Air-driven evaporators - stability issues

[desA]

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]

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...s/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]

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]

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]

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]

Thanks for an excellent review.

^ SX = ?

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]

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]

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]

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]

Thanks for an excellent review.

^ SX = ?

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]

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]

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]

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]

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]

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...

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

.
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]

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]

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'.

[desA]

[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...

[Chef]

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]

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]

[quote=desA;161968]

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

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

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]

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

Chef

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

[desA]

[QUOTE=Chef;161974]

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]

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]

[

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]

^ 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]

[QUOTE=mad fridgie;161984]

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...

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]

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]

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]

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]

How are you measuring the evap heat load!

[Gary]

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]

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...

[desA]

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.

[Chef]

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]

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

[mad fridgie]

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.

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]

^ 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...