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desA
13-07-2009, 02:17 AM
This is a branch off Gary's Refrigeration 101 thread.



Coil outlet superheat varies with type of system. Generally speaking, a freezer should have 6-8F/3.5-4.5K superheat, a cooler should have 8-10F/4.5-5.5K superheat, and an A/C should have 12-16F/6.5-9K superheat.

What figures for evap outlet superheat & condenser sub-cooling should be used for an air-to-water heat-pump?

At what operating temperature would the refrigerant mass charge determination be most appropriate - startup Tc,sat 35'C; or hot end Tc,sat 70'C, or mid range?

I have no given mass charge to work off in this case & have to determine it in-house.

Gary
13-07-2009, 02:25 AM
Cap tubes or TXV's?

desA
13-07-2009, 02:40 AM
^ A TXV system.

Gary
13-07-2009, 02:53 AM
Is this an existing system? Something you are designing?

Gary
13-07-2009, 03:04 AM
I would charge it in cooling mode to obtain 12-16F/6.5-9K superheat with the evap entering temp at design and the evap out temp 20F/11K below design (dT can be altered by adjusting blower speed). And the subcooling at 15F/8.5K.

Then flip it over to heat mode and see if any alterations need to be made.

desA
13-07-2009, 03:10 AM
I would charge it in cooling mode to obtain 12-16F/6.5-9K superheat with the evap entering temp at design and the evap out temp 20F/11K below design (dT can be altered by adjusting blower speed). And the subcooling at 15F/8.5K.

Thanks Gary.



Then flip it over to heat mode and see if any alterations need to be made.

It is an air-to-water heat pump, with the air being used as a heating source. There is no 4-way valve in this system.

Gary
13-07-2009, 03:19 AM
So... this is a water cooled air conditioner?... or maybe a water cooled refrigeration system?

Gary
13-07-2009, 03:21 AM
What would be the expected temp range of the air?

desA
13-07-2009, 03:26 AM
Is this an existing system? Something you are designing?

It is a new machine. The design is complete & the machine is currently in the build stages.

The design was developed around the ARI standards of 11.1K evaporator superheat, 8.33K condenser sub-cooling, as the first pass.

The thing is, practically, my test rig (similar spec, but tube-in-tube condenser - not my design) seems to prefer operating with a slightly lower evap superheat & condenser sub-cooling.

My thoughts on the test rig is that is may be running as it is, due to its being a cobbled-together system I inherited from a now defunct heat-pump manufacturer - it was a suck-&-see system. The evap/compressor/condenser combinations were ill-matched in my view. The other issue was that the previous manufacturer merely loaded in refrigerant as he saw fit - until 'it looks right'. There was no science behind what went in.

My new designs are finely matched in terms of evap/compressor/condenser & I'm looking for a high COP system. To get there, I need to be able to finely-determine the refrigerant mass charge.

The condenser sub-cooling can be optmised a little via the design flow-rate, entry temp for the water stream, as this affects the HX performance.

desA
13-07-2009, 03:31 AM
So... this is a water cooled air conditioner?... or maybe a water cooled refrigeration system?

This is a dedicated water heating system.

In Asia, our local ambient air temperature typically runs in the range of 25-35'C. This provides a useful heating source for heating water for home, hotels & guest-houses.

The unit currently in build can supply enough hot water for a 20 room guest-house, with a greatly reduced electrical consumption, compared to direct electrical heaters. Over here, the only viable heating source is hydro-produced electricity.

Most establishments use element-type heaters. The heat pumps can typically operate in the range of 14-25% of this electrical consumption. They have potential to be great power-savers.


What would be the expected temp range of the air?

Typically runs in the range of 25-35'C.

Gary
13-07-2009, 04:37 AM
The higher temp range on both high side and low side makes for some interesting design challenges.

You say the superheat is set at 11.1K, but at what evap air in and air out temps?... and at what saturated suction temp (SST)?

desA
13-07-2009, 04:46 AM
The higher temp range on both high side and low side makes for some interesting design challenges.

They certainly do. It seems to be a very fine design envelope to work in.



You say the superheat is set at 11.1K, but at what evap air in and air out temps?... and at what saturated suction temp (SST)?

The initial design selection of 11.1K for evap superheat comes from wanting to stick closely to industry standards, for reference purposes. Practically though, the evap, under a more heavily loaded refrigerant charge, at the upper end of the heating cycle, the superheat tends to reduce to something closer to around 5K. This tendency has been noted by other folks for heat-pumps.

The typical saturated suction temp (SST) range for these units is in the order of 10-17.5'C (R-134a) (start - end of heating cycle) - with air on temp around 28-33'C (measure in lab).

Gary
13-07-2009, 05:17 AM
Is there some means of further subcooling the liquid before it enters the TXV? Warm incoming liquid will take up valuable coil area in flashing down to SST.

desA
13-07-2009, 05:27 AM
^ Extremely valuable comment - indeed.

Thoughts:
1. Dedicated sub-coolers are something in the pipeline, based on the results of the early prototype machine. Definitely.
2. Sub-cooling of refrigerant too far closes the gap on useful air temp range before forcing defrost cycles.
3. Sub-cooling aids heat-pump COP,r.

This is a fine balancing act, in the aim for low-cost machines for the region.

Gary
13-07-2009, 05:28 AM
The initial design selection of 11.1K for evap superheat comes from wanting to stick closely to industry standards, for reference purposes. Practically though, the evap, under a more heavily loaded refrigerant charge, at the upper end of the heating cycle, the superheat tends to reduce to something closer to around 5K. This tendency has been noted by other folks for heat-pumps.


Given the high pressure pushing the liquid through, I would expect the TXV to have a strong tendency to hunt at the upper end of the heating cycle?

desA
13-07-2009, 05:40 AM
Given the high pressure pushing the liquid through, I would expect the TXV to have a strong tendency to hunt at the upper end of the heating cycle?

You are correct here. I've selected the TXV fairly tightly in terms of the compressor/TXV load curves.

Another driver for TXV disturbances is that the mass balances for the evap/condenser are different at start of cycle & end of cycle, from a thermodynamic & heat-exchanger design point of view. The system tries to re-establish this balance as the heating cycle progresses.

I use simple bulb-driven TXV's to smooth out any hunting as far as possible (slow, lazy response) & so what comes over are slow, low-amplitude, smooth waves towards the upper end of the heating cycle, rather than short, choppy swings. My test heat-up time is aggressive at 1.75h & so this tends to show up these issues fairly early.

At the top end of the heating cycle, the compressor is close to its upper performance envelope. :)

Gary
13-07-2009, 05:45 AM
I'm thinking a compressor designed for R404A might be more appropriate than one designed for R134A. It doesn't know what it is pumping, it just reacts to the pressures.

desA
13-07-2009, 05:59 AM
^ What's the critical pressure/temp for R404A?

For instance R-134a is ~101'C. The distance between the condenser Tc,sat & the critical temp (top of vapour dome), is what has me using R-134a at the moment. The present design heats water up to 65-70'C.

I'm very open to looking at other low-impact refrigerants which could perhaps provide more compact systems at lower pressure. A very good point - thank you. :)

Gary
13-07-2009, 06:04 AM
^ Extremely valuable comment - indeed.

Thoughts:
1. Dedicated sub-coolers are something in the pipeline, based on the results of the early prototype machine. Definitely.
2. Sub-cooling of refrigerant too far closes the gap on useful air temp range before forcing defrost cycles.
3. Sub-cooling aids heat-pump COP,r.

This is a fine balancing act, in the aim for low-cost machines for the region.

Keeping in mind that design is not my forte, here is how I would handle it:

I would install a vertical suction/liquid heat exchanger (HX).

The suction leaving the evap would enter the bottom of the HX and exit the top. The TXV bulb would be mounted at the exit.

I would install a pressure reducing valve (PRV) in the liquid line to reduce the liquid to a pressure that is more appropriate to R134A, say about 100-110psi.

This reduced pressure liquid would enter the top of the HX wher it would be heavily subcooled and exit the bottom, then on to the TXV.

desA
13-07-2009, 06:13 AM
^ That's very cunning. :)

So in essence:
1. Superheat the evap outlet vapour;
2. TXV controls on evap outlet temp;
3. Sub-cool the condenser exit liquid;
4. PRV reduces liquid pressure, before entry to TXV.

Basically an inter-cooler strapped across suction & liquid lines, plus pressure control pre-TXV.
With the HX, I could then reduce the evap footprint = good!
With the PRV, the TXV will control better.

Now, we're cooking... :)

May need to check on TXV at HX discharge - something I've read is niggling here - may need to use an electronic expansion valve due to an inverse relationship (homework for me).

Gary
13-07-2009, 06:15 AM
^ What's the critical pressure/temp for R404A?

For instance R-134a is ~101'C. The distance between the condenser Tc,sat & the critical temp (top of vapour dome), is what has me using R-134a at the moment. The present design heats water up to 65-70'C.

I'm very open to looking at other low-impact refrigerants which could perhaps provide more compact systems at lower pressure. A very good point - thank you. :)

I would continue to use R134A, but in a compressor that is designed and rated for R404A.

Gary
13-07-2009, 06:19 AM
1. Superheat the evap outlet vapour;
2. TXV controls on evap outlet temp;


No...

1. Superheat the HX outlet vapor.
2. TXV controls on HX outlet temp.

Superheat at compressor inlet remains as is.

desA
13-07-2009, 06:23 AM
I would continue to use R134A, but in a compressor that is designed and rated for R404A.

Can you elaborate further on this idea?

Gary
13-07-2009, 06:26 AM
4. PRV reduces liquid pressure, before entry to TXV.


When the pressure is reduced the liquid will want to flash. That's why it needs to then enter the HX to bring the temp down and return it to subcooled liquid state.

In addition to everything else, the HX, being vertical and bottom fed, will act as an accumulator in the off cycle.

desA
13-07-2009, 06:26 AM
No...

1. Superheat the HX outlet vapor.
2. TXV controls on HX outlet temp.

Superheat at compressor inlet remains as is.

Ok... that's fine. (My head was thinking of HX outlet monitoring, but my fingers typed evap). :o

The compressor inlet is the one that has to be held - agreed.

Gary
13-07-2009, 06:32 AM
Can you elaborate further on this idea?

Where a compressor that is designed for R134A under such high pressures might be close to kicking out on it's internal overload, a compressor that is designed for R404A would see these pressures as normal.

desA
13-07-2009, 06:46 AM
^ That's a very interesting point. I'll look into it & feed back.

I've tested the current compressor range beyond its normal envelope conditions without the internal kick-out activating. This is done by gradually increasing the Hi/Lo pressure cutout switch settings. These compressors are usually set around 10-15% higher for the internal cut-out, than the current operating range I'm working in. So there is still some margin left.

desA
13-07-2009, 07:49 AM
When the pressure is reduced the liquid will want to flash. That's why it needs to then enter the HX to bring the temp down and return it to subcooled liquid state.

In addition to everything else, the HX, being vertical and bottom fed, will act as an accumulator in the off cycle.

This idea is excellent. I like the pressure control & accumulator aspects. Thanks so much for your wisdom.

Gary
13-07-2009, 04:58 PM
^ That's a very interesting point. I'll look into it & feed back.

I've tested the current compressor range beyond its normal envelope conditions without the internal kick-out activating. This is done by gradually increasing the Hi/Lo pressure cutout switch settings. These compressors are usually set around 10-15% higher for the internal cut-out, than the current operating range I'm working in. So there is still some margin left.

Given improvements in heat transfer in the evap (better TXV control, lower superheat, minimal flashing, maximized fan speed, etc.) the SST should rise and that margin could quickly disappear. You may need a compressor with a higher operating (pressure) range.

desA
13-07-2009, 05:13 PM
^ Thanks Gary.

Is this where you'd be looking for a R404A compressor, but pumping R-134a?

The upper safe envelope limit for the current R-134a compressor is given as Tc,sat of 75'C, at Te,sat of 15'C - top right point.

This operating point is always a major top end issue to a standard heat-pump compressor & it would be nice to be able to run up a bit higher without problems. The discharge line temps & compressor base temps are currently well within the manufacturer's limits listed in their technical spec sheets - it's that upper temp limit of 75'C that concerns me.

I've taken precautions on the positioning of the compressor to keep the casing reasonably cool - looking to push the envelope margin a little, if necessary.

Gary
13-07-2009, 05:28 PM
^ Thanks Gary.

Is this where you'd be looking for a R404A compressor, but pumping R-134a?

Yes.


The upper safe envelope limit for the current R-134a compressor is given as Tc,sat of 75'C, at Te,sat of 15'C - top right point.

This operating point is always a major top end issue to a standard heat-pump compressor & it would be nice to be able to run up a bit higher without problems. The discharge line temps & compressor base temps are currently well within the manufacturer's limits listed in their technical spec sheets - it's that upper temp limit of 75'C that concerns me.

I've taken precautions on the positioning of the compressor to keep the casing reasonably cool - looking to push the envelope margin a little, if necessary.

Improvements in evap efficiency should have little to no effect on the high side limits (this is primarily a matter of water temp and flow volume) and the lower more stable superheat should in fact give you a lower discharge temp.

However, improvements in evap efficiency should result in higher refrigerant mass flow which equates to higher amp draw. It is the amperage limit and consequent tripping of the internal overload which we will be running up against.

Gary
13-07-2009, 05:39 PM
Te,sat of 15'C...

This is the limit that concerns me.

Gary
13-07-2009, 05:53 PM
The typical saturated suction temp (SST) range for these units is in the order of 10-17.5'C (R-134a) (start - end of heating cycle) - with air on temp around 28-33'C (measure in lab).

And apparently you are already exceeding the Te,sat 15C (SST) limit at the end of the cycle.

desA
14-07-2009, 02:59 AM
And apparently you are already exceeding the Te,sat 15C (SST) limit at the end of the cycle.

True & it does concern me. These results are from the test machine in my laboratory.

I originally considered the main reasons for this Te,sat value to be due to:
1. The evap is too large for the heat-pump service (by some 60% according to my calculations);
2. Local ambient temps are very high, pulling the Te,sat (SST) up towards the end of cycle;
3. Current refrigerant mass charge in test rig is not yet at optimal value. (With lower charge values Te,sat is reduced slightly);
4. The lab machine evap/compressor/condenser/piping are mis-matched (quite a large discrepancy, according to my calculations).

As I understand - from the compressor manufacturer's tech sheets - the right line of the compressor envelope may be crossed if compressor shell cooling is implemented.

For the record, the supplier of my lab machine, had supplied a number of units abroad, into hot climates, with some measure of success. (It was his business model that failed, with unscrupulous partners, rather than poor product.)

desA
14-07-2009, 04:03 AM
These Asian-condition AWHP machines are a very fine design balance, at the outer limits of the compressor envelope. Of that, there is no doubt.

What some folks will do, is to limit the achievable hot-water upper temperature limit. This will move the Te,sat (SST) & Tc,sat values to just within the published compressor envelope.

In this game, oil control is a must. Pipes are carefully sized & laid to ensure correct oil return.

Magoo
14-07-2009, 04:06 AM
Hi desA
Interesting post/topic, with evap superheat testing, setting the TEV is critical for system performance. Start by reading the air on temp., and the actual evap pressure converted to temp., this is system TD. Then read suction temp at TEV bulb versus the evap pressure/temp. The superheat of TEV should be 60 > 70 % of system TD. Can be set up during pull down or at design, any adjustments to TEV wait 15 minute for TEV to stabilize. Doing this method of checking gets rid of all the "rule of thumb " ideas
I would stick with R134a, but with high suction temps would consider a TEV injecting liquid before an accumulator to cool suction gas entry to compressor, reduces motor winding temps, and discharge temps.. to a controllable level. High discharge temps kill oil quality, and even consider fitting a head cooling fan and oil cooler coil above heads as well.

magoo

desA
14-07-2009, 06:26 AM
Thanks so much Magoo for your post.


