AWHP superheat & sub-cooling

[desA]

@ #396:

Many thanks for the Grainger & JohnsonControls links. The JC aspect will be useful over here, as they have a strong presence in the region. I'll try that route directly.

[desA]

Here is another small detail that can make a big difference:

When installing a device which uses a sensing bulb (TXV, temp control, etc.) the bulb should be mounted with the cap tube upwards.

If the cap tube is on the bottom, when the bulb gets warm the liquid portion of the charge can be pushed through the cap tube to the power element. Then the power element temp is sensed instead of the bulb temp.

It is also a good idea for the power element to be above the bulb with its cap tube on the bottom, so that any liquid can drain back to the sensing bulb.

Like the fine print in a contract, the devil is in the details.

Some excellent points. Thanks so much for that. It is these kinds of details that come from years of experience, to be sure.

For the lab TXV, the sensing bulb is located some 500mm below the TXV position, bulb on pipe upper portion - insulation-wrapped. The TXV power element lies horizontally, top-side. In other words, TXV inlet from bottom, TXV outlet horizontal.

I'm trusting that, with this arrangement, that the bulb liquid stays in the bulb.

[Gary]

Cap tube up on the sensing bulb is the important part.

On a TXV the problem is usually self-correcting. When the bulb gets cold enough it sucks the liquid back in and everything goes back to normal. But in the meantime, the valve does strange things and figuring it out can drive you crazy.

[Gary]

How is the testing coming along?... making any progress?

[desA]

^ Spent the past two days in something like 33-35'C (indoor) temperatures. Pretty knackered after all that.

Here's the run-down so far:
1. Scrounged two spare digital temperature readouts off an old lab machine.
2. Scrounged two thermocouples (wrap-around clamp-type).
3. Attached the T/C's to the condenser in/out, just below an elbow.
4. Wired up the temp readouts & T/C's. Noticed way too much digital bounce on the temp readouts.
5. Scrounged two digital controllers off spare lab machines.
6. Wired T/C's & power, set input filter, to filter out T/C noise. Ok.
7. Calibrated T/C's at 0'C (ice/water mix) & 100'C (boiling water). At zero, read fine - at 100'C, read around 95'C. Calibration curve to follow.
8. Added 100g of refrigerant - based on earlier low SC.

9. Ran heat-up & hold at following test points - Tc,sat:
50/55/60/65/70 - as per previous system settings.

10. Decreased water flow to raise dT,water from 2.3 - 3.1K, at Tc,sat=70'C.

11. As per (10) & trimmed evap fan back to around 70% of normal running condition, at Tc,sat=70'C.

I'm busy writing everything up, so that we can make sense of it all. I should have this up in the morning (head hurts from too much heat )

----------

A question, regarding the band thermocouples on the condenser inlet/outlet pipes (5/8"). I'm seeing an offset compared to the tank immersion probe, which climbs at water temps rise. The T/C calibration checks out, so I put this error down to pipe wall thickness.

The clamps have to fit on the outside of a 5/8" copper elbow. They are well lagged, & the Tw,in / Tw,out / T,tank probes were all set to read the same at startup, before the heat-pump was switched on.

What amount of temperature offset is typical for these band-type T/C's mounted on the outside of a pipe?

I need to correct for this, before we can make sense of the output data.

The other thermocouples are bare tip beaded end which I tape securely on the pipe surfaces (after sand-papering the surface). They are reasonably consistent. The banded T/C's seem consistent, but with growing error as temp rises.

[Gary]

Are all of the sensors insulated?

[desA]

The sensors on the refrigerant circuit are all taped in place & then securely wrapped in further tape. That is as far as it goes for the bare pipe. Where the lines have insulation, the insulation is replaced over the wrapped sensors.

So, basically, the sensors are fairly well protected from the surrounding air-stream, & even for the bare pipes, would be seeing a similar surface effect to the pipes themselves.

These readings were always very consistent - bead T/C's are good that way.

