Mikesolar,

Thank you so much for the retrofit articles! I wish we could sticky these. They are a must for anyone attempting projects with flammable refrigerants. The articles put safety first and discuss proper handling techniques for these potentially dangerous gases. Very well authored.

dc€x,

I have just recently become aware of Coolpack and have used the software a few hours, but I still haven't become proficient enough with it to get much meaningful data. AC_Hacker has started a tutorial thread about it. Hopefully he can teach us noobs something.

I am aware of the HC refrigerants and their applications except DME. Really? Dimethyl ether? This stuff has been taken out of everything it used to be in except spray paint and vcr tape head cleaner spray due to people getting high with it. I googled it and found that ASHRAE has labeled it R-E170. Wow, higher COP than R12? Maybe someday we will have sub-mini-splits running off ether!

Honestly, I believe the 3 winners in this new arena will be propane, butane, and co2. All are common and inexpensive, and they actually work well. I've experimented with propane (r-22 replacement) and butane (r-12 replacement), with good results with both gases.

The guentner knowhow article I cited a few days ago titled "Subcooling but Correctly" briefly covers the TX valve operating at its limits. It's near the end of the article. Basically, it says that with extreme subcooling (enough that the liquid line could sweat), the valve bangs against its seat when it shuts, much like when flash gas is fed into it. Either condition is to be avoided. I read the same thing somewhere else also but I can't find it now.

I found a suitable HX on ebay for $35 plus shipping.

http://i1326.photobucket.com/albums/...psf7ab4dcc.jpg

3/8" line wound around 1" pipe with reducers at each end. 11" overall length, Sized for 10 degF subcooling, 20 degF superheat at 1.5 ton capacity.

All major manufacturers recommend the same superheat/subcooling values as this example for a sizing reference. Most say this is all the transfer you should need. Some say to go up one size for extreme performance. Who am I to argue?

I believe I am ready to build a SLHX for my unit. Since the compressor is rated at 10500 btu, I will borrow from this example and wind 10 turns of 3/8 inch copper refrig. line around a 1" pipe as a liquid side exchanger. For the suction accumulator, I'll chop an evacuated and purged propane bottle down to about half its original size, plumb the shell, then braze it all up. Wish me luck, please.

What I did is to get the next size up tube from the suction line, put 2 reducing tees at either end and braze the sucker up. The tees had to have the stops drilled out so the original suction line would go right through. 8" was long enough for good heat transfer.

Mikesolar,

So did you get your heat pump tested/working? Haven't seen anything new on your thread. What transfer temps are you getting out of the suction exchanger?

I hate to disappoint but I haven't got it running yet. I am waiting for a hall effect flow meter to arrive which I will hook up to the top of a manual 4-way diverting/mixing valve that I am making out of PP (see attached photo). This will have two pumps one for the evap and one for the condenser going to/from the tanks in the next picture. Meanwhile the dog shows off her ribbons.

The parts will be mounted on the plywood so I can test a variety of heat pumps.

http://ecorenovator.org/forum/attach...0215-00543-jpg

http://ecorenovator.org/forum/attach...0215-00544-jpg

http://ecorenovator.org/forum/attach...0211-00537-jpg

I know this is an old thread but that don't change it's importance.

Actual Testing, not simulation:

http://fire.nist.gov/bfrlpubs/build00/PDF/b00014.pdf

See fig 4.4 R290 heating mode ICHX vs Basic (without ICHX)

ICHX looses about .4 COP

BUT, this also shows same .4 COP loss for for 4*C superheat.

This conflicts with many other papers tha show great gain in COP with 30-35*C superheat.

SO FOR NOW I DON'T BELIEVE IT!


See figure 4.2 R290/R600a isobutane 70/30% mixture sucks!

http://inpressco.com/wp-content/uplo...r48226-229.pdf

SLHX is very beneficial in this testing.

Scientific research in general is not as concrete or absolute as you might think. Most scientists and engineers are paid to provide research and published papers that support a certain bias or agenda. Numbers are twisted and parameters are modeled to produce a certain outcome, if not dictated outright.

Concerning r22 and propane, the pt curves are much like r12 and r134a. Close enough for me to call even, but far enough apart for engineering field days.

http://www.danfoss.com/NR/rdonlyres/...ompressors.pdf

p122

"For R290 experience shows the need for a
capillary flow rate almost similar to R404A. At
least this is a good starting point for optimization.
As with R134a, R404A and R600a the suction
line heat exchanger is very important for system
energy efficiency of R290, which it was not for
R22, see fig. 5. The figure shows increase of COP
with superheat from few K up to +32 °C return
gas temperature, where a range from +20 °C
to approx. +32 °C is usual for small hermetic
systems.
This large increase in COP for R290 is caused by a
high vapour heat capacity. In combination with
the need for keeping the refrigerant charge close
to maximum possible in the system, thus giving
no superheat at evaporator outlet, the suction
line heat exchanger has to be very efficient for
preventing air humidity condensation on the
suction tube. In many cases an elongation of
the suction line and capillary gives efficiency
improvements.
The capillary itself has to be in good heat
exchanging contact with the suction line for as
long a part of total length as possible.
At high superheat, with good internal heat
exchange, the theoretical COP of R290, R600a
and R134a is higher than for R22. At very low
superheat the COP of R290, R600a and R134a is
lower than for R22. The R290 behaviour is similar
to R134a, with respect to internal heat exchange."

