AWHP superheat & sub-cooling

[desA]

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]

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]

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]

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]

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 ]

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

[Gary]

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]

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]

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]

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]

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]

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]

-------------
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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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]

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...

[desA]

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'...

[desA]

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]

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]

#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]

^ 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]

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]

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]

Some folks set the design approach at 10K for the condenser, for instance.

We can do better. In fact we already have.

[desA]

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]

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]

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.