The problem is with these clamped T/C's. I wrapped them, after installation, because the wall between fluid & sensor is much greater than between sensor to air. This was to prevent excess air-side influence.

What I'm busy doing now, is to calculate the temperature offset from fluid to sensor position itself. I can then use this to correct the probe readings. I'll probably have something sensible on that side tomorrow, after I run up a spreadsheet.

[Gary]

In addition to taping the sensor, if you extend the taping upstream a few inches and downstream a few inches, this will lessen the influence of the air temp on the pipe.

[desA]

In addition to taping the sensor, if you extend the taping upstream a few inches and downstream a few inches, this will lessen the influence of the air temp on the pipe.

Good idea.

In a sense, this has to be done in order to practically hold a bead T/C in place, as they tend to worm their way out of location, as the pipe heats up. I usually tie the T/C down firmly around 25-40mm from the sensing position.

When I tape down the sensor bead, I usually wrap an inch, or so away as well.

A little extra tape is a good thing.

[desA]

Compressor operating conditions (re-visited)

I queried the name-brand compressor technical staff in both Asia & Europe on the correct safe operating envelope for the model presently used. I received a phone call from their Hong Kong compressor specialist & a few days later an e-mail from their local specialist.

I would like to confirm you that the approved operating envelope is +13 degree C for Evaporating temp and +65.6 degree C for Condensing temp. Please try to operate the compressor in this envelope. Then, the compressor life time will be as our design (approx. 10 years).

So, the recommendaton is fairly clear, on the continuous operating envelope. For short excursions outside the operating envelope, with Tcomp,disch well controlled, the one technical specialist considered this to be fine. When I mentioned the Tcomp,disch ~ 91'C @ Tc,sat=70'C, he was not at all unhappy.

Now, the ultimate questions are:
1. What maximum operating temperature should a heat-pump manufacturer advise his customers to use?
2. What would constitute allowable operating excursions - how severe & for how long?
3. What is the real reason for the Te,sat=13'C upper limit (the Tc,sat=65.6'C is clear)?

For the record:
I know of a number of instances in Australia, where a particular heat-pump manufacturer has had a number of compressor failures - many within 1-1.5 years of operation - using the same family of compressors used in the lab machine. In that case, there seems to be little control of maximum water discharge temperature, relative to Tw,out.

[desA]

The role of the condenser approach temp

If the squeeze is on Tc,sat being as close to 65.6'C under a continuous running condition, then the condenser approach temp needs to be as small as possible, in order to produce as high a Tw,out as possible.

So, I'd like, for the next few trials, to look at how to push the approach down as far as possible on the lab machine. The Tc,sat maxima will be 65'C - 70'C, with 70'C being allowed for short-time excursions of on-off heating cycles, with periodic recovery time.

[Gary]

Set the water regulating valve to maintain Tc,sat 65.6C.

Adjust the discharge line/fan control to maintain Te,sat 13C.

[desA]

Set the water regulating valve to maintain Tc,sat 65.6C.

Ok, good, makes sense.

So, what would happen if the loop circulation pump pushed harder than the heat-pump can tolerate, is that the water regulating valve will throttle back to only allow sufficient flow through the condenser. This protects the heat-pump against outside influences.

Adjust the discharge line/fan control to maintain Te,sat 13C.

Now, we are measuring Tcomp,sat, not Te,sat (or Pe,sat). In that case, surely we would rather measure Te,sat?

Further thoughts on Tcomp,disch - fan speed control
Progressing the Tcomp,dish - fan speed control a little further. What is the reasoning for the Te,sat=13'C limit in the first place? How does this relate to Tcomp,disch practically in the compressor?

As Te,sat rises, the compressor amperage drops off, COP rises, mass flow rises, heat rejected rises, evap capacity rises. Surely, a higher Te,sat should be a good thing, as long as Tcomp, suction is well controlled? Tcomp,suction would be reflected in Tcomp,disch.

