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.
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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.
Dedicated subcooling assumes an external source of cooling for the liquid. What exactly did you have in mind?
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.
Use the incoming water to first sub-cool in the sub-cooler, before passing into the condenser.Quote:
Dedicated subcooling assumes an external source of cooling for the liquid. What exactly did you have in mind?
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)
This can be tuned by adjusting the water flowrate to suit the system.Quote:
What temperature is the incoming water?
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.
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.Quote:
Are the condensers counterflow or crossflow?
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.
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.
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.
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 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).
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.
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
Touche'... You tell me you're not a designer... :D
What types of compressors would this work on?Quote:
As a back-up precaution to safeguard the compressor, I would suggest a crankcase pressure regulating valve (CPR).
How does this work? I've never worked with these yet.
^ 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.
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.
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.Quote:
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.
To make the machines in ultra-compact format is a real challenge - especially given the size of the current evaporator technology (archaic technology).
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?
Scroll compressor - reputable brand.Quote:
What type of compressor are you currently using?
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.
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.
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.
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.
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
Apologies for not replying earlier - I've been away for a few days.
Can you explain more on how the MOP option works? I've seen the option on offer - although not common in my present location.
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.Quote:
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.
Conservative design is absolutely essential, if the compressors are to have a decent lifetime.
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.
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 Magoo,Quote:
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
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.
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.
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.