... with evap superheat testing, setting the TEV is critical for system performance. Start by reading the air on temp., and the actual evap pressure converted to temp., this is system TD. Then read suction temp at TEV bulb versus the evap pressure/temp. The superheat of TEV should be 60 > 70 % of system TD.

For example:
Air on temp (Ta,in) = 28.1'C
Evap sat temp (Te,sat) = 14.0'C
TD = Ta,in - Te,sat = 28.1 - 14.0 = 14.1'C
0.6*TD = 0.6*14.1=8.46'C
0.7*TD = 0.7*14.1=9.87'C

So, acceptable superheat in range 8.5-9.9'C.


Can be set up during pull down or at design, any adjustments to TEV wait 15 minute for TEV to stabilize. Doing this method of checking gets rid of all the "rule of thumb " ideas

At which operating point in the heat-up cycle would you optimise the superheat settings?

For example, for a hot-water heat-pump, the vapour compression cycle moves from a condenser Tc,sat of around 35'C up to 70-75'C, for instance. During this excursion, the whole vapour compression cycle has to continually re-position itself - everything is in a state of dynamic instability & has to continually re-adjust itself to track the hot water temperature.

Where in this range would be the best place to optimise the system settings, charge determination & so forth?


I would stick with R134a, but with high suction temps would consider a TEV injecting liquid before an accumulator to cool suction gas entry to compressor, reduces motor winding temps, and discharge temps.. to a controllable level. High discharge temps kill oil quality, and even consider fitting a head cooling fan and oil cooler coil above heads as well.

Excellent points - well taken. Thank you.

At the moment, in a high-load test, the base of the compressor casing runs at 45.3'C, in the test rig - this has not gone above 55'C - well under the compressor manufacturer's guidelines for safe operation. The new machine has the compressor temps well managed.

Currently, the maximum discharge line temps are well under the allowable temp limit suggested by the manufacturer - they are being very closely monitored.

Gary
14-07-2009, 06:55 AM
Since the evap is not in fact being used for comfort cooling, I see no reason the compressor cannot be mounted in the path of the evap leaving air.

desA
14-07-2009, 07:36 AM
^ Touche... :)

desA
14-07-2009, 05:38 PM
What are the pro's and con's of introducing a dedicated de-superheater & sub-cooler into the system, instead of trying to do sub-cooling, condensing, sub-cooling in one single condenser unit?

Gary
14-07-2009, 05:50 PM
If the compressor is close coupled (short suction line) there is no need for de-superheating as the compressor inlet superheat will be the same as the evap outlet superheat. De-superheating is for long suction lines.

Gary
14-07-2009, 05:56 PM
Dedicated subcooling assumes an external source of cooling for the liquid. What exactly did you have in mind?

Gary
14-07-2009, 06:07 PM
A possible source of cooling for the liquid would be the evap leaving air. A spiral of copper tubing in the path of the cool air could act as a subcooler and if sized properly could provide just enough restriction to drop the liquid pressure, thus eliminating the PRV.

desA
15-07-2009, 02:50 AM
Dedicated subcooling assumes an external source of cooling for the liquid. What exactly did you have in mind?

Use the incoming water to first sub-cool in the sub-cooler, before passing into the condenser.

The thought process here is whether a stand-alone liquid sub-cooler/water pre-heater is more efficient/cost-effective than the sub-cooler integrated as part of the condenser itself.

Essentially, this same logic applies to a dedicated de-superheater.

So, in essence, which option is more efficient/cost-effective:
1. Integrated de-superheater/condenser/sub-cooler (single HX);
2. Separate de-superheater + condenser + sub-cooler (3 HX's)

desA
15-07-2009, 02:53 AM
A possible source of cooling for the liquid would be the evap leaving air. A spiral of copper tubing in the path of the cool air could act as a subcooler and if sized properly could provide just enough restriction to drop the liquid pressure, thus eliminating the PRV.

That's brilliant! I love that idea... :)

Now, that has got to be looked into. Using a air 'waste stream' to do further duty. You've made my day. Thank you so much.

Gary
15-07-2009, 04:08 AM
Use the incoming water to first sub-cool in the sub-cooler, before passing into the condenser.


What temperature is the incoming water?

desA
15-07-2009, 05:54 AM
What temperature is the incoming water?

This can be tuned by adjusting the water flowrate to suit the system.

Water-heating systems trade water temp rise & flow-rate in using the condenser heat output. The flowrate selected will determine the number of water passes through the heat-pump (heating cycles).

In other words, on the one extreme, (1) you have water moving from 20'C to 65'C in one pass through the heat-pump i.e. low volume flowrate, high dT,water.

The other extreme is (2) water coming in at say 60'C & leaving at say 65'C, at high flowrate.

A sub-cooler would be useful in case (1) & not much use in case (2), for instance.

So, to answer your question, with a Tc,sat ~ 70'C, & typical sub-cooling of 8.33K, Tc,sub ~ 61.67'C. Allowing for an entry approach of 5K (good HX), this means that the entry water could be around 56.67'C, or slightly higher - even up to 60'C with an excellent HX.

If the water flowrate is now reduced, allowing for less passes through the condenser (less heating cycles), then a lower inlet temp can be selected. Say we use Tw,in=45-50'C for example, then we can further sub-cool the refrigerant by using the incoming water stream.

Heating circuits are a balancing act & this can be used to advantage.

Gary
15-07-2009, 02:31 PM
Are the condensers counterflow or crossflow?

Gary
15-07-2009, 02:48 PM
Water-heating systems trade water temp rise & flow-rate in using the condenser heat output.

This is not an even trade. Lower flowrate will give you higher temp water, but far less heat transfer. It is a mistake.

desA
15-07-2009, 02:50 PM
Are the condensers counterflow or crossflow?

For most of these AWHP's, the condensers are either tube-in-tube coils, or plate heat-exchangers. Both can be treated in essence as counterflow units.

For a condensing phase unit, the cross-flow correction factor is taken as unity, in the phase-change region. It is only when you get to the sensible heat-transfer regions (de-superheating & sub-cooling), that the effect of cross-flow, counter-flow, or parallel flow, is noticed.

In general, the HX's are taken as counterflow.

desA
15-07-2009, 02:58 PM
This is not an even trade. Lower flowrate will give you higher temp water, but less heat transfer. It is a mistake.

This is true, for the condenser heat-balance, as we saw in Drew's swimming pool heat-pump.

The best way to actually test the real effect of these changes is to measure the heat-up time for water cycling around a heat-pump loop. In practice, I've actually found little noticeable difference in tank heat-up time using slower, or faster flows.

What happens is that the lower temperature of the entering water can be set to maintain the same log-mean-temp-difference across the condenser, so that the condensing temp is not actually affected.

In Drew's case, my test was set a Tc,sat=50'C, then the water flow was closed off - the Tc,sat rose to compensate. This is not a realistic test for circulating water, though. What typically happens is that at a similar Tc,sat~70'C the whole heating cycle finishes - with less passes through the heat-pump circuit, but more dTw in each pass.

Gary
15-07-2009, 03:41 PM
This is true, for the condenser heat-balance, as we saw in Drew's swimming pool heat-pump.

The best way to actually test the real effect of these changes is to measure the heat-up time for water cycling around a heat-pump loop. In practice, I've actually found little noticeable difference in tank heat-up time using slower, or faster flows.

What happens is that the lower temperature of the entering water can be set to maintain the same log-mean-temp-difference across the condenser, so that the condensing temp is not actually affected.

In Drew's case, my test was set a Tc,sat=50'C, then the water flow was closed off - the Tc,sat rose to compensate. This is not a realistic test for circulating water, though. What typically happens is that at a similar Tc,sat~70'C the whole heating cycle finishes - with less passes through the heat-pump circuit, but more dTw in each pass.

Edit: Oops... I need to re-think my statement... I'll just remove it and maybe nobody will notice... lol

Gary
15-07-2009, 03:43 PM
For most of these AWHP's, the condensers are either tube-in-tube coils, or plate heat-exchangers. Both can be treated in essence as counterflow units.

For a condensing phase unit, the cross-flow correction factor is taken as unity, in the phase-change region. It is only when you get to the sensible heat-transfer regions (de-superheating & sub-cooling), that the effect of cross-flow, counter-flow, or parallel flow, is noticed.

In general, the HX's are taken as counterflow.

Counterflow and crossflow can give us equal heat transfer, but counterflow gives us a water temperature gradient, which is far better for our purposes.

Gary
15-07-2009, 03:48 PM
I am trying to envision the water loop. The water leaves the condenser, travels around the point of use areas, then to a storage tank, then back to the condenser. Is this correct?

Where in this loop is the pump? Where is the makeup water? Where is the expansion tank?

I'm thinking a piping diagram would be useful here.

Gary
15-07-2009, 04:34 PM
This is true, for the condenser heat-balance, as we saw in Drew's swimming pool heat-pump.

The best way to actually test the real effect of these changes is to measure the heat-up time for water cycling around a heat-pump loop. In practice, I've actually found little noticeable difference in tank heat-up time using slower, or faster flows.

What happens is that the lower temperature of the entering water can be set to maintain the same log-mean-temp-difference across the condenser, so that the condensing temp is not actually affected.

In Drew's case, my test was set a Tc,sat=50'C, then the water flow was closed off - the Tc,sat rose to compensate. This is not a realistic test for circulating water, though. What typically happens is that at a similar Tc,sat~70'C the whole heating cycle finishes - with less passes through the heat-pump circuit, but more dTw in each pass.

Okay... let's try this again:

As water flow is increased there should be more heat transfer (decrease in SCT), however... every condenser has a point where a further increase in flow will result in equal or less heat transfer rather than more. If you can reach this point, your pump is oversized for the condenser.

To put it more simply, if lowering the water flow results in the same heat transfer, then you can save energy and initial cost by lowering it permanently (smaller pump). And if lowering the water flow beyond this point lowers the heat transfer, then why do it? Just enough and not too much.

Gary
17-07-2009, 06:37 AM
Basically an inter-cooler strapped across suction & liquid lines, plus pressure control pre-TXV.
With the HX, I could then reduce the evap footprint = good!


I would suggest that once the capacity of the coil is maximized, the coil could be sized such that ambient temp of 25C would result in Te,sat of 15C thus riding the upper limits of the compressor.

Of course, as the ambient rises the Te,sat will try to rise. So... with a variable speed fan, the airflow can be slowed to bring the Te,sat back down, holding it at a steady 15C, riding the compressor limits throughout the entire ambient range (25C-35C) and beyond.

As a back-up precaution to safeguard the compressor, I would suggest a crankcase pressure regulating valve (CPR).

desA
17-07-2009, 12:43 PM
I am trying to envision the water loop. The water leaves the condenser, travels around the point of use areas, then to a storage tank, then back to the condenser. Is this correct?

Where in this loop is the pump? Where is the makeup water? Where is the expansion tank?

I'm thinking a piping diagram would be useful here.

There are a number of variations on this water loop concept. Each system designer has their own ideas for what they prefer.

From a heat-pump perspective, the simplest to work with, conceptually, is as follows:
Tank->pump->condenser->tank. (Pump-around loop)

The water make-up feeds into the tank on level, as separate water pump can be used to pull out of the tank, as hot water is needed. This is a mixed tank concept & is inexpensive.

Some folks prefer different systems of water storage - some pressurised, others not. The least expensive route used over here is a simple pump-around loop - much like the petrochemical industry.

desA
17-07-2009, 12:53 PM
Okay... let's try this again:

As water flow is increased there should be more heat transfer (decrease in SCT), however... every condenser has a point where a further increase in flow will result in equal or less heat transfer rather than more. If you can reach this point, your pump is oversized for the condenser.

To put it more simply, if lowering the water flow results in the same heat transfer, then you can save energy and initial cost by lowering it permanently (smaller pump). And if lowering the water flow beyond this point lowers the heat transfer, then why do it? Just enough and not too much.

I agree with going for as small a pump-around pump as is necessary to just do the job - with some safety margin for wear & tear. It makes no sense to waste pumping power.

There seem to be two very different schools of though when it comes to heat-pumps:
1. Set dT,water across condenser in range of 3-5K & adjust water flow-rate accordingly (pump-around);
2. Pass water through heat-pump only 1 time - raise from say 20-65'C in one pass. Here the water flowrate is incredibly small.

With (1), the heat-pump cycle is changing continually throughout the heat-up cycle, whereas with (2), the heat-pump run is a steady-state operation.

If anyone can clearly state which option is the best from overall energy conservation & operability points of view, the frosties will be on their way. :D

desA
17-07-2009, 01:44 PM
I would suggest that once the capacity of the coil is maximized, the coil could be sized such that ambient temp of 25C would result in Te,sat of 15C thus riding the upper limits of the compressor.

Of course, as the ambient rises the Te,sat will try to rise. So... with a variable speed fan, the airflow can be slowed to bring the Te,sat back down, holding it at a steady 15C, riding the compressor limits throughout the entire ambient range (25C-35C) and beyond.

Touche'... You tell me you're not a designer... :D



As a back-up precaution to safeguard the compressor, I would suggest a crankcase pressure regulating valve (CPR).

What types of compressors would this work on?
How does this work? I've never worked with these yet.

Gary
17-07-2009, 02:13 PM
There are a number of variations on this water loop concept. Each system designer has their own ideas for what they prefer.

From a heat-pump perspective, the simplest to work with, conceptually, is as follows:
Tank->pump->condenser->tank. (Pump-around loop)

The water make-up feeds into the tank on level, as separate water pump can be used to pull out of the tank, as hot water is needed. This is a mixed tank concept & is inexpensive.

It seems to me that you would want to feed the make-up water directly to the heat pump to avoid cooling the tank water.

tank > check valve > feed water > pump > condenser > tank

desA
17-07-2009, 03:41 PM
^ That's what can be done, in practice. Would need a little more instrumentation - although not really a problem. It will only differ by (T,in+dT,w,cond) though.

The other problem with storage tank mixing, is that all the water ends up at a single mixed temp.

Some folks look at layered storage tanks which feed top-to-bottom in series. The colder water goes to the heat-pump & the warm water stays in the hot part - ready for use. There is a temp gradient across the tank. Problem is though, that the slightest inlet disturbance & the tank partially mixes. Not everyone likes this method.

Horses for courses.

Gary
17-07-2009, 04:00 PM
Touche'... You tell me you're not a designer... :D

I have zero experience in system design, but several decades of making problem systems work reasonably well despite the design, turning lemons into lemonade.

I find this particular project fascinating in that the temperature/humidity of the evaporator leaving air is not the end product... and that changes everything.

Gary
17-07-2009, 04:05 PM
What types of compressors would this work on?
How does this work? I've never worked with these yet.

A CPR valve places an upper limit on its outlet pressure, which is the inlet pressure for the compressor. It can be adjusted such that the maximum load can be placed on the compressor without overloading it.

desA
17-07-2009, 04:34 PM
A CPR valve places an upper limit on its outlet pressure, which is the inlet pressure for the compressor. It can be adjusted such that the maximum load can be placed on the compressor without overloading it.

Can this be applied to a scroll compressor?

Is this a valve on the suction line, then? If so, how is it set, or where does it take it's control signal from?

desA
17-07-2009, 04:37 PM
I find this particular project fascinating in that the temperature/humidity of the evaporator leaving air is not the end product... and that changes everything.

AWHP's are fascinating beasts & that is what got me into the technology in the first place. They are a little counter-intuitive at first, but, once you get going, they're a fascinating design challenge.

