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desA
25-09-2009, 04:34 AM
I would like to open a new thread exploring stability problems I've observed in air-driven evaporators.

In this thread, I would like to explore the following:
1. Industry norms for superheat (SH) against TD.
2. Industry norms for approach against TD.
3. Typical temp_cross = Ta,out-Te,sat.
4. Evap control methodologies eg. fan, superheat, throttle etc.
5. Typical (UA) values.
6. Fin design for moisture control.
7. TXV, or other, inlet control - valve hunting, or not.
8. Fixed-speed fan, or variable?
9. Why does a fixed-speed fan sometimes change speed, & at the same time an Te,sup (evap exit temp) short excursion is noted.
10. Industry norms for dTlm (log mean temp difference).
11. Norms for percentage of evap tube length allowed for superheating.

More can be added as the thread develops.

I'd like this thread to explore & press deeply into evaporator design & operational aspects. The more user experiences we can have, the better, during this exploration experience.

Thanks everyone.

:)

desA
25-09-2009, 07:26 AM
To start the ball rolling, I'll post a link from the Kuba site, regarding suggested SH set-up requirements.

http://www.kueba.de/en-us/Tools/K%C3%BCba-Expansion-Valve-Calculator/Pages/default.aspx

This correlates nicely with the, by now, well known Magoo Rule for setting evaporator superheat:

SH=(0.6 - 0.7)*TD

Will this rule work under all circumstances?
Is it possible to achieve under all circumstances?
What are the pro's & con's of using this SH setting?

Chef
26-09-2009, 03:39 AM
What is the evap design, how many passes, size of tube (diameter).
What is the mass flow rate at design and what is the variation of mass flow rate in a normal cycle.
What kind of fins does it have and is the tube rifled.
Is it a top inlet or bottom inlet.

Is the TXV externally equalised?

What are the instabilities you see - do you have any results to post.

Also a picture of the evap would be nice

Chef

desA
26-09-2009, 05:15 AM
Hi Chef,

Many thanks for entering the debate.


What is the evap design, how many passes, size of tube (diameter).
What is the mass flow rate at design and what is the variation of mass flow rate in a normal cycle.
What kind of fins does it have and is the tube rifled.
Is it a top inlet or bottom inlet.

See attached sketches (all rights of original fabricator acknowledged - no names mentioned):
http://i34.tinypic.com/2hf3oli.jpg
http://i33.tinypic.com/hrj91e.jpg



Is the TXV externally equalised?

Externally equalised TXV.


What are the instabilities you see - do you have any results to post.

In operation, fan runs at v,face=3.6m/s (inlet). Evap, according to calculations, over-boils, relative to compressor requirment. Can hear fan-speed corrections at times, during which the coil superheat adjusts slightly (no hunting), to a slightly different level. Machine then runs smoothly, until next 'adjustment event. Te,sat drift observed from typically 14'C (start-of-cycle) to 19'C (end of cycle).

Chef
26-09-2009, 01:06 PM
OK from what the pictures show is a 5 way split and each split is about 4.1m long. Assuming the inlet and outlet conditions from previous posts it is possible to see what is happenning but the mass flow rate will be needed.

It would probably be better to provide the compressor volume flow rate so mass flow rate can be adjusted to exit conditions of the evap as it changes through the cycle. Of course that would mean the SH and the suction pressure would need to be tabulated.

The reason to have the volume flow rate is because it dictates the flow regime inside the tubes of the evaporator. If the flowrate alters the regime might change - ie from slug to plug to annular etc. These changes in flow regime effect the heat transfer coefficients and hence where the refrigerant 'boils' off. If the regime is on or near a transition point then it may pop from one to the other and back again, this effects the superheat and the valve correspondingly responds. Now the flow rate has altered and the flow regime pops back to its original state. This may manifest itself as a cyclic response.

However you say it is stable with respect to hunting but this flow regime change (if it does indeed occur in your system) might not cause hunting. If the time constant of the valve is of an order of 3 or 4 times differant from the flow regime changes then it will appear stable from a hunting point of view. It may show small changes in SX but not the big variations normally associated with classic hunting.

So the next question is what is the time constant of your TXV, you can easily measure this with a small purtabation to the system (usually by suction throttling the compressor inlet by a few PSI) and timing how long it takes to re-establish equilibrium when the system is running in a constant stable condition. Maybe you could measure that paramter?

So armed with flowrate, SH, SX, and a time constant it should be possible to see if your evap is in or out of a fragile zone.

Mass flow equalisation among the 5 splits is another thing all together.

Chef

desA
26-09-2009, 02:02 PM
Thanks for an excellent review.

^ SX = ? :confused:

I can work on the time-constant, by perturbating the fan speed, via the dimmer switch. Can drop that down a fair wack, then bring it back up again.

Would that be enough?

desA
26-09-2009, 02:07 PM
A further thing to add is as follows:

All tests are at a quasi steady-state, whereby the Tc,sat is brought up close to the test point, then taken slowly into the test point & held there during the duration of the data recording.

With a full transient run, the density waves & system re-balancing act can be heard from time-to-time, via the fan speed adjustment - & watching the SH readout. The SH movement & re-balance takes in the order of 20-30sec.

desA
27-09-2009, 03:10 AM
OK from what the pictures show is a 5 way split and each split is about 4.1m long. Assuming the inlet and outlet conditions from previous posts it is possible to see what is happenning but the mass flow rate will be needed.

It would probably be better to provide the compressor volume flow rate so mass flow rate can be adjusted to exit conditions of the evap as it changes through the cycle. Of course that would mean the SH and the suction pressure would need to be tabulated.

Question:

Mass flow of what?

Under certain operating modes, the measured heat-transfer across the evaporator shows more mass evolution than the compressor swept volume rate requires. It is as though the evaporator 'over-boils', relative to the compressor requirements.

How does SH & its relationship to the TXV mass-flow control have any bearing on what the compressor requires.

In other words:
1. Does the evaporator over-boil & drive the compressor; or
2. Does the evaporator supply on demand, what the compressor requires.

Who is the boss?

mad fridgie
27-09-2009, 03:11 AM
What is the limit of your process variables, and what is your required result. (stability, maximum energy transfer, piping velocity etc)
As far as i can see by controlling the fan, your are attemting to maintain load (Q), are you also maintaining liquid pressure and temperature, If not then properties of the evap are going to change. Are you critical charge or using a liquid reciever.

Chef
27-09-2009, 03:57 AM
Thanks for an excellent review.

