less the flow through the cap tube will increase over the value assumed if it was just related to Dx and Sx. As the temperature of the water rises the SC is less but the pressure is higher so the flow is roughly the same and when its hot then lets assume we get to no SC and much higher pressure which gives a similar flow. So in this arrangement a cap might be good in a heat pump.

Take the point about ambient rangeing from -10 to 45C would cause problems in the evap area but are these temps really common?

Chef

[desA]

Not sure why you think a cap is not good for heat pumps? The range of temperature change on the evap is small and also quite small on the condenser and its always within very clear bounds so a cap system duly designed would operate over the range without much change so would be mostly running at its optimum.

This is precisely why I started this thread - so that we could all have a bash at knocking out much of the speculation & possible mis-conceptions we have come across.

I've personally been scared stiff of them, based on what I'd heard. Seeing a running system in action, with it stroking repeatably across the water temp range 28-60'C, is quite intriguing.

Let's learn what we can about cap tubes & heat-pumps.

[desA]

Thats the point about a cat tube in a heat pump - it is not a fixed pressure drop - quite the opposite.

When the water is 10C there will be a lot of SC available and so even if the pressure is less the flow through the cap tube will increase over the value assumed if it was just related to Dx and Sx. As the temperature of the water rises the SC is less but the pressure is higher so the flow is roughly the same and when its hot then lets assume we get to no SC and much higher pressure which gives a similar flow. So in this arrangement a cap might be good in a heat pump.

Excellent points.

Take the point about ambient rangeing from -10 to 45C would cause problems in the evap area but are these temps really common?

Seasonal variation - yes - daily variation - unlikely, except in a desert. Add into the mix the RH%...

[desA]

Current cap tube dimensions:

OD = 2.4 -2.45mm
L = 2246 mm = 2.246 m

These are from measurements of the existing system.

[Chef]

Surely, though, most designers would want to see some level of SC in the condenser, to create a liquid seal & ensure full condensation takes place (asuming condenser to be of adequate size in the first place)?

Could I ask you to please refer this to a log(p)-h or T-s chart, so that we can debate why this should be so?

Thats very true for TXV and would be ideal for a cap also but as they only have a single point at which they perform 'perfect' it means any system load or temperature changes means the SC will increase and so COP will fall or it will move into X and the COP will again fall. You can't design a cap tube to have the same SC over its operating range - its just the nature of the tube.

But you can size the condenser and tube to give the results as close to the ones you are looking for.

It is SC and X that control the device and so if you can control these 2 or at least understand how they work it will be a lot easier to design a system.

As for the PH diagram! I may be able to put it in words till I get to plotting it.
The system we will assume is simply compressor, condenser, cap tube and evaporator.
Higher SC values mean higher discharge pressure which is bad for COP.
Minimal SH means highest evap pressure which is good for COP.

Oh and about the cheque - just kidding as thats why I said post dated - like jan 2030 or similar! Maybe the humour got lost in translation.

Chef

[mad fridgie]

Thats the point about a cat tube in a heat pump - it is not a fixed pressure drop - quite the opposite.

When the water is 10C there will be a lot of SC available and so even if the pressure is less the flow through the cap tube will increase over the value assumed if it was just related to Dx and Sx. As the temperature of the water rises the SC is less but the pressure is higher so the flow is roughly the same and when its hot then lets assume we get to no SC and much higher pressure which gives a similar flow. So in this arrangement a cap might be good in a heat pump.

Take the point about ambient rangeing from -10 to 45C would cause problems in the evap area but are these temps really common?

Chef

Would you actually have a lot of sub-cooling, or just a low condensing pressure. To get the high level of liquid sub-cooling then the liquid must be backed up in the condenser, which must leave the evap short of refrigerant, I agree that with increased sub-cooling, the pressure drop will reduce, so would increase flow, but would you not then loose the liquid which is backed up. thus loosing the sub-cooling benefit.
Some form of equalbrium will be reached, I still believe that the suction will trend down proportianally to liquid pressure on a critically charged system on a cap. (greater than that of a modulating valve, and non critcal charge)

[Chef]

Some form of equalbrium will be reached, I still believe that the suction will trend down proportianally to liquid pressure on a critically charged system on a cap. (greater than that of a modulating valve, and non critcal charge)

Mad - you may not like cap tubes but you understand them very well.

Yes the evap will get shorter on liquid and be lower pressure and the system will fall off the COP obtainable from a TXV.

But h3 moves to the left on the PH diagram and so does h4 so the amount of heat pumped is still good and maybe even better than at design point.

Its not the best solution but it is the one desA is dealing with.

I dont particularly like caps either because of this SC and X problem but just got caught up in them.

Chef

[mad fridgie]

Taking in all the information, and allowing for the variables, if a cap is to be used (what ever sized is picked)
I believe the most important issue is the size of the suction accumulator. If a relatively large one is installed, we could effectively over charge the system, allowing for increased sub-cooling at lower pressure ratios, and low SH at higher compression ratios.
At the lower ratios SH has less importance in relation to compressor discharge.