[Gary]

As I see it the important criteria are compression ratio and discharge temp. I really don't know why they want 13C or 65.6C... but then they may have reasons that I am not aware of.

[desA]

As I see it the important criteria are compression ratio and discharge temp.

I agree with your logic here, entirely.

I really don't know why they want 13C or 65.6C... but then they may have reasons that I am not aware of.

I'm beginning to get the impression, by the answers given, that it has more to do with sticking to known long-run test data, rather than an understanding of the fundamentals & reasoning behind the various inter-connected factors.

Ironically, this name-brand compressor is arguably one of the oldest & most established in the industry - yet, when pressed, they all, without a deviation, revert to the textbook answers. The fact that the textbook answers & their technical bulletins leave a lot of room in-between, makes me wonder how much they really know, to be brutally honest.

Practically, I'm beginning to lose some level of confidence in their analytic skills, since the debate always shifts back to "well, this is what we know - fit in", rather than, "well, this is what we know, so far - these are the surrounding tests & their limits. Select your safety margin".

[desA]

The current experimental data for the lab heat-pump.

Run 7 - hole in compressor partition, condenser well insulated, filter-drier set to 30' off horizontal axis, 100g of refrigerant added.
Run 8 - as per 7, with water flow throttled to raise dTw to 3.1K.
Run 9 - as per 8, with evap fan speed reduced to 66.7% of starting velocity.

The water temps have been calibrated against previous runs using Tw,out & T,tank as the basis for correction. The dTw between new digital readouts determines Tw,in.

Also note additional sensors at Txv,in , Txv,out , Ta,evap (directly behind evap). The Ta,out value was always for the outlet of the box, here giving a measure of pirate air leakage into the box.

-----

Updated pic to show Tw,in values - these are calibrated against previous experimental measurements & the current dTw per digital readouts.

[desA]

Preliminary diagnosis at Tc,sat=65'C (open to debate)

1. Evap -
1.1 SH marginally higher (same TXV settings as before);
1.2 SH/TD has climbed from 0.54 up to 0.63 (controlling well across range);

2. Condenser -
2.1 SC low - dropped from 5.1K to 2.1K;

Add further refrigerant until condenser SC approx 5.5K.
No adjustment to TXV SH adjuster yet.

[Gary]

Run 8 - as per 7, with water flow throttled to raise dTw to 3.1K.
Run 9 - as per 8, with evap fan speed reduced to 66.7% of starting velocity.

Air flow and water flow are limiting factors, either of which, when reduced, should cause a drop in performance. The only reason we would reduce these flows is to safeguard the compressor and/or throttle the system in order to remain within the manufacturer's envelope.

The purpose of the water regulating valve is to throttle the flow at the beginning of the cycle so that the water is fully heated before delivering it to the tank. The flow should increase as the incoming water temp increases.

I don't understand your purpose in throttling the system air and water flows in these test runs.

[desA]

Originally Posted by desA
Run 8 - as per 7, with water flow throttled to raise dTw to 3.1K.
Run 9 - as per 8, with evap fan speed reduced to 66.7% of starting velocity.

Gary:
Air flow and water flow are limiting factors, either of which, when reduced, should cause a drop in performance. The only reason we would reduce these flows is to safeguard the compressor and/or throttle the system in order to remain within the manufacturer's envelope.

The purpose of the water regulating valve is to throttle the flow at the beginning of the cycle so that the water is fully heated before delivering it to the tank. The flow should increase as the incoming water temp increases.

I don't understand your purpose in throttling the system air and water flows in these test runs.

Run 8 - throttling the water flow
With a slowing down of the water flow, the condenser will force a large dTw on the waterside. This will alter the heat-balance in the condenser. Since the water-side is often not the dominating fluid in the condenser, the effect on overall performance can be minor, to nothing. On the other hand, the effect on the condensation can be quite a lot, forcing slightly better sub-cooling - or, so the heat-exchanger calcs predict.