^ SX = ? :confused:

I can work on the time-constant, by perturbating the fan speed, via the dimmer switch. Can drop that down a fair wack, then bring it back up again.

Would that be enough?

SX is suction pressure to the compressor like DX is its discharge - just depends on where you learnt the basics from I suppose to what one calls things.

The time constant really needs needs to be measured by purturbations to the exit pressure. It is the response of the feedback loop your trying to measure when you change the SH in a step fashion, but putting a wet cloth on the bulb will not work. You need both the SH and the pressure to change so it is usually done by closing a suction valve a little.

Chef

Chef
27-09-2009, 04:09 AM
With a full transient run, the density waves & system re-balancing act can be heard from time-to-time, via the fan speed adjustment - & watching the SH readout. The SH movement & re-balance takes in the order of 20-30sec.

Density waves??? It sounds here your referring to a change in the TXV flow which then alters the pressure at the inlet to the evap. The pressure change then travels along the evaporator till it is sensed by the bulb and begins to alter the system balance. Is this density waves?

This is not thesame time constant one gets if the SX is step changed.

Chef

Chef
27-09-2009, 04:43 AM
Question:

Mass flow of what?

Under certain operating modes, the measured heat-transfer across the evaporator shows more mass evolution than the compressor swept volume rate requires. It is as though the evaporator 'over-boils', relative to the compressor requirements.

How does SH & its relationship to the TXV mass-flow control have any bearing on what the compressor requires.

In other words:
1. Does the evaporator over-boil & drive the compressor; or
2. Does the evaporator supply on demand, what the compressor requires.

Who is the boss?

The mass flow of the compressor, but better defined by volumetric flow rate, SX and SH as it allows variations to be accounted for.

The only constant in the system process is the compressor volumetric flowrate, all the other parameters vary till a system wide balance is acheived. So this really means that condensor and evaporator will vary thier respective refrigerant charge as loads vary to get to this balance for prevailing temperatures at the evap and condensor.

The mass flow rate is very important as it is reqired to determine pressure drops which is proportional to m^2/d^5 (m=mass flow and d=diameter of pipe) and is also used to calculate the heat flow. Both of these are needed to estimate the evaps performance and what is happening inside at any point in time.

As mad fridgie points out the conditions in the evap will change even with fan control.

When you say 'over boils' - not sure what this is as an evaporator cant really decide to do something outside of the parameters that control it. And using a mass flow derived from calculations to compare to real measured values is dangerous at best!

If you do a refrigerantt charge balance between the evap and condensor it may help you see where your drift comes from.

Chef

mad fridgie
27-09-2009, 07:12 AM
I suspect your waves are more to do with changing conditions in your liquid feed to the TXV, thus Liquid to vapor ratio, is changing slightly by mass, but increasing largely by volumn. Causing a wide range of effects down stream.
Pressure drop, heat transfer co-efficients.

desA
27-09-2009, 09:45 AM
What is the limit of your process variables, and what is your required result. (stability, maximum energy transfer, piping velocity etc)

1. Maximum COP;
2. Best posible heat-transfer to suit (1);
3. Stability under very high TD conditions & excessive RH%.

Point (3) is NOT a simple given - there is a hidden 'devil' lurking in this.



As far as i can see by controlling the fan, your are attemting to maintain load (Q), are you also maintaining liquid pressure and temperature, If not then properties of the evap are going to change.

Fan control is an option, under a certain evap 'mode' of operation. It may not be under an alternative 'mode' of operation.



Are you critical charge or using a liquid reciever.

Critical charge - with some slack taken up by Gary's excellent trick of vertical filter-drier, acting as a 250g receiver...

desA
27-09-2009, 10:05 AM
SX is suction pressure to the compressor like DX is its discharge - just depends on where you learnt the basics from I suppose to what one calls things.

I'm a graduate Mechanical Engineer, off a Physics platform (PhD ABD - momentum waves). Some of the jargon is new to me & I'll ask you to pardon my ignorance... :o



The time constant really needs needs to be measured by purturbations to the exit pressure. It is the response of the feedback loop your trying to measure when you change the SH in a step fashion, but putting a wet cloth on the bulb will not work. You need both the SH and the pressure to change so it is usually done by closing a suction valve a little.


I have access to a service valve, in the suction line - so I can tweek this. How much of a perturbation do you suggest? Screw in x turns, then screw out x turns - measure?

desA
27-09-2009, 10:11 AM
Originally Posted by desA
With a full transient run, the density waves & system re-balancing act can be heard from time-to-time, via the fan speed adjustment - & watching the SH readout. The SH movement & re-balance takes in the order of 20-30sec.

Chef:
Density waves??? It sounds here your referring to a change in the TXV flow which then alters the pressure at the inlet to the evap. The pressure change then travels along the evaporator till it is sensed by the bulb and begins to alter the system balance. Is this density waves?

I have interpreted this as a periodic re-balancing of liquid from the condenser (HP side), to the LP side. The condenser tends to build up a liquid reserve over the course of a heating cycle. It will have to gradually re-balance during the next cycle.

By-the-way, the first heat-up cycle, has different characteristics to following cycles - over shorter hysteresis band, which are repeatable, but significantly different from the initial heat-up cycle.



This is not thesame time constant one gets if the SX is step changed.

Ok, fair-enough. We'll see how this evap & system respond to a sv perturbation. :)

Thanks for the advice...

desA
27-09-2009, 10:23 AM
The mass flow of the compressor, but better defined by volumetric flow rate, SX and SH as it allows variations to be accounted for.

The only constant in the system process is the compressor volumetric flowrate, all the other parameters vary till a system wide balance is acheived. So this really means that condensor and evaporator will vary thier respective refrigerant charge as loads vary to get to this balance for prevailing temperatures at the evap and condensor.

The mass flow rate is very important as it is reqired to determine pressure drops which is proportional to m^2/d^5 (m=mass flow and d=diameter of pipe) and is also used to calculate the heat flow. Both of these are needed to estimate the evaps performance and what is happening inside at any point in time.

As mad fridgie points out the conditions in the evap will change even with fan control.




When you say 'over boils' - not sure what this is as an evaporator cant really decide to do something outside of the parameters that control it. And using a mass flow derived from calculations to compare to real measured values is dangerous at best!

Thus far, interestingly-enough, my measure/derived - evap heat-loads - to refrigerant mass flowrate, are within a small percentage, when the system is running properly.

It is possible for an evap to be able to boil off more vapour than a compressor can draw off at any instant. If the evap is not tuned properly to the compressor, then some very, very odd things occur. Most likely this vapour re-condenses before the compressor can extract it.