To make the machines in ultra-compact format is a real challenge - especially given the size of the current evaporator technology (archaic technology).

Gary
17-07-2009, 05:07 PM
Can this be applied to a scroll compressor?

Is this a valve on the suction line, then? If so, how is it set, or where does it take it's control signal from?

Yes... it is on the suction line near the compressor.

As far as I know, the CPR is applicable for any type of compressor.

There is an adjustment stem on the CPR. Its control/sensing is mechanical and internal.

The fan speed strategy and CPR valve are redundant. Either could be used to optimize the compressor load.

The fan speed strategy has the added benefit of reduced energy for the fan motor. However, if the fan stuck in the full speed position in high ambients this could be disastrous for the compressor. Its a belt and suspenders kinda thing.

As to type of compressor, the only type I would tend to eliminate out of hand would be the rotary, because the compressor shell is part of the high side and runs very hot. We don't want to lose that heat to the air surrounding the compressor. We want that heat to go to the condenser.

What type of compressor are you currently using?

desA
18-07-2009, 08:47 AM
What type of compressor are you currently using?

Scroll compressor - reputable brand.

Gary
18-07-2009, 02:59 PM
Another load limiting strategy would be a TXV w/MOP (maximum operating pressure) charge in the power element.

Given the desire to ride the upper limits of the compressor, I'm surprised that some form of load limiting is not commonly used in these systems.

Gary
18-07-2009, 04:49 PM
There seem to be two very different schools of though when it comes to heat-pumps:
1. Set dT,water across condenser in range of 3-5K & adjust water flow-rate accordingly (pump-around);
2. Pass water through heat-pump only 1 time - raise from say 20-65'C in one pass. Here the water flowrate is incredibly small.

With (1), the heat-pump cycle is changing continually throughout the heat-up cycle, whereas with (2), the heat-pump run is a steady-state operation.


It seems to me that filling/heating the water in the storage tank would be a one time thing.

From that point on we are heating the feed water and/or maintaining the storage tank temperature.

This being the case, I vote for strategy #2, with the system automatically switching between the former and latter duties in accordance with the water level in the tank.

feed water > check valve > condenser > flow regulator > tank

-OR-

tank > pump > check valve > condenser > flow regulator > tank

The flow regulator would automatically restrict the water flow to maintain a constant 65C leaving water temp.

Gary
20-07-2009, 08:21 PM
There seem to be two very different schools of though when it comes to heat-pumps:
1. Set dT,water across condenser in range of 3-5K & adjust water flow-rate accordingly (pump-around);
2. Pass water through heat-pump only 1 time - raise from say 20-65'C in one pass. Here the water flowrate is incredibly small.

With (1), the heat-pump cycle is changing continually throughout the heat-up cycle, whereas with (2), the heat-pump run is a steady-state operation.


I would contend that the energy consumed is pretty much the same either way, with one exception: The energy consumed in pumping the water.

If the feed water is fully heated before it reaches the storage tank, then the local water supplier has paid to pump it, therefore not running the tank water pump to heat the feed water is a savings to the end user.

Gary
20-07-2009, 09:28 PM
Of course, as the ambient rises the Te,sat will try to rise. So... with a variable speed fan, the airflow can be slowed to bring the Te,sat back down, holding it at a steady 15C, riding the compressor limits throughout the entire ambient range (25C-35C) and beyond.


Thus far, assuming everything works as envisioned, we have a system which absorbs a very stable amount of heat in the low side.

In the high side, a high percentage of that heat is transferred to the water in the condenser, while a small percentage is transferred back to the waste air stream via the pressure reduction coil.

Perhaps we can control the variable speed fan to sense/minimize this waste heat, dropping the SST to pump only the heat that can be currently utilized by the condenser.

Then look for ways to improve the condenser heat transfer.

Gary
25-07-2009, 04:29 AM
It occurs to me that the pressures and loads being stable, the system is now ideal for a cap tube. A cap tube is not only less expensive, but it uses less refrigerant.

And the fan can be controlled off discharge line temp using a thermistor, rather than using a more expensive transducer to control off low side pressure.

Somehow, this beast keeps evolving... lol

desA
26-07-2009, 02:01 PM
Apologies for not replying earlier - I've been away for a few days.


Another load limiting strategy would be a TXV w/MOP (maximum operating pressure) charge in the power element.


Can you explain more on how the MOP option works? I've seen the option on offer - although not common in my present location.



Given the desire to ride the upper limits of the compressor, I'm surprised that some form of load limiting is not commonly used in these systems.

I agree here. If you look at some of the products coming out of Asia, I'll be very surprised if their compressors last any decent time at all.

Conservative design is absolutely essential, if the compressors are to have a decent lifetime.

Gary
26-07-2009, 03:49 PM
Can you explain more on how the MOP option works? I've seen the option on offer - although not common in my present location.


Where there is both liquid and vapor in an enclosed container an increase in temperature will cause an increase in pressure.

This relationship continues until all of the liquid becomes vapor, at which point the pressure becomes fixed regardless of any further increase in temperature.

By precisely manipulating the amount of refrigerant in the TXV bulb, a fixed bulb pressure limit can be set.

The TXV judges superheat by comparing the pressure in the coil to the pressure in the bulb.

An increase in coil pressure, when compared to a fixed bulb pressure is interpreted as a decrease in superheat, which tends to reduce refrigerant flow, which in turn reduces the coil pressure.

Thus equilibrium is reached at a predetermined coil pressure. The coil is at its maximum operating pressure.

When the evaporator load decreases, the bulb temperature decreases, liquid droplets form in the bulb and everything goes back to normal.

Magoo
27-07-2009, 04:18 AM
Hi desA
The primary consideration is compressor operational conditions, for longevity of system integraty.
Did the TEV superheat check I suggested stabalize performance, it has worked for me for decades.

desA
27-07-2009, 04:05 PM
Hi desA
The primary consideration is compressor operational conditions, for longevity of system integraty.
Did the TEV superheat check I suggested stabalize performance, it has worked for me for decades.


Hi desA
Interesting post/topic, with evap superheat testing, setting the TEV is critical for system performance. Start by reading the air on temp., and the actual evap pressure converted to temp., this is system TD. Then read suction temp at TEV bulb versus the evap pressure/temp. The superheat of TEV should be 60 > 70 % of system TD. Can be set up during pull down or at design, any adjustments to TEV wait 15 minute for TEV to stabilize. Doing this method of checking gets rid of all the "rule of thumb " ideas
magoo

Hi Magoo,

Thanks so much for your follow-up. I'll be running up my lab machine tomorrow, with a pre-calculated mass charge. I'll apply your TEV rules in setting the superheat & report back on the performance.

Gary
27-07-2009, 05:24 PM
By precisely manipulating the amount of refrigerant in the TXV bulb, a fixed bulb pressure limit can be set.


I should add that for your purposes, given the non-conventional pressures you need in order to maximize coil heat absorption, you would probably need to special order your MOP charge.

Gary
27-07-2009, 05:56 PM
You may have noticed that refrigeration coils are sized for TD's of 10-15F/5.5-8.5K, while A/C coils are sized for 35-40F/20-22K TD's. There is a very good reason for this: If an A/C coil were sized for 10-15F/5.5-8.5K TD it would be incapable of achieving acceptable humidity levels. Your home would be a cold swamp.

You have no such dehumidification needs in this system, therefore you can achieve much higher COP by sizing your coil in accordance with refrigeration practices as opposed to A/C practices.

Gary
27-07-2009, 06:37 PM
On the other hand, conventional sizing calculations assume a portion of the coil is used for flashing off the liquid. Since we are taking steps to eliminate flashing this changes everything.

I'm thinking you are going to have to size and adjust every component through a step-by-step trial and error process.

On the bright side, at the end of this process you may be in a position to devise your own set of unique formulas for this particular industry niche.

desA
28-07-2009, 09:27 AM
I would contend that the energy consumed is pretty much the same either way, with one exception: The energy consumed in pumping the water.

If the feed water is fully heated before it reaches the storage tank, then the local water supplier has paid to pump it, therefore not running the tank water pump to heat the feed water is a savings to the end user.

A very useful thought... :D

I wonder how many users will still prefer to pump around on a mixed tank, rather than plan the inventory properly & heat incoming.

desA
28-07-2009, 01:39 PM
Thus far, assuming everything works as envisioned, we have a system which absorbs a very stable amount of heat in the low side.

In the high side, a high percentage of that heat is transferred to the water in the condenser, while a small percentage is transferred back to the waste air stream via the pressure reduction coil.



Perhaps we can control the variable speed fan to sense/minimize this waste heat, dropping the SST to pump only the heat that can be currently utilized by the condenser.

This is very cunning. I've sourced a range of fan speed controllers that can allow fan speed optimisation. I'll work further on this - very interesting.



Then look for ways to improve the condenser heat transfer.

The current condensers are already pretty good in terms of their heat-transfer capacity. Very, very compact. This, plus a few other tricks, has allowed the current prototype model to reduce down to 54% of the previous box volume. Practically, this can go down a fair bit further still, judging from the piping layout.

I'm absolutely loving the challenge so far. It always gets interesting when you're spending your own money on such development projects. :D

desA
28-07-2009, 01:44 PM
It occurs to me that the pressures and loads being stable, the system is now ideal for a cap tube. A cap tube is not only less expensive, but it uses less refrigerant.

This is a very useful observation, indeed. A change of heating philosophy leads to a more compact, simpler system. Thanks for that.



And the fan can be controlled off discharge line temp using a thermistor, rather than using a more expensive transducer to control off low side pressure.


Agreed. I do prefer this method. My current prototype has allowance for two active thermistors. The second can be put to good use here.

This is fun... :D

desA
28-07-2009, 01:50 PM
Where there is both liquid and vapor in an enclosed container an increase in temperature will cause an increase in pressure.

This relationship continues until all of the liquid becomes vapor, at which point the pressure becomes fixed regardless of any further increase in temperature.

By precisely manipulating the amount of refrigerant in the TXV bulb, a fixed bulb pressure limit can be set.

The TXV judges superheat by comparing the pressure in the coil to the pressure in the bulb.

An increase in coil pressure, when compared to a fixed bulb pressure is interpreted as a decrease in superheat, which tends to reduce refrigerant flow, which in turn reduces the coil pressure.

Thus equilibrium is reached at a predetermined coil pressure. The coil is at its maximum operating pressure.

When the evaporator load decreases, the bulb temperature decreases, liquid droplets form in the bulb and everything goes back to normal.

Thanks so much for the detailed explanation. That would be quite some juggling act the bulb-manufacturer has to take care of. I'd imagine that the bulb location would also be critical here.

desA
28-07-2009, 01:53 PM
I should add that for your purposes, given the non-conventional pressures you need in order to maximize coil heat absorption, you would probably need to special order your MOP charge.

Ok, that will probably make for an interesting supply discussion.

Which manufacturers of TXV's would be open to supply optimised MOP options? In my view, this would actually be a useful service component in terms of service support. Perhaps also give a bit of competitive edge as well.

Some very interesting thoughts coming out of this discussion.

desA
28-07-2009, 02:01 PM
You may have noticed that refrigeration coils are sized for TD's of 10-15F/5.5-8.5K, while A/C coils are sized for 35-40F/20-22K TD's. There is a very good reason for this: If an A/C coil were sized for 10-15F/5.5-8.5K TD it would be incapable of achieving acceptable humidity levels. Your home would be a cold swamp.

You have no such dehumidification needs in this system, therefore you can achieve much higher COP by sizing your coil in accordance with refrigeration practices as opposed to A/C practices.

Can you walk me through this, step-by-step? I'm assuming by refrigeration coils & A/C coils, you're referring to the evaporator coils?

If we are to size for a smaller TD (e.g. 10-15F/5.5-8.5K TD), would it not end up being a lot larger - volume, or area, for the same heat-transfer rate?

This is no problem, in practice, as the fin heat-transfer coefficient can be improved dramatically with a new design concept currently on the drawing board.

desA
28-07-2009, 02:07 PM
On the other hand, conventional sizing calculations assume a portion of the coil is used for flashing off the liquid. Since we are taking steps to eliminate flashing this changes everything.

I'm thinking you are going to have to size and adjust every component through a step-by-step trial and error process.

On the bright side, at the end of this process you may be in a position to devise your own set of unique formulas for this particular industry niche.

Gary, thanks so much for the incredible input & insights you have provided on this thread. Between you & Magoo you have given me a huge amount to think about & experiment with. I really do see that innovations need to be developed in this market niche. The end result will hopefully be passed on in terms of smaller size & lower cost to the end user.

Gary
28-07-2009, 03:56 PM
Ok, that will probably make for an interesting supply discussion.

Which manufacturers of TXV's would be open to supply optimised MOP options? In my view, this would actually be a useful service component in terms of service support. Perhaps also give a bit of competitive edge as well.


This would be a moot point if a cap tube is used. And a cap tube will be much better for this application. The cap tube can be used for both pressure reduction and HX and this will reduce weight, expense and footprint. It is ideal.

The key to using a cap tube for this application is the evap heat load stabilization provided by the fan control.

Gary
28-07-2009, 04:01 PM
Can you walk me through this, step-by-step? I'm assuming by refrigeration coils & A/C coils, you're referring to the evaporator coils?

If we are to size for a smaller TD (e.g. 10-15F/5.5-8.5K TD), would it not end up being a lot larger - volume, or area, for the same heat-transfer rate?


We already have a target TD. At 25C incoming air temp, we want the SST to be 15C, therefore 25-15=10K.

Our ideal TD, assuming a minimum incoming air temp of 25C, is 10K.

If we were designing for cooler ambients, say 20C incoming air, we would want to size our coil for 20-15=5K TD. In trying to get our TD lower than 5K we would hit a point of diminishing returns.

The machine would work well in cooler ambients than we are designing for, but the SST would drop and we would no longer be riding the upper limits of the compressor, thus the COP would be reduced as the ambient temp reduces.

In other words, if we were designing for all ambients, we would want the evap TD to be 5K.

desA
28-07-2009, 04:16 PM
^ What happens when incoming air temp rises to 35'C, as is common in Asia & Africa?

Gary
28-07-2009, 04:36 PM
^ What happens when incoming air temp rises to 35'C, as is common in Asia & Africa?

The fan slows down to compensate for the rise in heat load. Less evap airflow = less evap heat load.

desA
29-07-2009, 02:40 AM
The fan slows down to compensate for the rise in heat load. Less evap airflow = less evap heat load.

Agreed.

This fan control concept is also applicable to the unsteady heat-pump cycle situation.

desA
29-07-2009, 02:42 AM
What are your thoughts on COP,hp optimisation?

How to force the heat-pump system to operate on maximum COP,hp throughout the heating cycle?

desA
29-07-2009, 02:24 PM
Hi desA
The primary consideration is compressor operational conditions, for longevity of system integraty.
Did the TEV superheat check I suggested stabalize performance, it has worked for me for decades.

I did the following today, on the lab machine:
1. Weighed in the calculated refrigerant charge sufficient for all internal components at selected operating conditions;
2. Let the system settle;
3. Tuned the TXV using the 0.6-0.7 times TD rule;
4. Let system settle between TXV adjustments;
5. Ramped system up from ambient water temp to 60'C.

The system ran as sweet as a bird. Smooth. The TXV tuning rule seems to be bang on, Magoo. I'm a very happy camper.

Tomorrow, I plan to run the machine up to my standard test point, hold it & then further fine-tune the TXV a 1/4 turn at most from its current position.

Thanks Magoo - I owe you a few cold ones.

Gary
29-07-2009, 02:51 PM
What are your thoughts on COP,hp optimisation?