This run was to see what, if anything changed with reduced flow & increased dTw. Judging from the experimental results, it did the exact opposite of what I had thought. I'll set that one aside until later, when the machine is balanced at full water-flow conditions.

Run 9 - throttling the evap air flow
Again, a test to gauge the effect on the system by using the evap fan control strategy to lower Te,sat.

Background on this particular evap
Original evap (@R22) : Qe=11.23kW
Re-rated to R-134a : Qe=13.03kW (I have doubts)
Design air in velocity : va,in = 2.03 m/s
Operating air velocity: va,in = 3.3-3.6 m/s

Thermodynamics @ Te,sat=12.5'C
@ Te,sat=35'C : Qe,reqd=7.0kW
@ Te,sat=55'C : Qe,reqd=5.7kW
@ Te,sat=70'C : Qe,reqd=4.6kW

Observations
1. The coil supplier re-rated the coil, which had been designed for R-22 service, to R-134a. He raised the performance estimate from 11.23kW to 13.03kW at va,in=2.03m/s.

2. The maximum thermodynamic evap requirement at Tc,at=35'C is Qe=7.0kW.

3. With the increased fan speed, over design requirement, the Q,evap is likely to rise even further to around:
Q'e = (3.3/2.03)^0.8*Q,e = (1.475)*Qe

4. Re-rating evap for new fan speed @ 3.3m/s:
Original evap (@R22) : Q'e=(1.475)*11.23=16.56kW
Re-rated to R-134a : Q'e=(1.475)*13.03=19.22kW

5. Compare Q'e=19.22kW to the Qe=7.0kW thermodynamic requirement. This huge mis-match must surely create a system imbalance. Practically, this evap is way too large for the current heat-pump it is feeding.

What effects would be expected from an over-size evaporator on the heat-pump thermodynamics?

[Gary]

What effects would be expected from an over-size evaporator on the heat-pump thermodynamics?

I suppose it would drive the compressor right up to its upper limits... but isn't that what we are trying to do?

[Gary]

It has been a very long time since I was in refrigeration school, but it seems like I recall there being a formula for measuring heat transfer in a coil. Something like CFM times dT times 1.08 equals BTU or some such. And there must be something similar for water flow.

[desA]

Originally Posted by desA
What effects would be expected from an over-size evaporator on the heat-pump thermodynamics?

Gary:
I suppose it would drive the compressor right up to its upper limits... but isn't that what we are trying to do?

Some thoughts on the heat-pump system

The thermodynamic balance calls for the following:

Q'e + W'in = Q'c

Now, if Q'e (evaporator) is vastly over-rated & keeps pumping thermal energy into the refrigerant circuit, there must be a trade between W'in (compressor) & Q'c (condenser), to retain thermodynamic equilibrium. The TXV will also influence the dynamics due to its control of superheat, via m'g (refrigerant mass flow).

If we push too much energy into the system, it has to be removed, somehow, & system equilibrium restored, as I see it. If the imbalance is large, then I'd expect a certain amount of circulating energy to move around the circuit. The TXV can probably handle some of this - as seen by the holding of Te,sat reasonably well during the first part of the heat-up cycle. After a certain point, where the system expects a drop-off in Q'e & it does not come, then the system must retain the energy & heat up.

There will have to be a corresponding condenser (Q'c) oversize to match the oversize Q'e, but, what will happen to the compressor under these conditions? What will the suction & discharge temps to/from the compressor look like, motor amps, power draw etc?

[desA]

What I propose to do, is to calculate a little further on the experimental data obtained so far & do the evaporator & condenser heat-balances. Then try to reconcile this with the compressor power consumption.

Let's see where this all goes.

[desA]

It has been a very long time since I was in refrigeration school, but it seems like I recall there being a formula for measuring heat transfer in a coil. Something like CFM times dT times 1.08 equals BTU or some such. And there must be something similar for water flow.

I'll do both of these balances & see what comes out of the analysis.

[Gary]

As I understand it, the thermodynamic equation always balances itself. A change on one side of the system causes a change on the other side of the system.