There exists, & I've measured it, a critical point at which the evap goes into a 'stall mode', where it goes daft.



If you do a refrigerantt charge balance between the evap and condensor it may help you see where your drift comes from.


The temperature drift is actually now pretty simple to understand. The evap is out of balance relative to the compressor requirements. The evap then seeks to adjust itself until its output matches the draw-off requirement of the compressor - when that system balance point is achieved, then the Te,sat settles there.

Something has to be in charge in the system & when the evap is the boss, then the system drifts - when the compressor is the boss, & evap subservient, then the system remains absolutely & utterly stable.

desA
27-09-2009, 10:25 AM
For the non-linear purists in our community:

How many unstable & stable bifurcation points would you estimate that an evaporator has?

Hence, how many possible 'modes' of operation are possible from an evaporator?

:) :)

mad fridgie
28-09-2009, 02:02 AM
With a face velocity of 3.6M/s you also need to take into consideration the effect of free water travelling through the coil. (not down the coil) Practically this will be rather "lumpy"
Can not think of a better word?

desA
28-09-2009, 02:31 AM
With a face velocity of 3.6M/s you also need to take into consideration the effect of free water travelling through the coil. (not down the coil) Practically this will be rather "lumpy"
Can not think of a better word?

I agree completely. That is actually a very good observation, to be honest.

Intermittent, sporadic, lumpy.

In my view, the fan selection for this evaporator is completely incorrect, considering the original design called for 2.03m/s.

What air face velocity range would you consider to be appropriate, under extremely high humidity conditions?

Chef
28-09-2009, 02:46 AM
I have access to a service valve, in the suction line - so I can tweek this. How much of a perturbation do you suggest? Screw in x turns, then screw out x turns - measure?

Try 1PSI or 2PSI change as a step function, ie just leave it at the new setting.

You should either see a nice smooth curve back to balance which shows the feedback is over the critical damping and is probably what you want.

If the response does 2 or 3 small and ever reducing oscilations it means its still stable and OK, if you get 5 or 6 or more reducing oscillations then its a bit close to the limit.

Chef

desA
28-09-2009, 02:51 AM
So armed with flowrate, SH, SX, and a time constant it should be possible to see if your evap is in or out of a fragile zone.

Can you expand more on the concept of 'fragile zone'?



Mass flow equalisation among the 5 splits is another thing all together.


Could you expand a little more on this, please?

mad fridgie
28-09-2009, 02:53 AM
To stop water carry over, we need to keep face velocity under 2.4M/s, so your original selection was correct.
You may infact find, that with falling water that your heat transfer co-efficient increases, compared to that of dry air. (not to be mistaken for phase change of the water)
Of course with reduced velocity comes reduce air mass flow, you will have to re test

desA
28-09-2009, 02:55 AM
Try 1PSI or 2PSI change as a step function, ie just leave it at the new setting.

You should either see a nice smooth curve back to balance which shows the feedback is over the critical damping and is probably what you want.

If the response does 2 or 3 small and ever reducing oscilations it means its still stable and OK, if you get 5 or 6 or more reducing oscillations then its a bit close to the limit.

Chef

Ok, will do. Many thanks. I just need to modify one of covers to give me easy access to the SV.

So, to confirm, I'll be observing the SX pressure gauge response, not the evap pressure gauge? (I have two set up).

desA
28-09-2009, 03:00 AM
To stop water carry over, we need to keep face velocity under 2.4M/s, so your original selection was correct.

Thanks for that. The evap we're dissecting is not one of my designs. I typically go slightly lower, still.



You may infact find, that with falling water that your heat transfer co-efficient increases, compared to that of dry air. (not to be mistaken for phase change of the water)

Agreed. Inter-fin velocity increase, & moist air higher Cp.



Of course with reduced velocity comes reduce air mass flow, you will have to re test

True.

mad fridgie
28-09-2009, 03:01 AM
5 leg distribution!
Even though each circuit is equal, the postion within the coil block effects the individual mass flow, as much to do with the inbalance of air flow across the coil, this becomes more noticible on close coupled units (fan and evap) So when testing actual velocity, you measure using agrid pattern to give an averaged face velocity

Chef
28-09-2009, 03:06 AM
.
It is possible for an evap to be able to boil off more vapour than a compressor can draw off at any instant. If the evap is not tuned properly to the compressor, then some very, very odd things occur. Most likely this vapour re-condenses before the compressor can extract it.

There exists, & I've measured it, a critical point at which the evap goes into a 'stall mode', where it goes daft..

If the evap boils of more than the compressor can 'suck' then the pressure would rise. The valve responds instantly to pressure changes and so would close until it establishes a new equilibrium.
Not sure if the vapour would re-condense though? Any data to show how this may be possible?

The 'stall mode' is interesting - maybe you could expand on this and table some readings of what is happening.

Chef

Chef
28-09-2009, 03:14 AM
So, to confirm, I'll be observing the SX pressure gauge response, not the evap pressure gauge? (I have two set up).

It is the SH the valves controls on so you want to measure when the superheat again stablises, measuring evap pressure alone may not give conclusive results.

Chef

desA
28-09-2009, 03:15 AM
5 leg distribution!

Would you go with more?



Even though each circuit is equal, the postion within the coil block effects the individual mass flow, as much to do with the inbalance of air flow across the coil, this becomes more noticible on close coupled units (fan and evap) So when testing actual velocity, you measure using agrid pattern to give an averaged face velocity

Add to this, that on this particular lab machine, the outer two pases are partially blocked by the casing 'design'. :eek:

desA
28-09-2009, 03:21 AM
[quote]
Originally Posted by desA
.
It is possible for an evap to be able to boil off more vapour than a compressor can draw off at any instant. If the evap is not tuned properly to the compressor, then some very, very odd things occur. Most likely this vapour re-condenses before the compressor can extract it.

There exists, & I've measured it, a critical point at which the evap goes into a 'stall mode', where it goes daft..



If the evap boils of more than the compressor can 'suck' then the pressure would rise. The valve responds instantly to pressure changes and so would close until it establishes a new equilibrium.

When the pressure rises, what happens to Te,sat?



Not sure if the vapour would re-condense though? Any data to show how this may be possible?

Answer to the question above may shed light on this.



The 'stall mode' is interesting - maybe you could expand on this and table some readings of what is happening.

Try dTlm -> 0/0... This is from measured values & no mathematical tom-foolary... :eek:

Chef
28-09-2009, 03:37 AM
Can you expand more on the concept of 'fragile zone'?