How to force the heat-pump system to operate on maximum COP,hp throughout the heating cycle?


Now that we are maximizing the heat absorption on the low side, what remains is to transfer that heat to the water flowing through the condenser.

As heat is transferred from the refrigerant to the water, the refrigerant temp drops, having lost heat, and the leaving water temp rises, having gained heat.

The two temperatures "approach" each other. Thus the difference between the SCT temp and the leaving water temp is called the approach temp.

By installing a water regulating valve at the condenser outlet (which senses and controls SCT) we can regulate the flow to give us a steady SCT of 75C, which is the upper limit of the compressor.

Since the SCT is thereby fixed at 75C, the variable in our approach becomes the water leaving temp. The better the heat transfer, the higher the leaving water temp.

We need to have a condenser large enough to give us 65C leaving water temp, giving us an approach temp of 75-65=10K approach.

By increasing the condenser size beyond this minimum we can increase water temp and reduce the approach, but this again involves a point of diminishing returns.

I'm thinking a reasonable approach target would be about 5K (75-70=5K).

desA
29-07-2009, 03:09 PM
By install a water regulating valve at the condenser outlet (which senses and controls SCT) we can regulate the flow to give us a steady SCT of 75C, which is the upper limit of the compressor.

Since the SCT is thereby fixed at 75C, the variable in our approach becomes the water leaving temp. The better the heat transfer, the higher the leaving water temp.

This is excellent system design logic. Thanks for this.

I'd envisage a simple water flow control valve, taking its signal off the condenser. For a concentric-tube type, this should be no contest as long as the temp probe is well-secured.

For a plate-type condenser, I wonder whether there would be a decent place to pick up the condensing temp SCT @ 75'C on the body? May have to infer the SCT from another suitable temperature, or perhaps convert the SCP pressure, in the controller.

My current design approach is typically around 10K, for an economical design. The cost effect can be easily developed versus approach - I'll look into that.

Gary
29-07-2009, 03:29 PM
The only problem I can foresee at this point is in maintaining the water temp in the storage vessel.

Since the condenser incoming water temp is high, we may not be able to get enough flow through our condenser to prevent the SCT from exceeding the 75C limit.

If this is the case we may need to override the fan control, shutting down the fan to decrease the heat transfer... but let's cross that bridge if/when we come to it.

Gary
29-07-2009, 03:33 PM
I'd envisage a simple water flow control valve, taking its signal off the condenser. For a concentric-tube type, this should be no contest as long as the temp probe is well-secured.


I would control the flow off high side pressure, rather than temperature. This is a common strategy for water cooled condensers and the valve is readily available. Controlling off pressure rather than temp assures that we are riding the upper limit of the compressor but not exceeding it.

desA
29-07-2009, 03:45 PM
The storage vessels are usually fairly large, for the larger heat-pumps. Some could be in the region of 4000L or so. There would be a fair amount of thermal lag there.

A sensor signal from the tank can be input into the heat-pump controller & action taken to shut back the pump, fan etc. Typically the water temp is used to shut the heat-pump off, when the pre-set temp has been reached.

Hi/low pressure trips are also installed. If the water flow drops too low & SCT rises above the critical value, the hi/low should catch it, if correctly set - to protect the compresor.

Basically, since the hot water is generally blended with cold water by the end user, the terminal tank temperature can undershoot, or even overshoot a little, without much problem - in general.

desA
29-07-2009, 03:47 PM
I would control the flow off high side pressure, rather than temperature. This is a common strategy for water cooled condensers and the valve is readily available. Controlling off pressure rather than temp assures that we are riding the upper limit of the compressor but not exceeding it.

Ok, good. That makes good sense.

Do you have any manufacturers, or model numbers that may be of use in this application? I'll research it.

Gary
29-07-2009, 04:00 PM
This is what it looks like:

http://www.grainger.com/itemdetail_largepic.html?item=XL-4LZ96.JPG

Gary
29-07-2009, 04:07 PM
Basically, since the hot water is generally blended with cold water by the end user, the terminal tank temperature can undershoot, or even overshoot a little, without much problem - in general.

Our strategy being to fully heat the feed water before sending it to the storage vessel, we would not want the hot water to be blended with cold water in the storage vessel.

desA
29-07-2009, 04:24 PM
^^ Thanks for the pic.

That would be fairly easy to implement. Install valve on water ?exit? line - install bulb in discharge line prior to condenser, or at entry.

desA
29-07-2009, 04:28 PM
Our strategy being to fully heat the feed water before sending it to the storage vessel, we would not want the hot water to be blended with cold water in the storage vessel.

What I was aiming at is that for many hot water applications - at least for hotels, is that the hot water from the storage vessel (sat 60-65'C) is piped into a blender/mixer head in the shower, where it is blended with cold water.

So, the storage tank temperature could actually be acceptable at an upper temp of 63-65'C say, with little noticeable difference to the person taking a shower.

Gary
29-07-2009, 04:31 PM
^^ Thanks for the pic.

That would be fairly easy to implement. Install valve on water ?exit? line - install bulb in discharge line prior to condenser, or at entry.

Yes, the valve would be installed at the water exit.

On the end of that cap tube in the picture there is a 1/4 inch flare nut, which would be connected to a standard 1/4 inch access fitting, preferably in the discharge line prior to the condenser. It might be a good idea to also install a small shutoff valve between the access fitting and the flare nut, in case the valve may someday need to be replaced.

desA
29-07-2009, 04:49 PM
^ Thanks, Gary. I'm very happy with using the water flow-control concept via a capillary tube. I loathe too much reliance on electronics, if at all possible.

Next on the list:

At startup temperatures, with a Tc,sat temperature of around 35'C, say, the required refrigerant mass charge could be around 1200g, for instance. At Tc,sat of around 70'C, the required mass charge would only be 1020g - per calculation.

The required mass charges for the cold & hot condition are different. How to set up a suitable refrigerant loop such that the charge difference doesn't end up swamping the condenser in the hot condition?

The mass charge in the evaporator seems to decrease as the cycle moves upwards towards Tc,sat=70/75'C, since the quality moves from around x=0.12 to around 0.4. The excess refrigerant then needs to move to the condenser, where it holds up.

Gary
29-07-2009, 05:02 PM
The refrigerant charge should be suitable for the hot condition, not the cold condition... and given the water regulating valve that hot condition would be very quickly reached.

Gary
29-07-2009, 06:28 PM
The refrigerant charge should be suitable for the hot condition, not the cold condition... and given the water regulating valve that hot condition would be very quickly reached.

Perhaps this answer is too simplistic. To elaborate:

A TXV system requires surplus refrigerant in order to respond to variations in heat load. That surplus is stored in a receiver, or takes up valuable space in the condenser if there is no receiver. In addition there is excessive liquid pressure at the TXV due to the desired high side temperature, requiring a PRV to reduce the liquid pressure.

In adding the fan control we eliminate the heat load variations and stabilize the low side pressure... and in adding the water regulating valve we stabilize the high side pressure. Thus we ride the upper limits of the compressor (Te,sat 15C, Tc,sat 75C) throughout the cycle.

Having stabilized the heat load as well as both low and high side pressures, the system is now ideal for a cap tube as the metering device, thus eliminating the TXV, PRV, receiver... and the surplus refrigerant.

desA
30-07-2009, 01:06 AM
The required mass charges for the cold & hot condition are different. How to set up a suitable refrigerant loop such that the charge difference doesn't end up swamping the condenser in the hot condition?

A few thoughts on this - let's say for a typical, unoptimized, low-cost circuit (open to critical review):
1. Set system charge at start-up temp mass requirements - to optimize start-up heat-performance;
2. Oversize condenser suitably such that its internal storage provides sufficient space to store the excess refrigerant for the high-temp end point.

At this point, the condenser acts as a receiver.

The refrigerant excess at hot condition is around (1200-1020)/1020*100 = +17.6%. For most condensers, the design oversurface is in excess of this value & so the condenser provides a natural mass storage receiver, at little additional on-cost.

Gary
30-07-2009, 01:16 AM
A few thoughts on this - let's say for a typical, unoptimized, low-cost circuit (open to critical review):
1. Set system charge at start-up temp mass requirements - to optimize start-up heat-performance;
2. Oversize condenser suitably such that its internal storage provides sufficient space to store the excess refrigerant for the high-temp end point.

At this point, the condenser acts as a receiver.

The refrigerant excess at hot condition is around (1200-1020)/1020*100 = +17.6%. For most condensers, the design oversurface is in excess of this value & so the condenser provides a natural mass storage receiver, at little additional on-cost.

You forgot to add:

3. TXV

Gary
30-07-2009, 01:31 AM
I'm not sure if there was a question in your last post.

TXV's excell at handling variable loads... and for this they need surplus refrigerant.

desA
30-07-2009, 01:36 AM
^ I'd imagine that a TXV would have to be in that kind of circuit, due to the mass charge migration issues.

What I was thinking through is the case of some of the typical lowish-cost systems I've come across, where no liquid receiver, or suction accumulators are present in the system.

In academic literature, the mass migration effect is known & some have even talked about the idea of storing the excess refrigerant outside the main circuit, until required. All kinds of ideas. The idea of using the heat-exchanger as a storage device probably allows this to be done, at low cost - although the TXV comes at a price.

--------
Cap-tube system (open to review):
So, for a cap-tube system, the main idea would then be to rather size the refrigerant mass charge for hot-load condition (not start-up) & run a little low on start-up heat-performance, knowing that most of the run time will be spent at the hot condition anyway.

desA
30-07-2009, 01:54 AM
I'm not sure if there was a question in your last post.

More of thinking out aloud. :o



TXV's excell at handling variable loads... and for this they need surplus refrigerant.

Fair-enough.

To extend the condenser storage idea & TXV a little further, on such systems. Would it be feasible to further over-size the condenser so that it sub-cools beyond the typical ~8K amount?

For example, to try & force a larger sub-cooling in the condenser through a combination of design over-surface & liquid-line design.

Gary
30-07-2009, 03:13 AM
More of thinking out aloud. :o



Fair-enough.

To extend the condenser storage idea & TXV a little further, on such systems. Would it be feasible to further over-size the condenser so that it sub-cools beyond the typical ~8K amount?

For example, to try & force a larger sub-cooling in the condenser through a combination of design over-surface & liquid-line design.

How does this benefit the system?

desA
30-07-2009, 03:45 AM
The basic idea is as follows:

Increasing the sub-cooling in the condenser will allow additional heat to the water stream. This will increase COP,hp slightly for very little on-cost.

Essentially, this is integrating the sub-cooler into the condenser through designed over-surface. As long as the amount of over-surface doesn't wreak havoc with the de-superheating & condensing operation, then it should, in principle, be possible.

I'm not sure just how far this can be pushed - I'd estimate that the 20-30% range would be within normal design limits. For ultra-compact condenser designs, the increase in condenser volume is marginal, although for concentric tube condensers, the impact could be much larger.

Magoo
30-07-2009, 04:23 AM
I am getting a tad confused here, generall parallel flow has a heat transfer lose, but are you using that to maintain higher condensing/ SDT. I would tend to go counter flow with water and control water outlet temp., with a variable water flow to acheive same result.
Years ago I was involved with the developemnt of a HW heat pump as at the experimental stage on R12 [ OK years ago ]. tube and tube counter flow. Had to be tube soldered to tube, due to health regs of double wall separation. Fractional HP compressor as a pre-heater for general electric hot water heating system [ domestic ]. ran like a dream cut hot water heating costs by around 50%.
Carrier Inc.. also then marketed a system called a "hot shot "as and add on to existing air con system. Acted as a de-superheater on discharge.
Both systems probably way ahead of there time as and energy saver. Now everyone wants to save costs. Strange how things turn out.

Gary
30-07-2009, 04:38 AM
The basic idea is as follows:

Increasing the sub-cooling in the condenser will allow additional heat to the water stream. This will increase COP,hp slightly for very little on-cost.


The cooling would be limited to the temp of the entering water. I'm thinking there would be little if any gain and in fact I would bet there would be a loss as compared to using that same oversized condenser without the excess subcooling... but I could be wrong.

desA
30-07-2009, 04:51 AM
^^ For heat-pumps, the condenser is typically either a tube-in-tube, or plate. The piping is usually run in a counterflow direction.

Sidebar:
Practically though, if the condensing part carries most of the load, then even a parallel flow condition (via water connections), will not make a huge difference - theorectically, at least. Reason is that the de-superheating & sub-cooling parts of the condenser typically take up around 15-20% of the heat-load. If all is well, then the volume used by de-superheating & sub-cooling will be close to the heat-load fraction, although this is not actually always the case (needs to be checked carefully).

------------
With a need to push up the sub-cooling portion of the condenser, it will occupy additional space, which will need to be carefully managed. Normally, liquid-liquid (sub-cooler) heat-exchange does not require much surface area, compared to say vapour-liquid (de-superheater), so the additional space required is not that large, in practice.

Some tricks may have to be played on the liquid-line pipework to hold back the liquid in the condenser a tad, to force the volume retention, for instance.

desA
30-07-2009, 04:55 AM
The cooling would be limited to the temp of the entering water.

This is very true - it is. The entry temperature is critical, in how far the sub-cooling can be pushed.



I'm thinking there would be little if any gain and in fact I would bet there would be a loss as compared to using that same oversized condenser without the excess subcooling... but I could be wrong.

Where would the loss come from, though?

desA
30-07-2009, 04:57 AM
I've put up a link to the TXV rule Magoo mentioned earlier. (I hope this is ok).

http://pdf.directindustry.com/pdf/gea-kuba-gmbh/evaporator-sg-english-french-german/17657-57696-_27.html

Gary
30-07-2009, 05:06 AM
Where would the loss come from, though?

I'm thinking there would be more gain from using the extra condenser area to lower the approach temp, therefore a loss by comparison.

There is very little heat transfer involved in subcooling a liquid.

desA
30-07-2009, 05:18 AM
^ Fair-enough - good thought.

What I'll do is to have a set of ratings done by my condenser supplier, for various combinations of water entry temp & Tc,sat. From that data we should have a more clear idea of the system trade-offs.

I've already done this for a fixed Tc,sat & variable water inlet temp, but will extend the study to see where the sensitivities lie. Good one...

------
During operation with a pump-around system, the water flow-rate setting will determine the number of water passes (times) through the condenser, & hence the dTw across the condenser.

If the condenser is sized to cope with the hot-condition such that the condensing area is sufficient for this, any additional over-surface can then be put to good use, without disrupting the condensing section.

Gary
30-07-2009, 04:12 PM
I've put up a link to the TXV rule Magoo mentioned earlier. (I hope this is ok).

http://pdf.directindustry.com/pdf/gea-kuba-gmbh/evaporator-sg-english-french-german/17657-57696-_27.html

I haven't tested the Magoo Method (superheat 60-70% of TD), but it makes good sense. It should work very well over a wide range of conditions.

desA
30-07-2009, 04:32 PM
Experimental feedback - AWHP test system - 30.07.09



Evaporator:
Te,sat = 13'C (sat temp)
Te,sup = 20.8'C (vapour exit temp)
Ta,in = 24.9'C (air inlet temp)
va,in = 3.5 m/s (air inlet face velocity to evap face)
Ta,out = 21.5'C (air outlet temp)
va,out = 9.5 m/s (air outlet velocity - fan discharge)

Condenser:
Tc,sup = 64.3'C (superheated vapour inlet to condenser)
Tc,sat = 50'C (condenser sat temp)
Tc,sc = 43.7'C (liquid exit temp)
Tw,out = 39.8'C (water outlet temp)

Other:
Tw,tank = 38.9'C (water storage tank temp - mixed, half-height)
Tcomp,base = 41.1'C (compressor base temp)

Electrical:
Current = 6.1 A
Voltage = 223V (single phase, ~50Hz)


This run was performed at a mass charge calculated from the internal volume at operating conditions - internal software. The cold start-up mass was used in this trial.