Upsizing both the evap and cond lessens the approaches and causes the compressor to work harder, which increases the energy efficiency.

A great deal of the rise in energy efficiency in recent years is due to upsizing the coils.

But this has a cost, and upsizing coils has diminishing returns.

[Deniver45]

Yes that is right with the abient temperature at 75 degrees. Of course on a cold day, an outdoor AC might have a super heat of 5 degrees.

[Gary]

What I propose to do, is to calculate a little further on the experimental data obtained so far & do the evaporator & condenser heat-balances. Then try to reconcile this with the compressor power consumption.

Let's see where this all goes.

If we were designing this for pool heating or space heating, then this would need to be a recirculating system. This seems unneeded for your target area.

It makes more sense to me to design for feedwater heating for applications which consume hot water (kitchens, bathrooms, etc.), with a relatively minor amount of recirculation to maintain tank temperature.

So far, all of our testing assumes recirculation.

On the water side the formula would be GPH * 8.33 * dT (oF) = BTUH

In heating feedwater, the incoming water temp and leaving water temp (indirectly controlled by the water regulating valve) should be relatively constant. Given constant dT, our primary variable would be GPH. Given a relatively small incoming air temp range (25-35C), we should be able to design a very stable system with a constant high COP.

What do you think?

[desA]

And let's not forget:

3. CPR valve.

...

The CPR would do the job and in fact would be the simplest. The evaporator pressure increases with the load, but the CPR will allow no more than setpoint pressure to enter the compressor.

The fan control strategy responds directly to the compressor heat and I'm thinking will be the best in safeguarding the compressor.

Gary,

Can we re-visit the CPR again, in combination with the Tcomp,disc - fan speed control?

What's going through my thoughts is to use the CPR as a way to absolutely ensure that a maximum Te,sat=13'C is ensured at entry to the compressor, should the evap pull up the Te,sat value from its start-up value. The fan control strategy can then be used as a trim, rather than as a pure safety mechanism. If, for any reason, the fan speed control still lets the Te,sat drift up above 13'C, then the CPR kicks in & holds the compressor inlet safe.

Do you perhaps have any links as to what exactly this device looks like & potential suppliers?

I'm beginning to see that this is a safety device put in place to prevent compressor issues.

[Gary]

Gary,

Can we re-visit the CPR again, in combination with the Tcomp,disc - fan speed control?

What's going through my thoughts is to use the CPR as a way to absolutely ensure that a maximum Te,sat=13'C is ensured at entry to the compressor, should the evap pull up the Te,sat value from its start-up value. The fan control strategy can then be used as a trim, rather than as a pure safety mechanism. If, for any reason, the fan speed control still lets the Te,sat drift up above 13'C, then the CPR kicks in & holds the compressor inlet safe.

Do you perhaps have any links as to what exactly this device looks like & potential suppliers?

I'm beginning to see that this is a safety device put in place to prevent compressor issues.

http://us.refrignet.danfoss.com/RA/D...C.PD.HH0.A1.22

[desA]

^ Can I ask you to perhaps e-mail this to me, as I don't seem to be allowed access into that part of the Danfoss site? I'll pm you my e-mail details (I think you also have these on record).

[desA]

That US Danfoss site is way better than the European one - to be sure... So easy to find stuff...

[Gary]

Does this work?:

http://www.ra.danfoss.com/TechnicalI...CPDHH0A122.pdf

[desA]

Is this the one?

http://us.refrignet.danfoss.com/ra/P...-00508BF7E573}

[desA]

Does this work?:

http://www.ra.danfoss.com/TechnicalI...CPDHH0A122.pdf

Perfect - works just fine. Thanks so much for that.

Now, what are the finer details about how to set this device, where to place it in the suction line (near evap, or near compressor) - & any other advice.

I'm going to get some ordered in asap. I want to see how this device runs in the lab machine. I'm starting to feel that this is a necessary part of a safe design principle in these heat-pumps, if long compressor life is required. Just has to be done.