Mass flow equalisation among the 5 splits is another thing all together.
Could you expand a little more on this, please?

The 'fragile zone' is where the 2 phase flow section of the evap is running on a boundary line between 2 differant flow regimes and small changes to the system can allow it to jump from one flow regime to another. The pressure drop will change and the heat transfer coefficients will change.


The 5 feeders to the evap can never have identical conditions in each all the time. The flow entering is 2 phase and so the amount of liquid and vapour entering each one will change, if for instance one pass gets nearly all vapour whilst another pass gets much more liquid the outputs from these passes will be differant manifesting itself as pressure and temperature fluctuations at the bulb and equalisation line. The severity of these fluctuations depends a great deal on the geometry of the header/splitter.
In the worst case a large slug of liquid may even allow liquid to exit one of the passes. This scenario can be mad worse if the evap is 'fragile' as noted above.

This is why the mass flow and evap pressure/Tsat and SH are needed to see where the evap lies.

Chef

mad fridgie
28-09-2009, 03:54 AM
Re 5 legs,
At your present condtions the pressure drop is not to great so leave as is, if you increase the no. of circuits, then you could increase the chance of overfeeding in one or more of the circuits, you again can loose control. If how ever you are looking at working at much lower evaporating pressures, then increasing maybe an option. Lower mass flow, less dense, increased velocity

Chef
28-09-2009, 04:34 AM
[quote]
When the pressure rises, what happens to Te,sat?

Try dTlm -> 0/0... This is from measured values & no mathematical tom-foolary... :eek:

Still not quite sure how the vapour re-condenses - can you explain it fully? Thanks.

Try dTlm -> 0/0 Again maybe you could explain this as I am not at all sure what this alludes too?

Chef

desA
28-09-2009, 05:33 AM
Try dTlm -> 0/0 Again maybe you could explain this as I am not at all sure what this alludes too?

Chef

http://i34.tinypic.com/2nukbnp.png

Yep... the 'oh dang' moment has arrived... it's one of those... :)

desA
28-09-2009, 05:40 AM
[quote=desA;161968]

Still not quite sure how the vapour re-condenses - can you explain it fully? Thanks.

Chef

Well, energy cannot disappear, it has to be used somewhere. The compressor can also not draw more than the volumetric throughput it is designed for.

If the heat-transfer into the evap - as measured - turns out to be some +40% over what the compressor requires, under the evap-pushes-compressor mode, versus ~ 1% under the compressor-pulls evap mode - what gives? The method of calculation in both cases is identical.

Does the compressor suck trough the extra vapour?
If not, where does it go - if it is generated at all?

Are there modes of heat-transfer where the heat-load requirement is greater per mass of refrigerant vapour produced, hence leading to 'spurious' m'g evolution rates?

Moist air Cp is used at all times at prevailing conditions.

Gary
28-09-2009, 07:03 AM
The compressor can also not draw more than the volumetric throughput it is designed for.


It can't? I'm pretty sure that it can.

I vote for the compressor sucks the extra vapor through. The amount of vapor a compressor can pump depends upon how dense that vapor is.

mad fridgie
28-09-2009, 07:27 AM
[

Well, energy cannot disappear, it has to be used somewhere. The compressor can also not draw more than the volumetric throughput it is designed for.

If the heat-transfer into the evap - as measured - turns out to be some +40% over what the compressor requires, under the evap-pushes-compressor mode, versus ~ 1% under the compressor-pulls evap mode - what gives? The method of calculation in both cases is identical.

Does the compressor suck trough the extra vapour?
If not, where does it go - if it is generated at all?

Are there modes of heat-transfer where the heat-load requirement is greater per mass of refrigerant vapour produced, hence leading to 'spurious' m'g evolution rates?

Moist air Cp is used at all times at prevailing conditions.
The displacement may not change but you do get difference in volumetric efficieny,
As your pressure changes then your density changes thus mass flow changes.
If there is a sudden increase change in inlet pressure to the evap, then for a short instance the vapour is below its saturation point thus could condense, I would but this down to a pure piece of therory, I can not see how this could be measure if all other processes are constant.
If the compressor can not suck all the vapor the pressure would rise reducing the pressure differential across the expansion device and reducing the LMTD of the evap thus reducing transfer energy. Equalibrium reached.
I think this what you were asking?
Thanks for the graph "lost me"

desA
28-09-2009, 07:31 AM
^ The compressor is a constant-volume device, is it not - as in m3/h?

If drawing from a certain Te,at & delivering to a Tc,sat, we have an instantaneous pressure ratio, hence mean density (or other place) in compressor.

It could draw more mass-flow if the vapour density changes within the compressor. How would it be able to do that to the tune of +50%?

I'm open to suggestions on this one, to be honest...

It's either that, or the evap is not evolving the mass flow as per latent + superheat on the refrigerant side, but is using the energy somewhere else...

desA
28-09-2009, 07:40 AM
[quote=desA;161976]
The displacement may not change but you do get difference in volumetric efficieny,
As your pressure changes then your density changes thus mass flow changes.
If there is a sudden increase change in inlet pressure to the evap, then for a short instance the vapour is below its saturation point thus could condense, I would but this down to a pure piece of therory, I can not see how this could be measure if all other processes are constant.

Agreed. This would apply only to an 'instantaneous' (transient) change, which would smooth out as the process moves towards thermodynamic equilibrium.



If the compressor can not suck all the vapor the pressure would rise reducing the pressure differential across the expansion device and reducing the LMTD of the evap thus reducing transfer energy. Equalibrium reached.

Te,sat drift answered... :D



Thanks for the graph "lost me"

Chef will know it well. It represents a bifurcation curve generated for a particular case of this same evap. The two branches you see represent two 'stable' modes of evap operation, as a function of the heat-transfer governing equations. The test points are always found to be on one of the branches, even though, in practice, the curves morph slightly with change in evap conditions.

Evaps are non-linear in their nature & hence present some quite extraordinary complications in their control.

If the process ends up on the 'wrong' branch, bad things happen - unstable process (stable evap). If it ends up on the 'other' branch, the process (& evap) is/are stable.

The dTlm='0/0' point is at the bifurcation point i.e. where the single line splits into two (a fork).

It is now fairly clear for 'one' (not all) of the underlying reasons for TXV hunting - the system tries to jump between the two branches on offer...