Jeff, that is absolutely wrong and is one of the worst generalizations I have heard in a long time. And this argument about scientists is one reason why the climate change deniers will never believe them.

YOU HAVE TO SEPARATE OUT THE SCIENTISTS AND ENGINEERS ABOUT WHOM YOU SPEAK. YOU HAVE JUST, POTENTIALLY, INSULTED MY WIFE WHO FIGHTS DAY IN AND DAY OUT TO KEEP BIAS OUT OF PUBLISHED PAPERS. (take that with a grain of salt)

Distinguish between proper wide spread peer reviewed papers by people with good reputations and the corporate apes that masquerade as scientists. They are the ones with the usually "undeclared" bias.

That said, everyone has a bias, and most scientists do some research to prove a theory. Their neck is on the line if they fake anything. It will almost NEVER be published in a legitimate paper.

The above paper in BBP post, if you read it, is by engineers at a large Indian university. They are not being paid by some nameless nebulous think tank with deep pockets filled from beyond the big curtain.

I'm not talking about the bagillions of tons of institutional research that gets properly reviewed and edited, and ends up being published in some article buried in a trade magazine no one reads. I estimate 99.7% of the university research papers are of this type. Relevant, who's to say? Most of them remain unknown.

What I'm talking about is the papers written by individuals employed by a certain industry or interested or vested entity. Some of these authors are in positions sheltered from repercussions, but most aren't. Logical papers can be written, that stand up to institutional scrutiny, with research to back the thesis, to support nearly any agenda. After a while in the system, ambitious professionals have learned what they can write about that will get noticed, and what not to write about that might end their careers. The peer review process is a walk in the park at the end.

The whole process reminded me of an episode of mythbusters while I was in school. Some of the stuff good researchers were spending their public-funded time on was just a joke. But it wasn't.

Have you been through it?

Not the entire process. I was a helpers helper, working on a small part of a larger project. At least half of what I did got rejected. It was something to do with communication electronics. I was etching and drilling PC boards.

From my research: Propane (R290) performs best with very high Superheat (33-35*C) at the compressor input.

Attaining this much Superheat without a (SLHX, IHX, LSHX) many names, would require that the TXV throttle down the refrigerant flow so that the last section of the Evaporator would not be evaporating any more refrigerant, but just be adding sensible heat.

This throttling down of refrigerant reduces Mass-flow in system, thus capicity drops.

It seems that attaining this much superheat at the compressor input should be accomplished with a SLHX.

Another great benifit in performance, is increasing Sub-cooling at the TXV, so bubbles are elemonated.

Finding SLHX optimum size testing has been difficult, I have seen examples from 1/3 as big as the condenser on down. ?


http://inpressco.com/wp-content/uplo...r48226-229.pdf

" Use of liquid-suction heat exchanger in the VCRS reduces the possibility of liquid
carry-over from the evaporator which could harm the
compressor, by superheating the refrigerant vapours after
evaporator, by the heat of hot liquid refrigerant available
after condensation. The refrigeration capacity of the
system is increased by 9% to 32% for different flow rates
of the refrigerant in the range of 12 to 16 LPH, which is
obtained due to the sub cooling of the liquid refrigerant in
the LSHX, leading to the efficient throttling. For lower
flow rates of the refrigerant the effectiveness of the LSHX
is maximum as compared to higher flow rates, because the
time available to exchange heat is more in case of the low
flow rate. The optimum effectiveness (0.65) is obtained at
40 °C of initial water temperature. Use of LSHX in the
vapour compression refrigeration system, increases the
COP of the refrigeration unit for different initial
temperatures of water by 10 % to 40 % . The compressor
work gets reduced by 3% to 11%, by using the liquid
suction heat exchanger in the vapour compression
refrigeration system that means less power consumption
by the refrigeration unit has been observed. The LSHX
sub-cools the liquid refrigerant before passing to the
expansion valve which leads to the rise in evaporator
capacity which is much more than the heat rejected in sub-cooling.
Thus the LSHX is found to be a device which is
useful for avoiding flashing of refrigerant in expansion
device that raises the amount of liquid refrigerant in
evaporator. Ultimately, it increases the refrigeration effect
and COP of the system (10 % to 40 %). The optimum
effect of the LSHX is seen at different flow rates of the
refrigerant for lower initial temperature of water (30 to 35
°C)."

The tiny amount of testing I have done has me sold on them with r134 and propane in cold climate heating applications. The unit I tried one out in put more heat inside at the same outdoor temperature and added a few more degrees to the bottom end of the useful range. How much, I can't say for sure, but it definitely did a better job than before it was modified.