The fan control strategy can then be used as a fine trimming option, with the basic system integrity being governed by the CPR.

[Gary]

Now, what are the finer details about how to set this device, where to place it in the suction line (near evap, or near compressor) - & any other advice.

In the picture, at the right side is a brass cap. Under this cap is the adjustment screw. The red label tells you which way to turn it.

Install this close to the compressor. The stamped in arrow shows the direction of refrigerant flow.

Wrap the body with wet rags while brazing it in.

[desA]

^ Ok, great.

I've been sleeping on this aspect a great deal & came to the firm conclusion that the CPR has to be there, as the basic compressor Te,sat<=13'C (via P,LO) protection mechanism.

To guarantee Tc,sat<=65.6'C, is where the fan control can play a role, as well as simply locking in a Tw,out maximum in the controller, to prevent Tc,sat overshoot. This will also be strongly referenced in the technical manuals for the heat-pumps.

If the user should select to over-ride the Tc,sat part, then he can, of course, gain extra performance, at his/her own risk - violating product warranty.

I will test all lab units up to Tc,sat=70'C/75'C to ensure reliability, internally - at least to see where things go in terms of the real unit performance. System tuning will be at Tc,sat=65/70'C.

Really, I think that this then allows the end-user to obtain a good, safe, cost-effective heat-pump. For higher water delivery temperatures, I will use the much higher-priced high Tc,sat compressors from Europe, or elsewhere - but, these will push the price of the whole machine pretty high. This would be a new, high-performance range of machines.

[desA]

Suction optimisation
Now that we've settled on a CPR as being a safety mechanism for the compressor suction Te,sat<=13'C (via P,LO), can we discuss the possible reasons for the particular Te,sat=13'C as the limit for compressor operation. (Reasoning : chopping Te,sat early, will constrain system performance.)

Could this be something along the lines of:
Tcomp,suct = Te,sat + SH = 13 + 7 = 20'C ?

In other words, a restriction on Tcomp,suct, more than anything else?

-----------

If this were the case, what would/could the effects of the following combinations be:

Tcomp,suct = Te,sat + SH = 14 + 6 = 20'C ?
Tcomp,suct = Te,sat + SH = 15 + 5 = 20'C ?

In other words, to enlarge the evap Te,sat window, but control SH tightly?

With a scroll compressor, with built-in accumulator (oil reservoir/motor housing) with refrigerant holding capacity in the range of 4kg, odd - what could the lowest safe SH be? One hears talk of SH in the range of 3-5K for some heat-pumps. Not sure if this applies to all, though.

[Gary]

^ Ok, great.

I've been sleeping on this aspect a great deal & came to the firm conclusion that the CPR has to be there, as the basic compressor Te,sat<=13'C (via P,LO) protection mechanism.

To guarantee Tc,sat<=65.6'C, is where the fan control can play a role, as well as simply locking in a Tw,out maximum in the controller, to prevent Tc,sat overshoot. This will also be strongly referenced in the technical manuals for the heat-pumps.

If the user should select to over-ride the Tc,sat part, then he can, of course, gain extra performance, at his/her own risk - violating product warranty.

I will test all lab units up to Tc,sat=70'C/75'C to ensure reliability, internally - at least to see where things go in terms of the real unit performance. System tuning will be at Tc,sat=65/70'C.

Really, I think that this then allows the end-user to obtain a good, safe, cost-effective heat-pump. For higher water delivery temperatures, I will use the much higher-priced high Tc,sat compressors from Europe, or elsewhere - but, these will push the price of the whole machine pretty high. This would be a new, high-performance range of machines.

I assume you meant to say "water regulating valve", not "fan control".

[Gary]

Suction optimisation
Now that we've settled on a CPR as being a safety mechanism for the compressor suction Te,sat<=13'C (via P,LO), can we discuss the possible reasons for the particular Te,sat=13'C as the limit for compressor operation. (Reasoning : chopping Te,sat early, will constrain system performance.)