(A lot of this is pretty much 'hot-off-the-press' as I've been burning the midnight oil developing the theory - in order to best understand the system dynamics).

mad fridgie
28-09-2009, 07:42 AM
I just want to clarify the question,
You are indicating that you are absorbing more energy in the evaporator (upto50%) than can be possibly absorbed by the compressor mass flow calculation.?

desA
28-09-2009, 07:51 AM
I just want to clarify the question,
You are indicating that you are absorbing more energy in the evaporator (upto50%) than can be possibly absorbed by the compressor mass flow calculation.?

Yes... using exactly the same calculation methodology for the evap-push-compressor versus comp-pulls-evap modes of system operation.

The measured evap heat-load, converted via (latent + sensible) in the refrigerant, to m'g [kg/s] is, in some cases +50% more than the compressor can use at the conditions Te,sat, SH, Tc,sat, SC.

desA
28-09-2009, 08:07 AM
I must also state for the record that, these observations, will in all likelihood not be seen under small TD regimes, in the cooler/temperate climates.

Come over to Asia & the high TD gives a hard kick.

I predict a plethora of compressor failures over the next not so many years as more AWHP's are used in Asia - copying European/US technology... :)

mad fridgie
28-09-2009, 08:29 AM
How are you measuring the evap heat load!

Gary
28-09-2009, 08:44 AM
I must also state for the record that, these observations, will in all likelihood not be seen under small TD regimes, in the cooler/temperate climates.

Come over to Asia & the high TD gives a hard kick.

I predict a plethora of compressor failures over the next not so many years as more AWHP's are used in Asia - copying European/US technology... :)

Hmmm... I was thinking that scroll was US technology... and the CPR, too. But I could be wrong.

desA
28-09-2009, 09:05 AM
Hmmm... I was thinking that scroll was US technology... and the CPR, too. But I could be wrong.

Never too sure nowadays - most stuff gets churned out in China... :rolleyes:

desA
28-09-2009, 09:12 AM
How are you measuring the evap heat load!

Ta,in (close to evap face)
Ta,out (rear of evap)
va,in (average over face)
RH%
A,face known

Q'e=m'aw*Cpaw*(Ta,i - Ta,o)

Obviously an amount for water condensed should be brought into the calc, to be tight about it.

The thing that gets to me is that using this method & balancing it against refrigerant loop:

Q'e = m'g*[(1-x)*hfg + Cpv*SH]

provides a mass balance within 1% of the compressor requirement, under stable system conditions.

Under Te,sat drift conditions, the difference can be as much as +50%... go figure where the heat, or mass went to.

:confused:

Chef
28-09-2009, 09:27 AM
mad fridgie

Please would you edit your post number 37.

You have quoted me as saying what was in fact said by DesA.

Thanks very much for that.

Chef

mad fridgie
28-09-2009, 09:55 AM
mad fridgie

Please would you edit your post number 37.

You have quoted me as saying what was in fact said by DesA.

Thanks very much for that.

Chef
Sorry Chef, (have changed now) not sure what happened there, always new it was DesA:o

mad fridgie
28-09-2009, 10:02 AM
Ta,in (close to evap face)
Ta,out (rear of evap)
va,in (average over face)
RH%
A,face known

Q'e=m'aw*Cpaw*(Ta,i - Ta,o)

Obviously an amount for water condensed should be brought into the calc, to be tight about it.

The thing that gets to me is that using this method & balancing it against refrigerant loop:

Q'e = m'g*[(1-x)*hfg + Cpv*SH]

provides a mass balance within 1% of the compressor requirement, under stable system conditions.

Under Te,sat drift conditions, the difference can be as much as +50%... go figure where the heat, or mass went to.

:confused:
I would say your problem lies in the reaction time and accuracy of your instrumentation.
This would account for the difference in your energy mass balance

desA
28-09-2009, 10:41 AM
^ Nope - easy answer.

I'd not think so, to be honest, as the system is held at a constant Tc,sat test point - fairly carefully.

The extreme results are fairly repeatable, in the mode where Te,sat drift occurs.

I'm going to have to sit down & measure each & every stream, water included over a long period of time - so that we can put this one to bed... :)

Chef
28-09-2009, 10:44 AM
Chef will know it well. It represents a bifurcation curve generated for a particular case of this same evap. The two branches you see represent two 'stable' modes of evap operation, as a function of the heat-transfer governing equations. The test points are always found to be on one of the branches, even though, in practice, the curves morph slightly with change in evap conditions.

Evaps are non-linear in their nature & hence present some quite extraordinary complications in their control.

If the process ends up on the 'wrong' branch, bad things happen - unstable process (stable evap). If it ends up on the 'other' branch, the process (& evap) is/are stable.

The dTlm='0/0' point is at the bifurcation point i.e. where the single line splits into two (a fork).

It is now fairly clear for 'one' (not all) of the underlying reasons for TXV hunting - the system tries to jump between the two branches on offer...

(A lot of this is pretty much 'hot-off-the-press' as I've been burning the midnight oil developing the theory - in order to best understand the system dynamics).

desA your coming up with some dramatic new revelations in heat transfer and fluid flow!

First is it possible you can explain the two axis on your graph, as they are calculated you should also include the original formulae. You have previously introduced the alpha and beta coefficients but are they the same in the plot you have posted.

I have not come across a curve like this but have seen heat transfer coefficients change during changes in the flow regime, these are almost parrallel but offset and can cause jumps in the evap condition. These curves are always sloping in the same direction leading to stabliity assuming all alse is industry standard. The flow can jump from one line to another but they always have a positive slope.

Many years ago there was a paper that purported a mythical negative pressure slope where an increase in flow reduced the pressure drop, an interesting concept but it has never been shown in experimental data.

You also mention that evaps are non linear in their behaviour, well condensers, compressors, pressure drops in pipes and TXV's are all non linear but it does not mean they can't be modelled or understood.

Looking forward to graph details.

Chef

desA
28-09-2009, 11:04 AM
desA your coming up with some dramatic new revelations in heat transfer and fluid flow!

Yes, I know this... :)

The idea is intrinsic in every heat-exchanger in existence!!! It is, however, accentuated in a system where a change between flow regimes occurs - such as in an evaporator.

Heat-exchanger design theory has a fundamental flaw - at the dTlm=0/0 point. Even if you look into the derivation of NTU-effectiveness method, it conveniently skips over the dTlm=0/0 point, or even ()/0 for that matter. There are two natures at work.

I first observed this in automotive cross-flow radiators.



First is it possible you can explain the two axis on your graph, as they are calculated you should also include the original formulae. You have previously introduced the alpha and beta coefficients but are they the same in the plot you have posted.