I'm very interested in hearing your comments on this unit.

The particular test unit is undersized on the condenser, in my view. It is an early test machine, using a tube-in-tube condenser, TXV & scroll compressor. Typically, under a dynamic heat-ramp test, the heat-up power to raise the storage tank water temp from ambient to hot temp temp is slightly lower than the compressor performance tables would predict. I have always suspected that this stemmed from a slightly under-sized condenser. The low sub-cooling value seems to bear this out.

Gary
30-07-2009, 06:16 PM
Evaporator:
Te,sat = 13'C (sat temp)
Te,sup = 20.8'C (vapour exit temp)
Ta,in = 24.9'C (air inlet temp)
va,in = 3.5 m/s (air inlet face velocity to evap face)
Ta,out = 21.5'C (air outlet temp)
va,out = 9.5 m/s (air outlet velocity - fan discharge)

Condenser:
Tc,sup = 64.3'C (superheated vapour inlet to condenser)
Tc,sat = 50'C (condenser sat temp)
Tc,sc = 43.7'C (liquid exit temp)
Tw,out = 39.8'C (water outlet temp)

Other:
Tw,tank = 38.9'C (water storage tank temp - mixed, half-height)
Tcomp,base = 41.1'C (compressor base temp)

Electrical:
Current = 6.1 A
Voltage = 223V (single phase, ~50Hz)

Evaporator:

dT = 24.9-21.5 = 3.4K/6.1F
TD = 24.9-13 = 11.9K/21.4F
SH = 20.8-13 = 7.8K/14F
Appr = 21.5-13 = 8.5K/15.3F

Condenser:

dT = 39.8-38.9 = 0.9K/1.6F
TD = 50-38.9 = 11.1K/20F
SC = 50-43.7 = 6.3K/11.3F
Appr = 50-39.8 = 10.2K/18.4F

I am assuming that the Tw,tank is the same as the cond entering water temp, although this may not be the case.

The subcooling at 6.3K is well within what I would consider to be a normal range (5.5-8.5K). This is largely a matter of refrigerant charge.

As to undersize/oversize, these are terms that are relative to accepted standards. Are there in fact any accepted standards for this type of system?

All seems to be working well at this point. It will be interesting to see how these numbers compare with (all else being equal) a much hotter incoming water temp, as will be seen at the end of the heating cycle.

An additional temp which should be monitored/recorded would be the discharge line temp near the compressor.

desA
30-07-2009, 11:59 PM
Thanks so much for your comments.


I am assuming that the Tw,tank is the same as the cond entering water temp, although this may not be the case.

This is a close-enough approximation & within the range of experimental measurement uncertainty.



The subcooling at 6.3K is well within what I would consider to be a normal range (5.5-8.5K). This is largely a matter of refrigerant charge.

The current test point (Tc,sat=50'C) is around mid-range for the current designs. The amount of sub-cooling increases as the heating cycle advances. Towards the point Tc,sat~70-75'C, the sub-cooling effect is far more pronounced - in excess of 10-12K. This is what makes me think that the condenser is under-size.



As to undersize/oversize, these are terms that are relative to accepted standards. Are there in fact any accepted standards for this type of system?

To be honest, the standards that seem to be out there, for this kind of machine, seem to be locally-applied, rather than an industry-wide standard. Most of it seems to be still manufacturer-specific, as far as I can determine.



All seems to be working well at this point. It will be interesting to see how these numbers compare with (all else being equal) a much hotter incoming water temp, as will be seen at the end of the heating cycle.

Ok - good. What I'll do next is to do a set of tests at fixed points across the whole heating cycle & feed back. This will allow a more clear picture of the overall performance.



An additional temp which should be monitored/recorded would be the discharge line temp near the compressor.

The Tc,sup = 64.3'C (superheated vapour inlet to condenser) should be re-named as 'superheated vapour exit from compressor', as this is the point where it is physically measured. I'm used to using the figure for condenser performance checks & so lumped it there. Apologies.

Gary
31-07-2009, 12:19 AM
Ok - good. What I'll do next is to do a set of tests at fixed points across the whole heating cycle & feed back. This will allow a more clear picture of the overall performance.


Excellent. Then we can see the changes throughout the system that occur due to the incoming water temperature rising.

There should also be a similar set of tests (start of cycle/end of cycle) with the incoming air at 35C, as this is the other design extreme.

desA
31-07-2009, 12:45 AM
There should also be a similar set of tests (start of cycle/end of cycle) with the incoming air at 35C, as this is the other design extreme.

May have to wait a while for that - we're in the rainy season at present. Probably get up to ~30'C in the next few days, with lots of humidity. (Fill up water buckets from the evap drain - lol)

Gary
31-07-2009, 12:55 AM
May have to wait a while for that - we're in the rainy season at present. Probably get up to ~30'C in the next few days, with lots of humidity. (Fill up water buckets from the evap drain - lol)

I recall the weather extremes in that area of the world. When I was over there, people were shooting at me, so that made it even more uncomfortable... lol

Gary
31-07-2009, 02:16 AM
You really should have some means of regulating the intake air temp to simulate the two extremes that you are designing for.

desA
31-07-2009, 04:43 AM
That's true.

I'm presently planning to relocate from SEA back to my homeland. Once settled, I plan to set up the lab with some level of environmental control - it has to be done. To add hot air is not a problem, it's more when you want to cool the air to just above 0'C, that it gets a little more costly.

Gary
31-07-2009, 05:14 AM
That's true.

I'm presently planning to relocate from SEA back to my homeland. Once settled, I plan to set up the lab with some level of environmental control - it has to be done. To add hot air is not a problem, it's more when you want to cool the air to just above 0'C, that it gets a little more costly.

What you need is a good heat pump. ;)

desA
31-07-2009, 05:47 AM
^ Hahaha... got 4 of those already. Point taken... :D

desA
31-07-2009, 05:57 AM
A few thoughts on using only AWHP's to create a hot/cold test capability.

Air cooling
Using a air-source AWHP, where the output air-stream is partially re-circulated back to the evap inlet would bring the air temp down. I wonder how far this could go, though - if ducted properly?

Air heating
Use the hot water storage to pre-heat the incoming air-stream, or use a bank of direct-element heaters to raise the incoming air stream. Ducted flow.

Any thoughts?

Gary
31-07-2009, 06:41 AM
A few thoughts on using only AWHP's to create a hot/cold test capability.

Air cooling
Using a air-source AWHP, where the output air-stream is partially re-circulated back to the evap inlet would bring the air temp down. I wonder how far this could go, though - if ducted properly?

Aside from the obvious coil freezing problems:

As the temp drops the heat load drops, so eventually the TXV would start hunting and then uncontrolled flooding.

The orifice could be downsized, but the smaller the orifice the slower the pulldown.

The temp could theoretically keep dropping until the compressor reached its lower limit, where it just can't pull a deeper vacuum or the liquid temp coming back from the condenser engages the entire coil in flashing.

If you have a target temp you want to stop at you could do this with an evaporator pressure regulator (EPR) valve.


Air heating
Use the hot water storage to pre-heat the incoming air-stream, or use a bank of direct-element heaters to raise the incoming air stream. Ducted flow.


A hot water coil could be used easily enough.

Gary
31-07-2009, 06:57 AM
Actually, nothing so exotic is needed. I'm thinking a cheap window A/C could be modified to provide the heating and/or cooling.

desA
31-07-2009, 08:32 AM
Good idea. A window-rattler.

Isn't this the beauty of being in the HVAC&R game... lol... :)

desA
31-07-2009, 04:21 PM
The performance tests from Tc,sat = 40'C through 75'C, in steps of 5'C, are done. I need to tabulate & graph the results. Hope to have these out this weekend.

At Tc,sat=40'C, the condenser sub-cooling is 3.8K & goes up to 15.2K at Tc,sat=75'C !!! :eek:

Gary
31-07-2009, 04:28 PM
The performance tests from Tc,sat = 40'C through 75'C, in steps of 5'C, are done. I need to tabulate & graph the results. Hope to have these out this weekend.

At Tc,sat=40'C, the condenser sub-cooling is 3.8K & goes up to 15.2K at Tc,sat=75'C !!! :eek:

Now you know what receivers are for.

Gary
31-07-2009, 04:38 PM
What was the leaving water temp at Tc,sat=75C?

desA
01-08-2009, 02:21 AM
What was the leaving water temp at Tc,sat=75C?

Tw,out = 60.9'C @ Tc,sat=75'C

Gary
01-08-2009, 02:49 AM
At Tc,sat=40'C, the condenser sub-cooling is 3.8K & goes up to 15.2K at Tc,sat=75'C !!! :eek:


Tw,out = 60.9'C @ Tc,sat=75'C

Hmmm... 15.2K subcooling with 14.1K approach.

This would seem like an opportune moment to test the excess subcooling theory. You could remove refrigerant until the subcooling is 8.5K and see if the approach increases or decreases.

desA
01-08-2009, 03:23 AM
^ Can you explain a little further on this, please?

Gary
01-08-2009, 03:39 AM
You were speculating earlier about using part of the condenser for excess subcooling. You now have excess subcooling backing up into the condenser.

If that excess subcooling is beneficial, then removing it would increase the approach temp at Tc,sat=75C.

If that excess subcooling is not beneficial, then removing it would decrease the approach temp at Tc,sat=75C.

desA
01-08-2009, 03:59 AM
Ok, fair-enough.

Let me finish the full test table & put it up (wip). We can then see the progression the whole way through the heat-up cycle. I also have the dynamic ramp test from the previous day, to measure overall (average) heating rate under that mass condition.

We can then make a solid decision on how much charge to remove.

I value your input greatly - it is an extremely valuable exercise for me.

Gary
01-08-2009, 05:05 AM
Sorry, I don't mean to rush you. You will of course, need a full spectrum of test tables for a base, in order to quantify the value of improvements.

desA
01-08-2009, 05:41 AM
http://i30.tinypic.com/34dn8gi.png

http://tinypic.com/r/34dn8gi/3

That's the recent raw data.

desA
01-08-2009, 06:11 AM
http://i27.tinypic.com/21c7ame.png

http://tinypic.com/r/21c7ame/3

Correction... @ Tc,sat=75'C swapped Tc,exit & Tw,out data.

desA
01-08-2009, 06:23 AM
We can be as critical of this design as we like. :)

It is a machine I bartered in trade for consulting work performed for a now defunct heat-pump builder. It is not my design, I've just fine-tuned its performance a little & tested its robustness under 3rd world conditions. I use it as my test basis on which to benchmark my next machine, which is almost complete.

desA
01-08-2009, 07:49 AM
Note that the data speaks with a forked tongue... :)

You will observe that the Tw,out & T,tank values cross at some point. This should be non-physical & will definitely affect the condenser Approach calculation.

It is something I've observed in earlier trials at one of my clients. In my view the temp probe for the water outlet from the condenser, although insulated, is losing accuracy as temp increases, whereas the storage water temp is mixed & therefore more representative.

I will upload a plot showing both the conventional (Approach=Tc,sat-Tw,out) & modified (Approach'=Tc,sat-T,tank), for clarity. It makes a huge difference in interpretation.

desA
01-08-2009, 07:54 AM
http://i28.tinypic.com/2hi3r6o.png

http://tinypic.com/r/2hi3r6o/3

The value (-Cross)=Tc,exit-Tw,out. This is a critical parameter for heat-exchanger design purposes.

(-Cross)'=Tc,exit-T,tank

desA
01-08-2009, 07:56 AM
Hi Nike123, I see you lurking... welcome to comment... :)

desA
01-08-2009, 10:57 AM
Update - water calibration curve:

What I have done is to add an additional thermocouple probe onto the condenser water outlet pipe. I will calibrate the two Tw,out probes against the T,tank probe tomorrow & adjust the experimental results accordingly.

Practically, though, I'd expect the useful data for the Approach to be for (Approach) to the left of the temp crossing point & for (Approach)' to the right of the temp crossing point. Let's see what skewing the calibration curves bring out.

Gary
01-08-2009, 02:51 PM
We have been assuming that T,tank is the same as Tw,in.

It is not possible for Tw,in (T,tank) to be higher than Tw,out.

Nor is it possible for Tw,in (T,tank) to be higher than Tc,exit.

I tend to believe the problem is T,tank.

To my thinking, it would be far preferable to measure Tw,in and assume that it is the same as T,tank than the reverse.

Gary
01-08-2009, 03:51 PM
The TXV is working very well despite the increase in liquid pressure. That's very good news.

At the end of cycle the compressor discharge temp is running dangerously high (103.9C). Since we are riding the upper limits of the compressor, we need to provide maximum cooling for it. The compressor is cooled by the superheated vapor. Therefore we need to run the lower limits of acceptable superheat. Going by Magoos rule, at end of cycle the TD is 25.2-14=11.2TD.

11.2*0.6=6.72SH.

The TXV superheat could be lowered to about 6.72K from its current 7.9K. This would lower the compressor discharge temp.

desA
01-08-2009, 04:05 PM
^Good point. Thanks for that. I'll drop that down a tad to suit.

What maximum discharge temp would you be comfortable with off a scroll compressor?

The Copeland document AE-1263-R3 states that discharge temp measured 6" from the compressor discharge should be max ~ 250'F (121.1'C), with internal compressor exit maxima no higher than 300'F (148.89'C).

On that basis, we still have a little spare room.

The other limits along the way are as follows:
1. Motor winding limit : 325'F (162.78'C)
2. Oil breakdown limit : 325'F (162.78'C)
3. Compressor oil film limit : 310-325'F (154.4-162.78'C)
4. Sump temperature limit : 200'F (93.3'C)

desA
01-08-2009, 04:11 PM
We have been assuming that T,tank is the same as Tw,in.

True... thanks for the reminder - was getting lost in there for a second.



It is not possible for Tw,in (T,tank) to be higher than Tw,out.

Nor is it possible for Tw,in (T,tank) to be higher than Tc,exit.

I tend to believe the problem is T,tank.

To my thinking, it would be far preferable to measure Tw,in and assume that it is the same as T,tank than the reverse.

Point taken.

The problem with these type K temp probes is that the accepted experimental uncertainty is +- 2.2'C. Add to that positioning inaccuracy, tank mixing & so forth - & we have a quagmire.

I plan to run up a calibration trial on all 3 probes tomorrow. May have to start with ice & go on up to boiling to try & find out the bias of each... headaches.

Gary
01-08-2009, 04:20 PM
^Good point. Thanks for that. I'll drop that down a tad to suit.

What maximum discharge temp would you be comfortable with off a scroll compressor?

Being from the old school, I was taught that if you spit on the discharge line and it boils you have a problem. But it seems that compressors are more tolerant these days.

Still, I like to keep the discharge a little cooler to ensure the compressor lives a long and happy life.

And a lower superheat picks up more heat in the evap, thus increasing COP.

desA
01-08-2009, 04:35 PM
^ Points well taken.

I'll set the superheat down in the morning. That's sound advice.

The scroll compressor concept seems to be doing fairly well, it seems. Copeland seem to be continually revising their operational envelopes, based on field feedback.

desA
01-08-2009, 04:42 PM
http://i26.tinypic.com/250ou8n.png

http://tinypic.com/r/250ou8n/3

What would be your comments on this chart?

(-Cross)=Tc,exit-Tw,out

From my heat-exchanger design experience, I would say that the condenser performance is beginning to drop off drastically after Tc,sat = 50'C, based on (-Cross), although the Approach drop-off seems to occur around Tc,sat = 55'C. The SC shows a rapid rise after Tc,sat = 55'C as well.