Could this be something along the lines of:
Tcomp,suct = Te,sat + SH = 13 + 7 = 20'C ?

In other words, a restriction on Tcomp,suct, more than anything else?

-----------

If this were the case, what would/could the effects of the following combinations be:

Tcomp,suct = Te,sat + SH = 14 + 6 = 20'C ?
Tcomp,suct = Te,sat + SH = 15 + 5 = 20'C ?

In other words, to enlarge the evap Te,sat window, but control SH tightly?

With a scroll compressor, with built-in accumulator (oil reservoir/motor housing) with refrigerant holding capacity in the range of 4kg, odd - what could the lowest safe SH be? One hears talk of SH in the range of 3-5K for some heat-pumps. Not sure if this applies to all, though.

In addition to everything else, the VIC will reduce the coil outlet superheat to near zero while still maintaining sufficient superheat at the compressor inlet.

[desA]

In addition to everything else, the VIC will reduce the coil outlet superheat to near zero while still maintaining sufficient superheat at the compressor inlet.

Very fair comment, indeed.

Let's start working on the VIC as the next target.

Do you have any specific ideas at this stage on the VIC - shape, vertical, tube-in-tube, diameters, length, suction gas on which side, liquid refrigerant on which side, internal turbulation etc.

[Gary]

Very fair comment, indeed.

Let's start working on the VIC as the next target.

Do you have any specific ideas at this stage on the VIC - shape, vertical, tube-in-tube, diameters, length, suction gas on which side, liquid refrigerant on which side, internal turbulation etc.

http://www.ra.danfoss.com/TechnicalI...1/RD6KA502.pdf

[desA]

^ Nice piece of equipment. I'll check the Danfoss price, but, I'm sure it's going to be incredibly high, with all their castings/forgings & such like.

What about using a sub-cooler for this job? I have 3 on hand of various sizes, using an internally twisted tube.

http://www.vaportec.co.nz/specsheets/subcoolers.jpg

[Gary]

^ Nice piece of equipment. I'll check the Danfoss price, but, I'm sure it's going to be incredibly high, with all their castings/forgings & such like.

What about using a sub-cooler for this job? I have 3 on hand of various sizes, using an internally twisted tube.

http://www.vaportec.co.nz/specsheets/subcoolers.jpg

If you have one with the right connection sizes, then... a heat exchanger is a heat exchanger.

[desA]

If you have one with the right connection sizes, then... a heat exchanger is a heat exchanger.

Let me go through carefully which one to fit in the lab machine. It will give us a good idea of what to expect down the track.

While we're at it, are there any other requirments that we should consider:
- minimum height;
- maximum pressure drop allowable (both sides);
- flow velocity;
etc.

Why I ask, is that, I recently commissioned some fin tooling & the first fin profiles could serve this purpose very nicely. I have a second round of tooling in the pipeline & this project could just give me the shove I need to get them cut & sent over. I can then have the 2nd generation of VIC's made up over here to best suit our requirements.

[Gary]

Let me go through carefully which one to fit in the lab machine. It will give us a good idea of what to expect down the track.

While we're at it, are there any other requirments that we should consider:
- minimum height;
- maximum pressure drop allowable (both sides);
- flow velocity;
etc.

Why I ask, is that, I recently commissioned some fin tooling & the first fin profiles could serve this purpose very nicely. I have a second round of tooling in the pipeline & this project could just give me the shove I need to get them cut & sent over. I can then have the 2nd generation of VIC's made up over here to best suit our requirements.

As I pointed out early on, I am not a design engineer. I would do this by trial and error.

That said, I want to maintain the same velocity on the suction side and minimize pressure drop on both sides, especially on the suction side.

[desA]

^ Fair-enough. I'll put together some specs in terms of pressure drops, performance & so on - then we can kick it around.

Some ball-park questions (can refine these later):
1. With it being mounted vertically - any idea of say a minimum vertical height you'd like to see? Is 250mm enough, for instance?