The theory to derive this curve is, in essence, very simple - incredibly simple. I'll have to write it up in a decent note & post it up for us to discuss. The recent incarnation of this Beta theory came out a study in trying to confirm the Magoo rule. You could imagine my dismay, when it produced two answers - every time - sometimes very close to each other - at other times far apart.

It was only when I plotted the curves a few days ago, that the nature of the bifurcation became clearly evident.

The alfa stands for the degree of temperature cross between the air-leaving temp & the refrigerant leaving temp. The beta values come from the Magoo rule:
SH=beta*TD.

There are TWO suitable beta values for every alfa!!! In other words, there are two possible SH values for the evaporator.

If the wrong beta is used, the system, as a whole is unstable - leading to Te,sat drift.



I have not come across a curve like this but have seen heat transfer coefficients change during changes in the flow regime, these are almost parrallel but offset and can cause jumps in the evap condition. These curves are always sloping in the same direction leading to stabliity assuming all alse is industry standard. The flow can jump from one line to another but they always have a positive slope.

I am able to demonstrate the movement between the beta1 (bottom one) & beta2 curves (top one) - experimentally, at will.



You also mention that evaps are non linear in their behaviour, well condensers, compressors, pressure drops in pipes and TXV's are all non linear but it does not mean they can't be modelled or understood.

Looking forward to graph details.


I'm currently developing the condenser instability theory. This is going to be a few levels tougher, in that there are more degrees of freedom.

Non-linear science is fascinating & I'm so stunned that it can evidence itself in a simple HFC-based cycle - & in stable solutions at that.

CO2 trans-critical systems suffer from bi-stability, & this can produce low COP, & high COP cycles, on an apparent
whim. The maths used there is transient in nature & obviously far more complex. This is generally really not necessary, however, as the transient 'stable' states often have a way of dissipating as thermodynamic 'equilibrium' is approached - sometimes not...

desA
28-09-2009, 01:09 PM
Condenser multi-stability bifurcation theory now complete - turned out to be easier than I'd first thought.

I'll work on how to get that to make sense experimentally & we'll open another thread later on that...

:)

Gary
28-09-2009, 03:14 PM
If the compressor can not suck all the vapor the pressure would rise reducing the pressure differential across the expansion device and reducing the LMTD of the evap thus reducing transfer energy. Equalibrium reached.


At the same time, that same pressure rise would increase the density of the vapor entering the compressor, causing the compressor to pump more vapor.

desA
28-09-2009, 03:29 PM
^ It is the evap moving towards its stable point, by raising Te,sat (in order to reduce TD) & the compressor's apparent willingness to move along with it towards higher & higher output power, that is the mechanism at work here.

This mechanism has no self-limiting step (danger lies ahead for high TD's) - unless mechanical blocks, or throttles are put in place to limit the flows. Hence the success of the CPR & valve, on the previous thread - each with it's own 'issues'.

Under the unrestricted scenario, the evap seems to follow the Beta1 curve (the lower one), if fan-controlled (so far this seems to hold), or climb up to the bifurcation point at dTlm=0/0. I have loosely termed this the 'dither point'.

If the throttle is installed, the evaporator is free to float on its own path to limit TD. It then seems to be able to push itself up onto the Beta2 (top curve).

The Magoo Rule firmly places the evaporator on the Beta2 curve!!!

http://www.kueba.de/en-us/Tools/K%C3%BCba-Expansion-Valve-Calculator/Pages/default.aspx

It seems that Kuba has some hidden depth in their products... :)

I'll continue to explore further & report back on the experimental evidence & how these bifurcation curves shift slightly to match the circumstances. At each & every experimental test point computed thus far, at the combination of (alfa, beta), the data WILL lie on either the Beta1, or Beta2 line. A very, very interesting topic.

Gary
28-09-2009, 03:35 PM
Ta,in (close to evap face)
Ta,out (rear of evap)
va,in (average over face)
RH%
A,face known

Q'e=m'aw*Cpaw*(Ta,i - Ta,o)

Obviously an amount for water condensed should be brought into the calc, to be tight about it.

The thing that gets to me is that using this method & balancing it against refrigerant loop:

Q'e = m'g*[(1-x)*hfg + Cpv*SH]

provides a mass balance within 1% of the compressor requirement, under stable system conditions.

Under Te,sat drift conditions, the difference can be as much as +50%... go figure where the heat, or mass went to.

:confused:

I'm not sure where specific heat factors into these equations, but as I recall specific heat is extremely complex and non-linear as well.

desA
28-09-2009, 03:51 PM
Originally Posted by desA
Ta,in (close to evap face)
Ta,out (rear of evap)
va,in (average over face)
RH%
A,face known

Q'e=m'aw*Cpaw*(Ta,i - Ta,o)

Obviously an amount for water condensed should be brought into the calc, to be tight about it.

The thing that gets to me is that using this method & balancing it against refrigerant loop:

Q'e = m'g*[(1-x)*hfg + Cpv*SH]

provides a mass balance within 1% of the compressor requirement, under stable system conditions.

Under Te,sat drift conditions, the difference can be as much as +50%... go figure where the heat, or mass went to.

Gary:
I'm not sure where specific heat factors into these equations, but as I recall specific heat is extremely complex and non-linear as well.

I never knew that, thanks so much for that input. The Cp values are highlighted in the above eqns.

Cp, eh... well, I never. Ok, for moist air - there will be some trading between the air & water vapour. I wonder what gives in the refrigerant vapour? I'll look into this.

The other thing to remember is that there are a number of transient 'stable' states in complex systems, but these are often (should be) damped out by the more stable parts of the system, as it moves towards thermodynamic steady-state... or, do they?

:D

Gary
28-09-2009, 04:00 PM
Some years back, I attempted to devise a simple table plotting inlet superheat vs discharge temp in order to judge the relative efficiency of compressors, the theory being that inefficiency shows up as excess discharge temp.

The complexity of specific heat caused the idea to be scrapped. It became way too complex to be a useful troubleshooting tool.

desA
28-09-2009, 04:23 PM
^ Do you perhaps have any links to the specific heat information that you were using? This would be very interesting.

Gary
28-09-2009, 04:26 PM
^ Do you perhaps have any links to the specific heat information that you were using? This would be very interesting.