[mad fridgie]

Mad - you may not like cap tubes but you understand them very well.

Yes the evap will get shorter on liquid and be lower pressure and the system will fall off the COP obtainable from a TXV.

But h3 moves to the left on the PH diagram and so does h4 so the amount of heat pumped is still good and maybe even better than at design point.

Its not the best solution but it is the one desA is dealing with.

I dont particularly like caps either because of this SC and X problem but just got caught up in them.

Chef

You are have struck on a really good point, one I have pondering on for a while (not real work) and a light has just switched on.
Why do we bother controlling heat pressure (cooling applications) if a piece of refrigeration is just allowed to reach equalbrium would it be cheaper to run.
(only suitable for refrigeration not related to producrs that are moisture sensitive)
Even with TXV a certain pd is required, why do we hold the SCT high, just to keep the SST high.
Umm I might just have a little play. Thanks chef!

[desA]

As for the PH diagram! I may be able to put it in words till I get to plotting it.
The system we will assume is simply compressor, condenser, cap tube and evaporator.
Higher SC values mean higher discharge pressure which is bad for COP.
Minimal SH means highest evap pressure which is good for COP.

Now, I think what is important to note here, is that the optimum COP & maximum Q'cond do not generally coincide.

For the evaporator, the SH to Te,sat relationship is levered off T,cross & Ta,out. SH will naturally decrease across a heating cycle.

For the condenser, sub-cooling is added to the condensing heat-transfer. So, a small increase in Tc,at with increased sub-cooling may actually be beneficial, up to the point at which SC begins to swamp the condensing area. After this point, the trade would be non-beneficial.

Oh and about the cheque - just kidding as thats why I said post dated - like jan 2030 or similar! Maybe the humour got lost in translation.

No problem - thought it good to set the record straight, though.

[desA]

Results of yesterday's run.

Tc,exit = exit from tank coil.

[mad fridgie]

They are really stable, I presume that when you say Te sup, that is actual temp and not superheat over Te sat.

[mad fridgie]

You are using a tank with a coil round it, so you are getting lots of sub-cooiling or coil pressure drop.
can you add service port to the cond outlet pipe.
If this is pressure drop, (not sub-cooling) then stability would be more expected.

[Gary]

Presumably the liquid is being subcooled by the bottom turns of the coil. In order for Tc,exit to be cooled to 40C, the water at the bottom of the tank must be less than 40C. Seems there is considerable stratification inside the tank.

[desA]

They are really stable, I presume that when you say Te sup, that is actual temp and not superheat over Te sat.

Te,sup = measured directly at evap exit.
SH1 = Te,sup - Te,sat

Tc,suc = suction line measured 150mm before compressor - on pipe OD.

[desA]

You are using a tank with a coil round it, so you are getting lots of sub-cooiling or coil pressure drop.
can you add service port to the cond outlet pipe.
If this is pressure drop, (not sub-cooling) then stability would be more expected.

The coil exit temp is fairly consistent across the whole heating cycle, moves ~2K from start to finish. Measured in temp pocket on outside of pipe.

[desA]

Presumably the liquid is being subcooled by the bottom turns of the coil. In order for Tc,exit to be cooled to 40C, the water at the bottom of the tank must be less than 40C. Seems there is considerable stratification inside the tank.

Typically, these coils wind their way down from the top of the tank. The exit line then runs vertically upwards from the base of the tank, in the insulation barrier. There would be some temp loss to atmosphere, I'd expect.

Measurements on these tanks typically shows at least 10K temp stratification from top to bottom.

[Gary]

Typically, these coils wind their way down from the top of the tank. The exit line then runs vertically upwards from the base of the tank, in the insulation barrier. There would be some temp loss to atmosphere, I'd expect.

Measurements on these tanks typically shows at least 10K temp stratification from top to bottom.

I was thinking the vertical run crosses over the warmer tubes, and if anything would gain temp.

Possibly you could measure the temp of the bottom of the tank? Maybe open the drain and measure the temp of the water coming out?

[Gary]

The coil exit temp is fairly consistent across the whole heating cycle, moves ~2K from start to finish. Measured in temp pocket on outside of pipe.

Note also that it coincides with the Te,sat drift.

[desA]

I was thinking the vertical run crosses over the warmer tubes, and if anything would gain temp.

Knowing the way these tanks are made, the coil may, or may not be touching along its length. I wouldn't rely on it.

Possibly you could measure the temp of the bottom of the tank? Maybe open the drain and measure the temp of the water coming out?

Ok, good idea. I'll modify the current piping arrangment, so that I can bleed off at the base during the test. It'll probably take a few days, as I'll be out of the lab during te last few days of this week.

[desA]

Updated cap tube dimensions:

ID = 1.397 mm
L = 2246 mm = 2.246 m

These are from measurements of the existing system.

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.

[mad fridgie]

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