2. What about evap liquid retention? Would we want to take up some of the liquid in the VIC during off-time, rather than have it migrate to compressor lower shell?

3. For the suction gas, would we say, for instance, come out of the evap at saturate condition, then superheat all in the VIC (7K)?

4. What kind of target liquid sub-cooling dT would we be looking for across the VIC (say 5K)?

[Gary]

Some ball-park questions (can refine these later):
1. With it being mounted vertically - any idea of say a minimum vertical height you'd like to see? Is 250mm enough, for instance?

I would like to see the suction line go up near the height of the evap. The VIC needs to fit within that vertical line and be long enough to get the job done.

2. What about evap liquid retention? Would we want to take up some of the liquid in the VIC during off-time, rather than have it migrate to compressor lower shell?

The VIC will make migration less likely. If you don't have migration problems without the VIC, then you won't have migration problems with the VIC.

3. For the suction gas, would we say, for instance, come out of the evap at saturate condition, then superheat all in the VIC (7K)?

Yes.

4. What kind of target liquid sub-cooling dT would we be looking for across the VIC (say 5K)?

We want the liquid temp at the TXV inlet to be close to Te,sat.

[desA]

Originally Posted by desA
4. What kind of target liquid sub-cooling dT would we be looking for across the VIC (say 5K)?

Gary:
We want the liquid temp at the TXV inlet to be close to Te,sat.

Some basic heat-balances across the VIC
For perfect heat-exchange liquid-to-refrigerant, based on the stream balances & available specific heat property data, the liquid line temperature drop is:

dTliq = (1.122)*dTvap

@ dTvap=7K => dTliq = (1.122)*(7) = 7.86K !!!

These are the physics of heat-transfer, unfortunately.

What is really required in order for the VIC to bring the liquid temperature down further, is to have it take on some of the evaporation duty, off the evaporator.

The other alternative is to simply just use the evap off air to reduce the liquid line temperature, through a finned line cooler.

[Gary]

Some basic heat-balances across the VIC
For perfect heat-exchange liquid-to-refrigerant, based on the stream balances & available specific heat property data, the liquid line temperature drop is:

dTliq = (1.122)*dTvap

@ dTvap=7K => dTliq = (1.122)*(7) = 7.86K !!!

These are the physics of heat-transfer, unfortunately.

What is really required in order for the VIC to bring the liquid temperature down further, is to have it take on some of the evaporation duty, off the evaporator.

And with the TXV bulb being mounted at the heat exchanger exit instead of the evap exit, that's what it will do.

Given just sensible heat transfer, trading subcooling for superheat, the heat exchanger is a straight trade-off, with little if any gain.

In moving the TXV bulb to the outlet, the heat exchanger becomes an extension of the evaporator, with latent heat transfer on the suction side.

[desA]

And with the TXV bulb being mounted at the heat exchanger exit instead of the evap exit, that's what it will do.

Given just sensible heat transfer, trading subcooling for superheat, the heat exchanger is a straight trade-off, with little if any gain.

True. The calculations are showing this to be the case.

In moving the TXV bulb to the outlet, the heat exchanger becomes an extension of the evaporator, with latent heat transfer on the suction side.

The nice things about having the TXV at outlet is that it forces an area exchange in the evap/VIC system to ensure consistent superheat at exit. The check-&-balance.

Now, the next thing to work on is how much of the evaporator load to shift across to the VIC & then design/size the VIC so that it boils off the refrigerant properly, without flooding, or liquid droplet carry-over. The balance part is no real sweat, given where we want the final liquid temp to head.

The trade is going to be between a pool-boiling design & convective boiling design - I'm going to have to think about this a little more. Actually, if pool-boiling is present, then I'm going to have to think carefully a little more about the liquid routing, as the concept of parallel versus counterflow no longer applies under phase change - only for the sensible heat section of the unit.

The VIC is morphing into VICE (vertical intercooler evaporator)...