It's been a very long time, but I'll see if I can find something.

desA
28-09-2009, 04:39 PM
^ Thanks very much for that... :)

Gary
28-09-2009, 05:24 PM
I haven't been able to find my notes yet, but perhaps this may get you started:

http://hyperphysics.phy-astr.gsu.edu/HBASE/thermo/debye.html

http://hyperphysics.phy-astr.gsu.edu/HBASE/thermo/spht.html#c1

Gary
28-09-2009, 08:15 PM
This mechanism has no self-limiting step (danger lies ahead for high TD's) - unless mechanical blocks, or throttles are put in place to limit the flows. Hence the success of the CPR & valve, on the previous thread - each with it's own 'issues'.

Under the unrestricted scenario, the evap seems to follow the Beta1 curve (the lower one), if fan-controlled (so far this seems to hold), or climb up to the bifurcation point at dTlm=0/0. I have loosely termed this the 'dither point'.

If the throttle is installed, the evaporator is free to float on its own path to limit TD. It then seems to be able to push itself up onto the Beta2 (top curve).

The Magoo Rule firmly places the evaporator on the Beta2 curve!!!

http://www.kueba.de/en-us/Tools/K%C3%BCba-Expansion-Valve-Calculator/Pages/default.aspx

It seems that Kuba has some hidden depth in their products... :)

I'll continue to explore further & report back on the experimental evidence & how these bifurcation curves shift slightly to match the circumstances. At each & every experimental test point computed thus far, at the combination of (alfa, beta), the data WILL lie on either the Beta1, or Beta2 line. A very, very interesting topic.

Actually, the Kuba rule calls for .5*TD to .7*TD.

For rock stable operation, I would suggest using an AEV. However, changes in heat load will result in changes in superheat. Possibly this can be counteracted by fan control, the control points coinciding with .5*TD and .7*TD... although I suspect that lower SH would be entirely acceptable with an AEV in control.

desA
29-09-2009, 02:22 AM
I haven't been able to find my notes yet, but perhaps this may get you started:

http://hyperphysics.phy-astr.gsu.edu/HBASE/thermo/debye.html

http://hyperphysics.phy-astr.gsu.edu/HBASE/thermo/spht.html#c1

These refer to essentially metallic specific heats. I've not yet been able to locate suitable references on liquid/2-phase/gaseous specific heats for refrigerants.

Never fear, at some point, I'll check in the Chemstations database, but that would, in all likelihood be a simple polynomial expansion expression.

desA
29-09-2009, 02:35 AM
Actually, the Kuba rule calls for .5*TD to .7*TD.

http://www.kueba.de/en-us/Tools/K%C3%BCba-Expansion-Valve-Calculator/PublishingImages/Kalkulator-Seite-Schnitt.jpg

Kuba call for a Beta=0.65 setting, as their standard.

http://i34.tinypic.com/2nukbnp.png"]http://i34.tinypic.com/2nukbnp.png

The Beta2 curve corresponds to Beta > 0.48, climbing up to slightly in excess of 0.78.


For rock stable operation, I would suggest using an AEV. However, changes in heat load will result in changes in superheat. Possibly this can be counteracted by fan control, the control points coinciding with .5*TD and .7*TD... although I suspect that lower SH would be entirely acceptable with an AEV in control.

The lower SH (Beta1) curve leads to Te,sat drift - although the fan control will help track on this curve.

Gary
29-09-2009, 02:52 AM
http://www.kueba.de/en-us/Tools/K%C3%BCba-Expansion-Valve-Calculator/PublishingImages/Kalkulator-Seite-Schnitt.jpg

Kuba call for a Beta=0.65 setting, as their standard.



With .5 to .7 as their "ok" range.

desA
29-09-2009, 04:06 AM
With .5 to .7 as their "ok" range.

What you will find is the following:
1. The bifurcation curves tend to 'slide' along the linear part slightly (back, or forth), with system changes, cycle heat-soak & so forth.
2. The bifurcation point is reasonably close to Beta = 0.5, or there-abouts.
3. To set a safe operating point, it would be wise to be some 'distance' away from Beta=0.5.

Kueba seem to be happy at 0.65. Magoo has good experience in the range 0.6 - 0.7.

I would say that both have a solid theoretical, as well as practical basis for their usage. I would stay some distance away from Beta=0.5 as this means a potential constant flipping between two modes of operation - hence system instability.

Gary
29-09-2009, 04:54 AM
Then every A/C system I have ever seen, heard of, or worked on is... wrong.

It is going to take a lot more than bifurcation curves to convince me that a typical A/C coil with 20K/36F TD should have 13K/23.4F superheat.

desA
29-09-2009, 05:30 AM
It is the > 20K, off a small dTlm coil, that I would be concerned about.

In other words, if you operate in a region where TD < 20K, you may not experience excessive temp drift - or at least enough to worry te compressor. Of course, the 'distance' between operating Te,sat & the compressor window limit will also come into play.

If you are going to use low SH, but have a Te,sat initial of say 4.4'C, & final Te,sat of 9K - this will be of little, or no concern.

Remember, too - & this is the nut-cracker - aircon & refrigeration systems are self-governing in terms of the evaporator, in that TD lowers as the cooled space gets colder. The system moves into a 'safe region' as a function of operation.

For an air-driven heat-pump, the external air is outside your control & this is precisely where hot Asian conditions prove to be the 'compressor killers'. These are different beasts & deserve to be treated with some respect.

For the record, I will tell you that this evaporator under study now experiences an initial temperature drift of no more than 1K on the first heat-up run - moves up to around 12-12.5K after extensive heat-cycling & is rock solid thereafter. No tricks, or special requirements, other than correct tuning. The system even sounds 'smooth'.

This has been tested repeatably over many days, under varying TD's. It remains sweet & well-behaved, with no Te,sat drift.

The tuning was driven off the back of the knowledge of the evap bifurcation characteristics. The previous Te,sat control strategies have also been tested against this new concept, & mapped. They are beginning to make sense.

To fully tune the system to where I would like it to remain, will require some innovation from certain equipment manufacturers, with whom I'm now working closely. Once this is place, I'll be more comfortable to begin attempting TD's up to around 35K. This makes for a fairly safe AWHP & opens up new markets.

Gary
29-09-2009, 05:43 AM
Are you still testing with the same system?... the one where the condenser is piped wrong?

desA
29-09-2009, 05:49 AM
Yep...

We can discuss the condenser matters on another thread, in a few weeks. I have to be abroad for a round 10 days & will begin developing alfa-beta-delta curves for the condenser.

The bifuraction theory is complete, but the condenser is a bit more complex to explore practically, since 3 possible pinch points occur. This makes for many more operational options.