Gary
01-08-2009, 05:04 PM
I would attribute the dropoff in performance to the excessive subcooling. Liquid expands when heated. The liquid has no receiver to expand into, therefore it backs up into the condenser, taking up valuable space.

Gary
01-08-2009, 06:43 PM
At Tc,sat 75C, we can remove refrigerant to reduce the subcooling. If the superheat does not rise as a result, then the system has sufficient liquid refrigerant to feed the evap.

If reducing subcooling results in high superheat then the system needs a receiver.

desA
02-08-2009, 02:59 AM
A few thoughts on condenser size & liquid back-up & sub-cooling in the longer term.

The current condenser in this AWHP is a set of two tube-in-tube condenser coils placed in parallel. What if the condenser surface area were to be enlarged by adding a third coil in parallel?

My logic here is that is should be better to err on over-surface for the condenser & so maximise the heat-output from the system. According to the compressor performance charts, the system is capable of pushing around 10-15% more performance than the current heat-pump can deliver.

(We can discuss system balances evap/compressor/condenser a little further down the track)

desA
02-08-2009, 03:14 AM
I would attribute the dropoff in performance to the excessive subcooling. Liquid expands when heated. The liquid has no receiver to expand into, therefore it backs up into the condenser, taking up valuable space.

A few observations on these heat-pump dynamics as the heating cycle advances:

1. At startup condition, the evaporator has inlet quality of x=0.12 kg/kg.
2. At hot condition, the evaporator has inlet quality of 0.40 kg/kg.
3. As the heating cycle advances, & the evaporator inlet quality rises, the refrigerant mass charge begins to migrate from the evaporator towards the condenser - purely to attempt to maintain thermodynamic & mass balance equilibrium.
4. This mass migration then causes the condenser to become 'flooded' as it fills up & acts as a receiver.
5. If the level of flooding (as evidenced by the amount of sub-cooling) becomes excessive, then the sub-cooling area begins to encroach on the area required for condensation & the overall condenser performance begins to drop off.
6. If the condenser were to be enlarged sufficiently to cope with the flooding effect i.e. act as a receiver, then sufficient area would still remain for the condensing phase to occur unimpeded.

Interestingly, the mass migration effects can be observed during a heating cycle, by observing the compressor amperage behaviour. At times, the compressor seems to take on more load & can be heard to 'dig in' & pull slightly harder. During this time, small swings in condenser Tc,exit temp can be seen. Once this event settles, then the system stabilises at the new equilibrium point, settles down & all is smooth again. These events occur throughout the cycle & can be seen on the amperage-time trace as small waves.

This is a consequence of real-time system dynamics (unsteady) imposed on a thermodynamic system that presupposes & is designed for, a 'steady' operating regime.

This is why I prefer a thermal bulb to drive the TXV in such situations & not use an electronic variant. Thermal bulbs have intrinsic thermal inertia & a slow response - this helps to smooth the refrigerant waves in the system. If the wrong controller (fast response) were to be used, some very interesting wave dynamics could result. Sometimes, slower is better - in my view, at least.

Gary
02-08-2009, 05:16 AM
A few thoughts on condenser size & liquid back-up & sub-cooling in the longer term.

The current condenser in this AWHP is a set of two tube-in-tube condenser coils placed in parallel. What if the condenser surface area were to be enlarged by adding a third coil in parallel?

My logic here is that is should be better to err on over-surface for the condenser & so maximise the heat-output from the system. According to the compressor performance charts, the system is capable of pushing around 10-15% more performance than the current heat-pump can deliver.


No doubt... bigger is better.

But we haven't seen what it will do with 35C incoming air yet.

Gary
02-08-2009, 05:40 AM
A few observations on these heat-pump dynamics as the heating cycle advances:

1. At startup condition, the evaporator has inlet quality of x=0.12 kg/kg.
2. At hot condition, the evaporator has inlet quality of 0.40 kg/kg.
3. As the heating cycle advances, & the evaporator inlet quality rises, the refrigerant mass charge begins to migrate from the evaporator towards the condenser - purely to attempt to maintain thermodynamic & mass balance equilibrium.
4. This mass migration then causes the condenser to become 'flooded' as it fills up & acts as a receiver.
5. If the level of flooding (as evidenced by the amount of sub-cooling) becomes excessive, then the sub-cooling area begins to encroach on the area required for condensation & the overall condenser performance begins to drop off.
6. If the condenser were to be enlarged sufficiently to cope with the flooding effect i.e. act as a receiver, then sufficient area would still remain for the condensing phase to occur unimpeded.

There is no "area needed for condensation" as such. The vapor fills the available area and the more transfer surface, the more transfer. However the heat transfer per unit of surface diminishes, so there must necessarily be a point where the excess subcooling does less harm than good.

Still, I am not convinced that the added subcooling is needed. The optimum subcooling seems to be about 7K@75C for this system. If, at 7K@75C subcooling, there is sufficient refrigerant flow for the evaporator over the entire cycle, then the excess subcooling is not needed.

It also occurs to me that if the water were regulated such that the Tc,sat were held at 75C, then the subcooling could be stabilized at 7K, without infringement on the condenser.

Yet another point for the heat-it-in-one-pass school of thought.


Interestingly, the mass migration effects can be observed during a heating cycle, by observing the compressor amperage behaviour. At times, the compressor seems to take on more load & can be heard to 'dig in' & pull slightly harder. During this time, small swings in condenser Tc,exit temp can be seen. Once this event settles, then the system stabilises at the new equilibrium point, settles down & all is smooth again. These events occur throughout the cycle & can be seen on the amperage-time trace as small waves.

This is a consequence of real-time system dynamics (unsteady) imposed on a thermodynamic system that presupposes & is designed for, a 'steady' operating regime.

This is why I prefer a thermal bulb to drive the TXV in such situations & not use an electronic variant. Thermal bulbs have intrinsic thermal inertia & a slow response - this helps to smooth the refrigerant waves in the system. If the wrong controller (fast response) were to be used, some very interesting wave dynamics could result. Sometimes, slower is better - in my view, at least.

Not too fast. Not too slow. Not too much. Not too little. Not too big. Not too small. Balance is everything.

desA
02-08-2009, 06:02 AM
No doubt... bigger is better.

But we haven't seen what it will do with 35C incoming air yet.

This is true. I'll have to work on that in the lab, or wait for monsoon season to stop... lol :D

Supposedly, the performance is to go up with increasing incoming air temp. My concern for this particular AWHP is that the evap is specified at maximum refrigeration capacity of around 13.5 kW, with an air entry face velocity of 2.032m/s (I kid not). The current measured air face velocities are in the range of 3-3.3 m/s.

The system thermodynamic balance calls for the following, at say Te,sat=12.5'C:
1. At Tc,sat=40'C :
Q'evap = 6.7kW
Q'cond = 7.3kW

2. At Tc,sat=50'C :
Q'evap = 6.1kW
Q'cond = 6.9kW

3. At Tc,sat=75'C :
Q'evap = 4.2kW
Q'cond = 6.1kW

What would be the effects of an oversize evaporator on this circuit?

Gary
02-08-2009, 06:24 AM
The compressor is currently nearing its 15C limit with 25C incoming air. It may in fact exceed this limit with 35C incoming air. With a larger evap it would most certainly exceed it.

Gary
02-08-2009, 06:27 AM
The system thermodynamic balance calls for the following, at say Te,sat=12.5'C:
1. At Tc,sat=40'C :
Q'evap = 6.7kW
Q'cond = 7.3kW

2. At Tc,sat=50'C :
Q'evap = 6.1kW
Q'cond = 6.9kW

3. At Tc,sat=75'C :
Q'evap = 4.2kW
Q'cond = 6.1kW


It's easier to make the water warm than it is to make the water hot.

desA
02-08-2009, 06:32 AM
But we haven't seen what it will do with 35C incoming air yet.

For the latest AWHP build (own design), I am hoping to be able to test in my builder's environmental test chamber. This is rigged up to be able to manipulate air temps from Asian to European conditions.

I'm trusting that we can go up to a steady inlet of 35'C & test what goes on with the system.

desA
02-08-2009, 06:38 AM
The compressor is currently nearing its 15C limit with 25C incoming air. It may in fact exceed this limit with 35C incoming air. With a larger evap it would most certainly exceed it.

Agreed. This is my contention with 'one evap fits range' philosophy.

For the new machines, the evap is critically sized & not more. I'm aiming for better air speed control as well.

desA
02-08-2009, 06:39 AM
Hi Chef, please feel free to contribute... :)

Gary
02-08-2009, 06:41 AM
How are we doing with the compressor current draw? Still within the specs?

desA
02-08-2009, 07:07 AM
How are we doing with the compressor current draw? Still within the specs?

Good point. Bingo!!! :eek:

According to the compressor specs at Te,sat=12.5'C:

1. At Tc,sat=40'C : Current@230V = 5.83A - measured 4.9A @ 210V;
2. At Tc,sat=50'C : Current@230V = 6.86A - measured 6.25A @ 201V;
3. At Tc,sat=55'C : Current@230V = 7.49A - measured 7.2A @ 195V

4. At Tc,sat=60'C : Current@230V = 8.25A - measured 8.4A @ 192.5V
5. At Tc,sat=65'C : Current@230V = 9.19A - measured 9.6A @ 189.5V
6. At Tc,sat=70'C : Current@230V = 10.33A - measured 11.1A @ 191.5V
7. At Tc,sat=75'C : Current@230V = 11.72A - measured 13.25A @ 189.5V

Now that becomes more clear, doesn't it?

1. There is some level of current-voltage trade due to poor local power supply (known - a constant source of irritation);
2. The system starts to choke up after Tc,sat ~ 55'C.

Chef
02-08-2009, 10:11 AM
Hi Chef, please feel free to contribute... :)

Your thread with Gary does not need any interferance and the data you post is very useful, certainly the most detailed for a long while - I am just watching to see how it finally pans out.

However I would like to hear more about the fluctuations you see in the compressor current versus time trace.

Just sitting on the porch.

Chef

desA
02-08-2009, 11:50 AM
However I would like to hear more about the fluctuations you see in the compressor current versus time trace.

I'll detail that aspect a bit further into the study, with a fine current-time plot. The refrigerant 'wave' events certainly become fairly clear in the region where the sub-cooling has become dominant.


Just sitting on the porch.

Take care. :)

Gary
02-08-2009, 12:25 PM
Have you checked the voltage at the mains? You could have an undersized wiring problem in addition to the known power supply problems.

Is the compressor 50hz or 60hz? Is the local power supply 50hz or 60hz? Single phase or three phase? Does the compressor have multiple power ratings?

desA
02-08-2009, 01:32 PM
Have you checked the voltage at the mains? You could have an undersized wiring problem in addition to the known power supply problems.


I'll check the mains voltage again tomorrow at the main incoming feed.

In this part of the world, the main issue is that the power infrastructure has not kept abreast of the housing development & so some areas seem to be on the power borderline. What happens here is that during peak periods the available voltage drops off - it can go as low as 187V, on the odd occasion. I've even had so bad that a single airconditioner & computer cannot be operated at the same time, before the UPS begins to cry foul... lol

I tend to perform my testing at times of the day where the incoming power is close to 220V at start of test.

For the long test of the other day, it took many hours spread over the day due to my testing technique where I bring up the heat-pump load near to a test point & stabilise at the point for a period, before taking readings.

The typical power cycle over a day follows a sine wave, with lows in morning & evening.

The local power supplier is aware of this & attributes it to a critically-loaded transformer. The advice was - "Why don't you look for another house in another area where we have better power?"... LOL

My answer is - "No problem. I'll move back to my home country." Hence my planned move in around 4-5 months. :)


Is the compressor 50hz or 60hz? Is the local power supply 50hz or 60hz? Single phase or three phase? Does the compressor have multiple power ratings?

Compressor = 50Hz
Local power supply = 50Hz

Single phase
Single power rating for this compressor motor, as far as I know. I'm led to understand that the motor is changed to suit difference supply voltages & frequency, as required. A three-phase variant of this compressor is also available, but, frankly I've been loathe to go to 3-phase out of concern for phase balance problems - it's bad-enough on single phase.

Gary
02-08-2009, 02:00 PM
As I'm sure you are well aware, the key to accurate testing is controlling the variables... and in this you seem to be currently at a disadvantage.

I suppose for the time being we will just have to play the cards we are dealt, and do the best we can under the circumstances.

Gary
02-08-2009, 02:13 PM
Hmmm... maybe you can run the 40-55C tests again at a time of day when the voltage is down to about 190V to see if the performance is dropping as a result of the lower voltage, or is in fact attributable to the subcooling.

desA
02-08-2009, 02:51 PM
As I'm sure you are well aware, the key to accurate testing is controlling the variables... and in this you seem to be currently at a disadvantage.

I suppose for the time being we will just have to play the cards we are dealt, and do the best we can under the circumstances.

What I did check was the power (W'=V*I) draw by the compressor (calculated) during the run & this is almost bang-on to the quoted power for the compressor at Te,sat=12.5'C, at each point across the test range.

So, it seems that the compressor input power is actually close to the supplier's estimates. Seemed odd at first, but I did double-check my calcs. I'll plot these separately & post the plot.

desA
02-08-2009, 02:53 PM
Hmmm... maybe you can run the 40-55C tests again at a time of day when the voltage is down to about 190V to see if the performance is dropping as a result of the lower voltage, or is in fact attributable to the subcooling.

This is easy to do - I'll run these off in early peak power draw-off - probably tomorrow. Then we'll know for sure... :)

desA
02-08-2009, 03:06 PM
http://i28.tinypic.com/2zrkcq8.png

http://tinypic.com/r/2zrkcq8/3

Here's the plot. The manufacturer's curve is the lumpy one... :)

W'c,in,m = manufacturer's compressor power estimate [kW]
W'c,in,c = calculated V*I power from experiment [kW]

desA
02-08-2009, 03:12 PM
Oversized evaporator:

Today I tested the logic of the over-sized evaporator raising the Te,sat, by applying an insulation blank over around 40% of the evap inlet surface.

The Te,sat value dropped from 15'C (today) down to 12.5'C & settled there. I let the system stabilise for around 15 minutes & then removed the blank. Te,sat then gradually returned to 15'C.

I'm wondering if it would be worth doing a test with the evap blanked off to see the effect on the overall system performance, more out of a learning experience, than much else.

desA
03-08-2009, 03:43 AM
I'd like to consolidate the control strategy so far. Please correct me, or add in the missing information.

Strategy review for riding compressor limits:

1. Lower limit : Tcomp,inlet <= 15'C
1.1 Limit suction pressure such that Tcomp,inlet < 15'C;
and/or
1.2 Control evap fan speed such that Tcomp,inlet < 15'C.

2. Lower limit : Tc,sat <= 75'C (or desired Tc,sat,op < 75'C)
2.1 Control condenser water flow such that Tc,sat < 75'C.

Gary
03-08-2009, 04:22 AM
The upper limit for the suction pressure at the compressor inlet (CPR valve outlet) should be the saturation pressure corresponding to 15C.

The fan should be controlled by a temperature sensor on the discharge line (6 inches from the compressor). The ideal setting is yet to be determined. That setting would be the discharge temp at which Tc,sat is 75C, Te,sat is 15C, and superheat is 7K.

The water regulating valve should control the flow such that the Tc,sat is 75C.

desA
03-08-2009, 05:23 AM
Super, Gary. Good, now we've got that settled. Thanks so much.

desA
03-08-2009, 05:34 AM
A few more thoughts on the evap fan speed control.