Has anyone seen inside the newer Zoneline heat pump PTAC units? I saw a brand new one at work today. They have a quiggle-ator in them! This device not only serves as a suction/liquid heat exchanger, but also as a charge compensator. It looks like a compressor muffler, and has cap tubes in the top and bottom of it. The wide part of the muffler is brazed to the outdoor hx line that leads to the reversing valve. I found a faraway pic of a smaller unit on display:

http://www.alpinehomeair.com/partner...2D15_500_0.jpg

If I can remember, and find some time, I'll go back and stare at the rat's nest between the compressor and outdoor hx. I'll take some pics and try to untangle the maze.

These units also have a heat pipe sandwich with the indoor hx in the middle. They call it "super-dry" or something similar. The main purpose of this rig is to increase dehumidification.

Since one half of the heat pipe sandwich is upwind from the evaporator, it will always be warmer than the other half downwind. The heat pipes have R410A in them, and are connected at the top and bottom. When the unit is working as an air conditioner, the downwind hx condenses refrigerant, due to the temperature drop of the evaporator coil. This forces a pressure drop which carries over to the upwind hx. Due to the decreased pressure, the refrigerant boils out of the upstream hx and pre-cools the ambient air before it reaches the evaporator coil.

GE says this gimmick nearly doubles the amount of water the evap coil can condense from the air. They claim a 10 degree or better reduction of wet bulb temperature at the evaporator coil surface.

Now correct me if I'm wrong, but this sounds like the opposite of what we are trying to do with a suction line exchanger. Doesn't this thing shunt some of the load around the evaporator? Also, wouldn't forcing a 10 degree drop in evaporator temperature raise the effective temperature split and kill efficiency and COP?

Follow up video:

http://vid1326.photobucket.com/album...psds07dpdt.mp4

I stood and looked at it up close for a few minutes (until someone ran me off) and I can't be 100% sure how this thing is plumbed. There are no active metering devices, all metering is done with cap tubes. I believe some of them are shunted in cooling mode.

http://www.danfoss.com/NR/rdonlyres/...ompressors.pdf

r290 pressure level and critical temperature are
almost like R22. However, the discharge
temperature is much lower. This gives the
opportunity to work at higher pressure ratios,
means lower evaporating temperatures, or at
higher suction gas temperatures.
Refrigerant R290 is used with polyolester oil in
Danfoss compressors, so material compatibility
is almost identical to R134a or R404A situation
from oil side. R290 is chemically inactive in
refrigeration circuits, so no specific problems
should occur there. Solubility with ester oil
is good. Direct material compatibility is less
problematic.

To keep the
refrigerant flow speed within the recommended
range of 3 to 5 m/s it may be necessary to adopt
the cross flow sections

Special care has to be taken when
designing the accumulator in the system. When
using R22 or R134a the refrigerant is heavier than
the oil used, while with R290 the refrigerant
is less heavy, as can be seen in the data table 1.
This can lead to oil accumulation if the
accumulator is too large, especially too high,
and has a flow path which does not guarantee
emptying sufficiently during startup phase of the
system.

the suction
line heat exchanger is very important for system
energy efficiency of R290, which it was not for
R22, see fig. 5. The figure shows increase of COP
with superheat from few K up to +32 °C return
gas temperature, where a range from +20 °C
to approx. +32 °C is usual for small hermetic
systems.
This large increase in COP for R290 is caused by a
high vapour heat capacity. In combination with
the need for keeping the refrigerant charge close
to maximum possible in the system, thus giving
no superheat at evaporator outlet, the suction
line heat exchanger has to be very efficient for
preventing air humidity condensation on the
suction tube. In many cases an elongation of
the suction line and capillary gives efficiency
improvements.
The capillary itself has to be in good heat
exchanging contact with the suction line for as
long a part of total length as possible.
At high superheat, with good internal heat
exchange, the theoretical COP of R290, R600a
and R134a is higher than for R22. At very low
superheat the COP of R290, R600a and R134a is
lower than for R22. The R290 behaviour is similar
to R134a, with respect to internal heat exchange.

Generally the same rules for evacuation and
processing are valid as for R22, R134a or R404A
systems. The maximum allowable content of non
condensable gases is 1 %.
Too high level of non condensables increases
energy consumption because of higher
condensing temperature and a portion of the
transported gas being inactive. It can additionally
increase flow noise


The 3rd paragraph on Suction Line Accumulator, oil being heavier than Propane & probable accumulation of oil in Accumulator makes me VERY hesitant to use the LARGE 60oz one that I have bought!

I have a large Refrigeration Research BH750 for use as SLHX, 7.5hp (19K btu/h)

The suction line pipe in the middle of it is 15" long x 1.625" ID (31.1 cubic inch)

I'm thinking of positioning it vertical with evaporator output 3/4" pipe into bottom, so it would also be a burp / slobber reciever tank.

This would be a 6.76x increase in pipe volume for 15", thus the refrigerant would slow down by about the same ratio.

Good idea or bad ???