Chef
29-09-2009, 05:50 AM
The theory to derive this curve is, in essence, very simple - incredibly simple. I'll have to write it up in a decent note & post it up for us to discuss. The recent incarnation of this Beta theory came out a study in trying to confirm the Magoo rule. You could imagine my dismay, when it produced two answers - every time - sometimes very close to each other - at other times far apart.
..

How are you doing with the write up? Would be interesting to see the methodology you have used.

Chef

Gary
29-09-2009, 06:09 AM
With an AEV holding Te,sat steady, it would be possible to hold SH=.65*TD by controlling fan speed off Tair,in.

desA
29-09-2009, 07:28 AM
How are you doing with the write up? Would be interesting to see the methodology you have used.

Working on it. I'll be abroad for around 10 days, but plan to write up along the way - time permitting.

The concept is growing each day.

desA
29-09-2009, 07:35 AM
With an AEV holding Te,sat steady, it would be possible to hold SH=.65*TD by controlling fan speed off Tair,in.

This is a good idea. I'll continue working up the technology & see how exactly fan speed co-operates under each mode of operation.

In the Beta2 mode, I found fan roll-off to produce some interesting results, as the evaporator held itself steady. Te,sat, of course, drops lower with lower fan speed.

sterl
01-10-2009, 02:12 AM
The original Coil Loading shows a sensible heat ratio of about 62% and a pretty large Delta-T at a high laent....i haven't run the psychormetircs, but with a constant-speed compressor: this thing is going to respond drastically to a small decrease in air flow...Initially TXV is going to undershoot as Suction pressure drops with reduced air flow; then overshoot as refrigerant-dry section of coil becomes incapable of maintaining a temperature below the Dew Point of the air...THIS REALLY NOTICEABLE when a fairly high capacity single coil is working on a fairly small room, that is, volume of air...

As to Flow Regimes:

Suggest you look up articles (multiple) by Mr. Bruce Nelson of Colmac on DX coils....With R-134 at these temperature, you can eliminate about 50% of the phase-momentum regimes unless you have a lot of subcooling at the TXV Inlet.

And a one on one, hermetic or semi-hermetic circuit will quite easily accomodate some short term excess flow of liquid to the evaporator...The motor cools off very quickly.

desA
01-10-2009, 02:54 AM
^ Thanks for your input, Sterl.

Do you perhaps have an idea of the SH/TD ratio recommended by Colmac Coil? (SH = superheat; TD = temperature difference)

Joey.zhang
08-10-2009, 03:49 PM
Ta,in (close to evap face)
Ta,out (rear of evap)
va,in (average over face)
RH%
A,face known

Q'e=m'aw*Cpaw*(Ta,i - Ta,o)

Obviously an amount for water condensed should be brought into the calc, to be tight about it.

The thing that gets to me is that using this method & balancing it against refrigerant loop:

Q'e = m'g*[(1-x)*hfg + Cpv*SH]

provides a mass balance within 1% of the compressor requirement, under stable system conditions.

Under Te,sat drift conditions, the difference can be as much as +50%... go figure where the heat, or mass went to.

:confused:






I think for this instability issue we need pay attention on
1. TXV. Check its performance, its bulb
As you know, TXV, NOT EEV, operates and controls the refrigerant flow depending on SH (including SS). Its response time will effect the system stability directly.
2. Evaporator performance.
--Uneven liquid to the circuits of the Evap. It will also effect the actual SH.

--Heat transfer coefficient. Check if Heat transfer coefficient has been change during the period.
a. Air side, Water drop on the evap.
b. Refrigerant side, mass flow changes, but it affects slight on experience.
3. Liquid return. The chiller will be instable if liquid return happens to compressors.
Other opinion
4. Mass flow exiting the evap must be same to that entering the compressors. It is subject to the suction pressure, that is, SX minus pressues drop, due to the constant suction volume for compressors.
5.I don't think there will be vapour re-condense in the Evaporator per 1st and 2nd law of thermodinamics.

Joey.zhang
08-10-2009, 03:59 PM
I never knew that, thanks so much for that input. The Cp values are highlighted in the above eqns.

Cp, eh... well, I never. Ok, for moist air - there will be some trading between the air & water vapour. I wonder what gives in the refrigerant vapour? I'll look into this.

The other thing to remember is that there are a number of transient 'stable' states in complex systems, but these are often (should be) damped out by the more stable parts of the system, as it moves towards thermodynamic steady-state... or, do they?

:D



I think you need to calculate the enthalpy differece of air in/out instand of their specfic heat.

Q=m*(h'a,out - h' a,in)

On the other hand, it's difficult to test the temperature of air in/out exactly as the air flow or velocity. They will affect enthalpy and mass flow of the air.

desA
12-10-2009, 10:26 AM
^ Thanks Joey, for your excellent input. Much obliged.

desA
12-10-2009, 01:23 PM
Originally Posted by Chef
How are you doing with the write up? Would be interesting to see the methodology you have used.

desA:
Working on it. I'll be abroad for around 10 days, but plan to write up along the way - time permitting.

The concept is growing each day.

A brief update on the stable bifurcation theory.

The theory has actually become pretty advanced, now - with refinement to dimensionless parameters which apply to ANY heat-exchanger. To do the concept justice, I think that it would be preferable to expand the findings into a decent journal publication, or conference proceeding.

Each evap, or condenser I've tested thus far, falls onto an appropriate bifurcation curve - it's becoming pretty interesting.

desA
14-10-2009, 09:27 AM
A challenge

If anyone would like to participate in some research into evap bi-stable phenomena (bifurcation), please pm me the following information (minimum):
1. Te,sat (evap temp at LP - inside evap);
2. Te,sup (evap superheat temp at TXV bulb);
3. Ta,in (air inlet temp to evap);
4. Ta,out (air outlet temp - at evap exit - just behind unit).

Additional information will also be useful in tuning your evap:
a. Evap inlet face area [m2];
b. Evap air inlet velocity [m/s];
c. TXV, capillary, or EEV?

Based on the information received, I'll plot the bi-stability curves for the evap & note where your current system setting lies. We can then discuss ways to improve, or otherwise, the evap performance.

Please note - this information must be extracted from experimental, measured information - not from an evap designer's concept - they must be actual, hard measurements. The evap can come from ANY refrigeration, aircon, or heat-pump circuit - please state, though, which one, with your information.

You can elect to keep the information private & confidential, or to publish the results to this thread. The choice is entirely yours & will be honoured at all times. There is no cost for this service...

:)