Theoretical development
I've developed an algorithm linking the various factors that affect the Te,sat value in the evap. They are as follows:
1. Ta,in = air inlet temp;
2. U = overall heat-transfer coefficient;
3. A = heat-exchanger surface area;
4. m'a = air mass flowrate;
5. Q'e = evap heat load.

We can infer things like m'a, from a knowledge of Ta,in & Ta,out (air outlet temp) & Q'e (evap heat load).

We can also derive Q'e from the refrigerant side properties, if required, & entry/exit conditions, or we can work this out from other relationships.

How is the Tcomp,inlet - fan speed control issue managed in practice?

--------------
The main issue here is that the thermodynamic requirement for Q'e (evap) reduces from the start-up load, to the final hot load. This needs to be accommodated into a control strategy - if fine fan speed control is required.

--------------
Practical implementation
An alternative thought is to simply just use the Tcomp,inlet value as the marker & force the fan speed to reduce gradually so that Tcomp,inlet <= 15'C. How would this control logic be implemented in practice? The 'internal variables' can take care of themselves, with this approach, as long as the rate of change of fan speed is gentle.

Gary
03-08-2009, 06:20 AM
The discharge temp is affected by high and low pressures, inlet superheat, heat of compression and motor heat. If we are to ride the upper limits of the compressor, the discharge temp is as good a way as any to do this.

desA
03-08-2009, 06:32 AM
Mmmhh... that's an interesting point.

So, measure the compressor discharge temp, but control/adjust what parameters?

desA
03-08-2009, 06:35 AM
Practical implementation
An alternative thought is to simply just use the Tcomp,inlet value as the marker & force the fan speed to reduce gradually so that Tcomp,inlet <= 15'C. How would this control logic be implemented in practice? The 'internal variables' can take care of themselves, with this approach, as long as the rate of change of fan speed is gentle.

What I had in mind here is perhaps some sort of simple PLC logic, which tests Tcomp,inlet. If Tcomp,in > 15'C, shut down fan one notch, if Tcomp,in < 15'C speed up one notch.

Is this more easily done in a simpler way, through standard components, rather than a PLC?

Gary
03-08-2009, 06:38 AM
By slowing the fan we can keep from exceeding our ideal discharge temp.

Gary
03-08-2009, 06:41 AM
What I had in mind here is perhaps some sort of simple PLC logic, which tests Tcomp,inlet. If Tcomp,in > 15'C, shut down fan one notch, if Tcomp,in < 15'C speed up one notch.


That's pretty much the same logic we would use for discharge temp control.

desA
03-08-2009, 06:42 AM
By slowing the fan we can keep from exceeding our ideal discharge temp.

Good point.

Now is there any downside under cold inlet air conditions that we should take care of?

desA
03-08-2009, 06:44 AM
That's pretty much the same logic we would use for discharge temp control.

Ok, great.

What is the least expensive way to implement this kind of logic, to control fan speed?

For instance, there are various fan-speed controllers on the market - some of which are programmable, with simple logic.

Gary
03-08-2009, 07:12 AM
Ok, great.

What is the least expensive way to implement this kind of logic, to control fan speed?

For instance, there are various fan-speed controllers on the market - some of which are programmable, with simple logic.

I have no idea what anything costs.

desA
03-08-2009, 07:30 AM
^ :)

Perhaps : What would be the simplest way to implement the logic, that would not require much, if any, use of electronics?

Gary
03-08-2009, 03:24 PM
^ :)

Perhaps : What would be the simplest way to implement the logic, that would not require much, if any, use of electronics?

The simplest would be a thermostat which would turn off the fan at discharge line setpoint and then turn it back on as the temp drops.

Or perhaps a two speed fan motor could be used. The thermostat would kick it down to low speed at setpoint and then back to high speed as the temp drops.

desA
03-08-2009, 03:58 PM
The simplest would be a thermostat which would turn off the fan at discharge line setpoint and then turn it back on as the temp drops.

Or perhaps a two speed fan motor could be used. The thermostat would kick it down to low speed at setpoint and then back to high speed as the temp drops.

Ok, this is pretty straightforward to implement. Thanks for that - keeps it simple.

I've seen so many electronic control malfunctions over the years, that I like to keep to as simple a system as possible.

desA
03-08-2009, 04:07 PM
Can I ask you to re-visit the correct charge selection for a heat-pump.

For instance, say the Tc,sat=35'C charge requirement was 1220g, & the Tc,sat=70'C requirement was 1015g.

The previous discussions seems to prefer loading to the 1015g charge loading option rather than the 1220g option. Why select the lowest value & perhaps not some value part way up the range?

I can understand the 1220g start-up loading causing excessive condenser floodback & sub-cooling.

Would an option be to charge 1015g + add additional gas until say sub-cooling = 7K (or even 8.5K), at Tc,sat=70'C?

This could mean that the sub-cooling at start-up could be quite low. Would this present a problem, at all?

Gary
03-08-2009, 04:26 PM
Ok, this is pretty straightforward to implement. Thanks for that - keeps it simple.

I've seen so many electronic control malfunctions over the years, that I like to keep to as simple a system as possible.

As an experimental system, I would go mechanical as much as possible, then when it is perfected consider going electronic.

Hmmm... When the "Cap and Trade" legislation goes through, energy prices will go right through the roof here in the ObamaNation. Maybe I should build one of these systems and put it up in my attic.

desA
03-08-2009, 04:38 PM
As an experimental system, I would go mechanical as much as possible, then when it is perfected consider going electronic.

Fair comment. My reason for keeping things as simple & robust as possible, is that these machines are destined for 3rd world & developing nations, in the main. Getting technicians skilled in electronics will be a rarity.

I'll think about offering a an electronic option to the more developed folks.



Hmmm... When the "Cap and Trade" legislation goes through, energy prices will go right through the roof here in the ObamaNation. Maybe I should build one of these systems and put it up in my attic.

First time I'd heard the term ObamaNation... :)

These are fun machines to develop & operate. Have fun.

Gary
03-08-2009, 05:57 PM
Unfortunately I've grown quite lazy in my retirement, but I might talk myself into building a AWHP yet.

I would start off with my all time favorite experimental platform, i.e. the smallest cheapest window shaker I can find. Unfortunately these come with a rotary compressor, which is the least suitable for this application, but I can always switch to a scroll after I have destroyed the rotary by taking it beyond its limits.

I once chilled denatured alcohol down to -30C with one of these little beasts (heavily modified). Then I stuck it in a window in the middle of winter (Michigan) and dropped the alcohol down to -60C. Cascading off mother nature. :)

desA
04-08-2009, 01:31 AM
The upper limit for the suction pressure at the compressor inlet (CPR valve outlet) should be the saturation pressure corresponding to 15C.

The fan should be controlled by a temperature sensor on the discharge line (6 inches from the compressor). The ideal setting is yet to be determined. That setting would be the discharge temp at which Tc,sat is 75C, Te,sat is 15C, and superheat is 7K.

The water regulating valve should control the flow such that the Tc,sat is 75C.

Further to this specification:

The refrigerant mass charge should be sufficient for the condenser to provide a liquid sub-cooling of 8.33-8.5K, at a Tc,sat = 75'C.

(The 8.33K comes from the ARI specification).

Is there anything else that should be added to this?

Magoo
04-08-2009, 02:14 AM
Hi desA.
After you have finalized the system can you publish the design criteria for a hot water heat pump system.
As Gary stated, us old people will need to have one given the nanny state situation with global warming taxes, and Obamanation stuff.
I congratulate you for a very interesting and informative post. Keep up with the good work.
Cheers magoo.
ps, I interprute you are now in South East Asia, and going back to South Africa. A bit like frying pan into the fire.[joke ]

desA
04-08-2009, 02:32 AM
Hi desA.
After you have finalized the system can you publish the design criteria for a hot water heat pump system.

With pleasure - it's the least I can do, after all the wonderful input we've had on this thread.


I congratulate you for a very interesting and informative post. Keep up with the good work.
Cheers magoo.

Thank you for your very wise input on the TXV super-heat setting - this rule has been extremely helpful.


ps, I interprute you are now in South East Asia, and going back to South Africa. A bit like frying pan into the fire.[joke ]

:D It is a little worrying, I must say. Have ailing parents & am the only child. What to do?

Gary
04-08-2009, 03:10 AM
Further to this specification:

The refrigerant mass charge should be sufficient for the condenser to provide a liquid sub-cooling of 8.33-8.5K, at a Tc,sat = 75'C.

(The 8.33K comes from the ARI specification).


I'm confused. ARI has specs for an AWHP? And if so, why do we care?

Gary
04-08-2009, 03:19 AM
The system thermodynamic balance calls for the following, at say Te,sat=12.5'C:
1. At Tc,sat=40'C :
Q'evap = 6.7kW
Q'cond = 7.3kW

2. At Tc,sat=50'C :
Q'evap = 6.1kW
Q'cond = 6.9kW

3. At Tc,sat=75'C :
Q'evap = 4.2kW
Q'cond = 6.1kW



Here is another set of specs that has me confused. How is any of this relevant to what we are doing?

desA
04-08-2009, 03:46 AM
I'm confused. ARI has specs for an AWHP? And if so, why do we care?

Now, this is a perfectly valid question. I'll go into a little of the Designer's Philosophy & then back out again.

Designer's philosophy
- Designing something new with no rules in place, no firm standards, since there are an infinite number of design variables to consider.
1. Select nearest fit logical set of rules/standards;
2. Modify said rules from preliminary experience;
3. Build prototype unit;
4. Test/experiment;
5. Based on performance, modify original design rules
6. Iterate items (3) - (5) until technology stable.
7. Define & develop industry standards.

The ARI rules fit into step (1) above. In addition, the compressors used are often used in the air-conditioning industry.

-------------
Backing out, to reality
- In step (6) now;
- Question:
What amount sub-cooling should be considered appropriate, throughout the heating range, that will be sufficient, to provide optimum performance & why?

desA
04-08-2009, 03:51 AM
Here is another set of specs that has me confused. How is any of this relevant to what we are doing?

See the previous post, regarding appropriate initial rules/specifications (item 1).

In this particular case, you will see that the quoted data is pegged on Te,sat & spans the Tc,sat working range.

Gary
04-08-2009, 03:54 AM
Now, this is a perfectly valid question. I'll go into a little of the Designer's Philosophy & then back out again.

Designer's philosophy
- Designing something new with no rules in place, no firm standards, since there are an infinite number of design variables to consider.
1. Select nearest fit logical set of rules/standards;
2. Modify said rules from preliminary experience;
3. Build prototype unit;
4. Test/experiment;
5. Based on performance, modify original design rules
6. Iterate items (3) - (5) until technology stable.
7. Define & develop industry standards.

The ARI rules fit into step (1) above. In addition, the compressors used are often used in the air-conditioning industry.
[/B]

Okay. The next question would be, at what point is the subcooling measured? Are they referring to SC at the condenser outlet, receiver outlet or TXV inlet? It makes a huge difference.

Gary
04-08-2009, 04:00 AM
See the previous post, regarding appropriate initial rules/specifications (item 1).

In this particular case, you will see that the quoted data is pegged on Te,sat & spans the Tc,sat working range.

The quoted data depends upon a great many more variables than Te,sat and Tc,sat. First and foremost on the list would be the temperature of the liquid entering the coil.

I'm thinking the air coil and vertical HX would do wonders for this coil.

Gary
04-08-2009, 04:08 AM
-------------
Backing out, to reality
- In step (6) now;
- Question:
What amount sub-cooling should be considered appropriate, throughout the heating range, that will be sufficient, to provide optimum performance & why?

Whatever SC fully feeds the coil (maintains SH), without interferring with the condenser, under all conditions.

desA
04-08-2009, 04:12 AM
Okay. The next question would be, at what point is the subcooling measured? Are they referring to SC at the condenser outlet, receiver outlet or TXV inlet? It makes a huge difference.

From what I've been able to determine so far, the answer is not that clear.

Different rules are followed by different compressor designers:
1. Brand name - air conditioning rating conditions:
11.1K superheat / 8.3K subcooling / 35'C ambient air over

This will vary depending on compressor application, some will have 0K subcooling.

2. Bitzer - rating conditions:
*According to EN12900 (20°C suction gas temp., 0K liquid subcooling)

No mention is made of where the subcooling is measured.

Now that you mention it, & thinking through this aspect further, where is the most logical place that an air-conditioning standard would define the subcooling to be measured?

Gary
04-08-2009, 04:23 AM
To fully feed the coil a TXV needs solid liquid at its entrance. This occurs at 5.5-8.5K SC (at the TXV). Since 5.5-8.5K SC is the point at which there is solid liquid, we would not want more than 8.5K SC at the condenser outlet in order to avoid backing liquid up into the coil.

Gary
04-08-2009, 04:35 AM
Now that you mention it, & thinking through this aspect further, where is the most logical place that an air-conditioning standard would define the subcooling to be measured?

For A/C design purposes:

Probably at the metering device inlet since it has a profound effect on the capacity of the coil.

For our purposes:

It can be assumed that the liquid will gain considerable subcooling between the condenser outlet and the TXV inlet. Thus far, our ideal SC seems to be 7K at the condenser outlet at 75C Tc,sat. I assume that this will rise well above the 8.5K minimum long before it reaches the TXV inlet.

Gary
04-08-2009, 05:31 AM
The system thermodynamic balance calls for the following, at say Te,sat=12.5'C:

Ok, let's assume that the TXV inlet SC is 8.5K


1. At Tc,sat=40'C :
Q'evap = 6.7kW
Q'cond = 7.3kW

40-8.5=31.5C liquid at TXV inlet


2. At Tc,sat=50'C :
Q'evap = 6.1kW
Q'cond = 6.9kW

50-8.5=41.5C liquid at TXV inlet


3. At Tc,sat=75'C :
Q'evap = 4.2kW
Q'cond = 6.1kW



75-8.5=66.5C liquid at TXV inlet

I would assume that at 75C Tc,sat, with 7K SC at the condenser outlet, if we further cool the liquid along the way such that the liquid line temp at the TXV is 31.5C for a subcooling of 75-31.5=43.5K SC, then our capacity would be in the neighborhood of:

Q'evap = 6.7kW
Q'cond = 7.3kW

I bet my numbers are closer than theirs.

desA
04-08-2009, 05:33 AM
To fully feed the coil a TXV needs solid liquid at its entrance. This occurs at 5.5-8.5K SC (at the TXV). Since 5.5-8.5K SC is the point at which there is solid liquid, we would not want more than 8.5K SC at the condenser outlet in order to avoid backing liquid up into the coil.

Good points.

Now, is the liquid line between condenser outlet & TXV inlet insulated, or not?

If not insulated, then allowance will have to be made for some heat-exchange between pipe & surroundings - in, or out - depending on the local ambient air conditions & the point in the heat-up cycle.

Gary
04-08-2009, 05:40 AM
Good points.

Now, is the liquid line between condenser outlet & TXV inlet insulated, or not?


No... in fact just the opposite. This is where we want to add the spiral air coil, using the waste cool air stream to further subcool the liquid as much as possible.

desA
04-08-2009, 09:43 AM
To fully feed the coil a TXV needs solid liquid at its entrance. This occurs at 5.5-8.5K SC (at the TXV). Since 5.5-8.5K SC is the point at which there is solid liquid, we would not want more than 8.5K SC at the condenser outlet in order to avoid backing liquid up into the coil.

Good - that settles that one, then.

As a first pass, let's work on the following:

Start-up condition : SC = 5.5K
Hot condition : SC = 8.5K

I'll do further experimental runs & remove gas until we get to the 8.5K condition at Tc,sat = 75'C, then see how the start-up SC settles.

desA
04-08-2009, 09:45 AM
For A/C design purposes:

Probably at the metering device inlet since it has a profound effect on the capacity of the coil.

For our purposes:

It can be assumed that the liquid will gain considerable subcooling between the condenser outlet and the TXV inlet. Thus far, our ideal SC seems to be 7K at the condenser outlet at 75C Tc,sat. I assume that this will rise well above the 8.5K minimum long before it reaches the TXV inlet.

For the test machine, the liquid line from condenser outlet to TXV inlet, is insulated.

So, we can probably assume that we've maintained SC ~ 8.5K from condenser exit, to TXV inlet.

desA
04-08-2009, 09:52 AM
I would assume that at 75C Tc,sat, with 7K SC at the condenser outlet, if we further cool the liquid along the way such that the liquid line temp at the TXV is 31.5C for a subcooling of 75-31.5=43.5K SC, then our capacity would be in the neighborhood of:

Q'evap = 6.7kW
Q'cond = 7.3kW

I bet my numbers are closer than theirs.

The sub-cooled refrigerant only contributes to the heat-transfer, if it is used to heat the incoming water stream either before entry into the condenser, or within the condenser itself.

The additional effect of sub-cooling/water interchange is to raise the COP,hp.

COP,hp = (Q'desup + Q'cond + Q'subcool) / (W'comp + W'fan)

By making use of the additional sub-cooling, we effectively raise Q'subcool & COP,hp.

:)

desA
04-08-2009, 09:58 AM
No... in fact just the opposite. This is where we want to add the spiral air coil, using the waste cool air stream to further subcool the liquid as much as possible.

For this application, I'd rather see either a dedicated sub-cooler, or have this sub-cooler built into the condenser.

At this point, to blow, what could be useful pre-heat, into the air stream, would be a waste. If the economics of a sub-cooler were to prove uneconomical, then the question I'd like to ask is this:

"Does sub-cooling of the refrigerant liquid beyond the 8.5K liquid-only temp benefit the process - say by lowering Te,sat to under 15'C?"

If the answer is "Yes", then it will be useful to have an uninsulated line & force additional cooling.

desA
04-08-2009, 02:53 PM
http://i31.tinypic.com/20r66a1.png

http://tinypic.com/r/20r66a1/3

Latest experimental runs.

Run #2 - re-set TXV, SH=0.6*TD, original charge
Run #3 - TXV @ SH=0.6*TD, blew off charge for 60 sec from LP side, to lower mass charge.

Gary
04-08-2009, 03:21 PM
Cooling the liquid before the TXV reduces flashing in the evap.

This should be easy enough to test. Do a run with the liquid line insulated and a run without the liquid line insulated and compare the results.

desA
04-08-2009, 03:30 PM
Will we always have a situation where the liquid line is being cooled by the air-stream?

For instance, at the start-up condition, with a warm air-stream, is it not possible that the liquid line actually gets heated for a period until Tc,sat lifts sufficiently above the Ta,out temp?

Practically, I'm trying to make sense of why this particular machine has its discharge line insulated.

desA
04-08-2009, 03:40 PM
Cooling the liquid before the TXV reduces flashing in the evap.

Can you expand a little further on this, please.



This should be easy enough to test. Do a run with the liquid line insulated and a run without the liquid line insulated and compare the results.

Easy to do - I'll sort it out tomorrow.

I'll do a first run, as is - to benchmark against (on the day) & then do a second run without insulation.

Where would you think best to operate the machine, to see the full effect of this change? Tc,sat=70/75'C?

Gary
04-08-2009, 04:08 PM
I think the far superior alternative is the vertical HX. This transfers the heat from the liquid to the suction, neither gaining nor losing heat. However, the resulting cold liquid at the TXV inlet minimizes flashing in the coil, in effect increasing active coil surface area.

desA
04-08-2009, 04:15 PM
I think the far superior alternative is the vertical HX. This transfers the heat from the liquid to the suction, neither gaining nor losing heat. However, the resulting cold liquid at the TXV inlet minimizes flashing in the coil, in effect increasing active coil surface area.

Is this what is referred to as the SGHX - Suction Gas Heat Exchanger?

Does this then mean that the evaporator size can be further reduced (super-heating section removal) & most of the super-heating take place in the SGHX?

desA
04-08-2009, 04:32 PM
Cooling the liquid before the TXV reduces flashing in the evap.


Ok, I've been working through the log(p),h diagram, & it's fairly obvious that additional sub-cooling, whether used in a HX, or not, will move the liquid to the left of the liquid saturation line. This will, in turn, reduce the 2-phase quality entering the evaporator (x smaller).

This should then use up some of the evap over-capacity, in having to evaporate the entering low quality liquid.

Good - that's sorted in my head now. I get to it tomorrow. Thanks so much for that applied wisdom. :)

Gary
04-08-2009, 04:41 PM
Is this what is referred to as the SGHX - Suction Gas Heat Exchanger?

Does this then mean that the evaporator size can be further reduced (super-heating section removal) & most of the super-heating take place in the SGHX?

Normally a SGHX has the TXV bulb mounted upstream, increasing the compressor inlet superheat. We want to mount the TXV bulb downstream, maintaining the compressor inlet superheat while decreasing flashing in the coil. Let's call it suction/liquid heat exchanger SLHX. I suspect you won't find this strategy in your design books.

And yes... this would allow us to reduce the coil size. Both liquid cooling to prevent flashing and superheating takes place in the SLHX.

On second thought let's call this a VSLHX as the vertical aspect adds important coil trapping advantages.

Gary
04-08-2009, 06:23 PM
http://i31.tinypic.com/20r66a1.png

http://tinypic.com/r/20r66a1/3

Latest experimental runs.

Run #2 - re-set TXV, SH=0.6*TD, original charge
Run #3 - TXV @ SH=0.6*TD, blew off charge for 60 sec from LP side, to lower mass charge.

The improvement from decreasing the coil outlet superheat from run#1 to run #2 is somewhat obscured by the increase in Ta,in, but I think it is pretty much a given that reducing coil outlet superheat improves evap capacity. And clearly the compressor is running cooler in spite of the increases in Ta,in and Te,sat. Also, the V*A decreased from 2511 to 2374 with higher Te,sat, which means the COP increased.

Run #3 clearly demonstrates that reducing the condenser outlet subcooling (by reducing mass charge) improves the performance of the condenser, reducing the approach (at Tc,sat=75C) from 9.15K to 8K, while not causing an increase in superheat.

Also worth noting is that we have increased Te,sat to 19C (well beyond the Manufacturers 15C limit) while the compressor is in fact drawing less current (V*A compared to run#1) and running cooler as well. This is a happy compressor that is nowhere near overload conditions.

desA
05-08-2009, 03:22 AM
Normally a SGHX has the TXV bulb mounted upstream, increasing the compressor inlet superheat. We want to mount the TXV bulb downstream, maintaining the compressor inlet superheat while decreasing flashing in the coil. Let's call it suction/liquid heat exchanger SLHX. I suspect you won't find this strategy in your design books.

And yes... this would allow us to reduce the coil size. Both liquid cooling to prevent flashing and superheating takes place in the SLHX.

On second thought let's call this a VSLHX as the vertical aspect adds important coil trapping advantages.

I like this strategy very much. It should allow evap coil size reduction & allow us to control the compressor suction inlet temp well.

Would this impact the overall evap fan control strategy, with input of compressor discharge?

You've mentioned 'vertical' HX as being important. Can you perhaps explain this a little more?

Would oil migration be affected by this strategy?

Gary
05-08-2009, 04:28 AM
I like this strategy very much. It should allow evap coil size reduction & allow us to control the compressor suction inlet temp well.

My thinking is that we should try to get the approach down to about 5K for both evaporator and condenser and see if the compressor can handle it. If it can then we can downsize or upsize the entire system proportionately.

At this point, the evaporator approach is 7.5K and the condenser approach is 8.0K, so we are getting closer.


Would this impact the overall evap fan control strategy, with input of compressor discharge?

The fan strategy is yet to be determined. Let's follow the trail and see where it leads us.


You've mentioned 'vertical' HX as being important. Can you perhaps explain this a little more?

The vertical heat exchanger forms a U shape with the evaporator which traps any liquid refrigerant in the coil on the off cycle.


Would oil migration be affected by this strategy?

The suction side of the VSLHX needs to be sized to maintain sufficient velocity to move the oil upwards.

Hmmmm... this is way too many initials. How about we call our gadget a vertical intercooler (VIC)?

Gary
05-08-2009, 05:29 AM
Practically, I'm trying to make sense of why this particular machine has its discharge line insulated.

Our intention is to absorb as much heat as possible in the evaporator and then transport it to the condenser. We don't want to lose any of that heat along the way. That's why the discharge line should be insulated.

desA
05-08-2009, 08:27 AM
Our intention is to absorb as much heat as possible in the evaporator and then transport it to the condenser. We don't want to lose any of that heat along the way. That's why the discharge line should be insulated.

Agreed... :)

desA
05-08-2009, 11:35 AM
Excellent analysis. Thank you.


The improvement from decreasing the coil outlet superheat from run#1 to run #2 is somewhat obscured by the increase in Ta,in, but I think it is pretty much a given that reducing coil outlet superheat improves evap capacity. And clearly the compressor is running cooler in spite of the increases in Ta,in and Te,sat. Also, the V*A decreased from 2511 to 2374 with higher Te,sat, which means the COP increased.

Agreed.

"clearly the compressor is running cooler in spite of the increases in Ta,in and Te,sat." - As shown by the lower Tcomp,disch values.

"V*A decreased from 2511 to 2374 with higher Te,sat, which means the COP increased." - This is a big bonus.


Run #3 clearly demonstrates that reducing the condenser outlet subcooling (by reducing mass charge) improves the performance of the condenser, reducing the approach (at Tc,sat=75C) from 9.15K to 8K, while not causing an increase in superheat.

The increase in T,tank temperature from Run#1 - #2 - #3 is also clear - as well as Tw,out 60.9'C -> 65.85'C -> 67.0'C. The condenser is definitely performing better.


Also worth noting is that we have increased Te,sat to 19C (well beyond the Manufacturers 15C limit) while the compressor is in fact drawing less current (V*A compared to run#1) and running cooler as well. This is a happy compressor that is nowhere near overload conditions.

This exercise has certainly taught me something.

Get the system in balance first, then measure the overall performance of that system.

desA
05-08-2009, 01:26 PM
My thinking is that we should try to get the approach down to about 5K for both evaporator and condenser and see if the compressor can handle it. If it can then we can downsize or upsize the entire system proportionately.

At this point, the evaporator approach is 7.5K and the condenser approach is 8.0K, so we are getting closer.

This is a useful strategy.

I do wonder, though, if the evap & condenser designs could restrict a little. Some folks set the design approach at 10K for the condenser, for instance.

Let's keep going... :)

desA
05-08-2009, 01:28 PM
The fan strategy is yet to be determined. Let's follow the trail and see where it leads us.


Fine. Makes sense... Good - we'll put that one on hold.

desA
05-08-2009, 01:33 PM
The vertical heat exchanger forms a U shape with the evaporator which traps any liquid refrigerant in the coil on the off cycle.

I did wonder about that. This is cunning, in that it prevents migration to the compressor & saves having to use a suction accumulator. The current circuit does not use one.



The suction side of the VSLHX needs to be sized to maintain sufficient velocity to move the oil upwards.

Fair-enough. The suction side calculation is straightforward @ greater than 5 m/s to cope with vertical uplift. The internal flow velocity in VIC will need to be carefully monitored.



Hmmmm... this is way too many initials. How about we call our gadget a vertical intercooler (VIC)?

VIC has a certain 'ring' to it... :D

desA
05-08-2009, 02:17 PM
http://i27.tinypic.com/14wrv5g.png

http://tinypic.com/r/14wrv5g/3

The latest runs.

#4 = repeat of #3 on the day - insulated discharge line.
#5 = removed insulation on discharge line, improved section of suction line insulation.

Corrected to read - 'liquid line' not 'discharge line'... :confused:

desA
05-08-2009, 02:25 PM
Comments
1. I expected the Te,sat to drop off slightly with removal of the insulation. This did not occur - but remained the same. Interesting.

2. Compressor superheat was reduced. Expected, as evap was forced to work harder.

3. Condenser sub-cooling reduced. As a result of increase evap load, more liquid was held back in evap, with less migrating to the condenser.

Further observations
a. Air inlet temp reached 32.9'C today.
b. The compressor base reached 63.2'C - the highest so far.
c. We had a 'power challenge' towards the end of Run #4, with some voltage-current trading. The amperage shoots up as the voltage drops toward 185V.

desA
05-08-2009, 02:31 PM
b. The compressor base reached 63.2'C - the highest so far.

The test machine compressor is housed in an enclosure with the tube-in-tube heat-exchanger. The outside panels of this housing are insulated.

This design may be fine for colder climates, but is not desirable in hot Asian climates.

Why a manufacturer would choose to house a compressor & uninsulated heat-exchanger in the same insulated chamber, defeats me.

Gary
05-08-2009, 02:47 PM
#4 = repeat of #3 on the day - insulated discharge line.
#5 = removed insulation on discharge line...

Now I'm confused. I thought we were talking about removing the insulation on the liquid line, not the discharge line.

desA
05-08-2009, 03:13 PM
^ Apologies - it's finger trouble. I'll correct the printout - it's been a long day of testing.

The insulation was indeed removed from the liquid line... :)

Gary
05-08-2009, 03:24 PM
The test machine compressor is housed in an enclosure with the tube-in-tube heat-exchanger. The outside panels of this housing are insulated.

This design may be fine for colder climates, but is not desirable in hot Asian climates.

Why a manufacturer would choose to house a compressor & uninsulated heat-exchanger in the same insulated chamber, defeats me.

The condenser should definitely be insulated.

desA
05-08-2009, 03:40 PM
The condenser should definitely be insulated.

I agree. This test machine has a spiral tube-in-tube condenser, with the refrigerant condensing on the outside of a spiral inner tube. The water flows inside the inner tube.

I plan to use some mineral wool insulation, or alternative high-temp insulation, around this condenser. To test the effect of their configuration.

It seems that the machine designer had perhaps thought that the container outer insulation would have been sufficient. After a long day of testing, the compressor/condenser gallery outer box surfaces become fairly hot to the touch - even with the insulation.

I'll take a picture of the set-up tomorrow & post it for reference. In my view, it's something best avoided.

Gary
05-08-2009, 04:00 PM
Some folks set the design approach at 10K for the condenser, for instance.


We can do better. In fact we already have. :)

desA
05-08-2009, 04:03 PM
http://i29.tinypic.com/2604pee.jpg

http://tinypic.com/r/2604pee/3

Picture of a similar machine. The current test heat-pump in my lab, has only two condenser coils in parallel. The insulation line appears to be missing insulation in this pic - the production machines did have insulation installed, if memory serves correctly.

The other panels comprise a thin outer metal sheet, with an inner insulation layer. When the machine is assembled, this compartment is totally closed.

desA
05-08-2009, 04:05 PM
We can do better. In fact we already have. :)

Yes, agreed. Much to my amazement. Frankly, I did not think it possible.

You, sir, are a master craftsman. :)

Gary
05-08-2009, 04:07 PM
At best the question of liquid line insulation is a trade-off. On the one hand we are losing the liquid heat, but on the other hand we reduce flashing in the evap. Given these test results, I would have to say it is too close to call.

I would expect no such ambiguity with the VIC. Not only will it bring the liquid temp down close to Te,sat maximizing the evap capacity, but it will also recover the heat from the liquid, transferring it to the suction. We win in both directions.

I expect the VIC to make a major difference.