View Full Version : How to increase pool heater capacity?
A company has installed a heat pump style pool heater for a client who has complained that it just doesnt reach temerature ( 30.5 degrees C).
I am trying to increase the capacity as much as i can without changing any components. The suppliers of the heat pump have said that i should decrease the water flow rate through the heat exchanger? This they say will increase my head pressure and increase the temperature difference across the heat exchanger. 'This would increase my capacity.' While this initially sounded right i now, after thinking about it, disagree. If we lift our HP the compressors capacity decreases. At the moment we have a 47 degreeC HP with a water in of 27 and out of 30.
Shouldnt we rather increase the water flow rate through the heat exchanger , lowering our HP , which is an increase in capacity?. Even though our HP is lower and our TD will be less due to the increased flow rate and the improved compressor capacity will be an overall system improvement?
In other words if I lower my HP from 47 degreesC to 40 DegreesC I increase my compressor capacity from 20kw to 25kw. This i can acheive by increasing the water flow rate which increases the capacity of my water cooled condenser.
Therefore having a Standard 40 DegreeC head pressure with its corrosponding flow rate is better from a heating capacity point of view than a high head pressue with its slower flow rate???:rolleyes:
What compressor (model) are you using?
What is the evaporator saturation pressure/temp?
What is your condenser saturation pressure/temp?
Let's first see what the thermodynamics tells us, before playing too much.
By-the-way, increasing water flow velocity within the allowed pressure-drop generally increases the heat-transfer in the heat-exchanger (condenser), due to an increase in the water-side heat-transfer coefficient, as well as an increase in the log-mean-temp-difference (dTlm).
Slowing down flow purely trades off with a larger dT, water, but reduces dTlm - not where you want to go, generally.
Practically though, the following questions will tell us if the pump is large-enough for the job at hand:
1. Pool volume, or surface area x depth;
2. Days to heat;
3. Covered, or bare?
4. Ambient air temp;
5. Water starting temp.
The pump may just be way too small & then there is precious little you can do without adding in a larger heat-pump.
nike123
16-05-2009, 12:58 PM
At the moment we have a 47 degreeC HP with a water in of 27 and out of 30.I would say that you have dirty condenser, noncondensables in condenser or overcharge since TD of condenser is to high! With that water in temperature your condensation temperature should be 42°C. Check condenser for dirt.
Check system for noncondensables or overcharge.
What is subcooling?
Thanks for the prompt reply.
I have a feeling that we are border line with regard to capacity. However I would like to confirm that their is nothing further i can do to improve performance. The suppliers theory of decreasing the water flow rate does worry me.
The way i understand it from a technicians point of view is that the lower the HP the more the compressors capacity due to the increased refrigerant flow rate. The compressor dictates the system capacity? The lower the HP the more refrigerant flow the higher the capacity.
If I increase the flow rate through the heat exchanger I lower the HP and even though the refrigerant isnt as hot more energy is exchanged?
Therefore even though the TD across the coil is decreased the water change rate of the pool is increased and with the compressors improved capacity the overall capacity is improved?
I would say that you have noncondensables in condenser or overcharge since TD of condenser is to high! With that water in temperature your condensation temperature should be 42°C.
Check system for noncondensables or overcharge.
What is subcooling?
Couldnt the HP be to high due to the suppliers insisting on a 3 degree TD across the coil?If I increase the water flow rate the HP would drop( incresaing the compressors capacity) , the TD would decrease , but overall the system would deliver more heat? Currentley the HP is 47 degreesC , LP is -2 degreesC, 3 degreesC across heat exchanger. Ambient is 10 DegreesC. This is a standard 'off the shelf' pool heat pump coupled to a standard pool pump. I believe the heater is under sized , but would like to understand the theory of how the different Head Pressures and flow rates would influence capacities? Im not sure of the sub cooling.
nike123
16-05-2009, 02:04 PM
Couldnt the HP be to high due to the suppliers insisting on a 3 degree TD across the coil?If I increase the water flow rate the HP would drop( incresaing the compressors capacity) , the TD would decrease , but overall the system would deliver more heat? Currentley the HP is 47 degreesC , LP is -2 degreesC, 3 degreesC across heat exchanger. Ambient is 10 DegreesC. This is a standard 'off the shelf' pool heat pump coupled to a standard pool pump. I believe the heater is under sized , but would like to understand the theory of how the different Head Pressures and flow rates would influence capacities? Im not sure of the sub cooling.
3°C across heat exchanger is not TD. That is condenser secondary side delta t (Δt).
TD is difference between SCT (saturation condensation temperature) and water in temperature and for water condenser should be around 15K.
Condenser (Δt) should be around 5K for best efficiency and it is adjusted by water flow.
3°C across heat exchanger is not TD. That is condenser secondary side delta t (Δt).
TD is difference between SCT (saturation condensation temperature) and water in temperature and for water condenser should be around 15K.
Condenser (Δt) should be around 5K for best efficiency and it is adjusted by water flow.
A decent condenser will have a TD around 10-12'C, if at all possible. If it needs to sit above 15K, then please check the items nike123 suggested.
Perhaps you could refer to the water-side temp rise as dT,w. This is typically around 3-5'C.
To the OP, please don't forget to check that the supplied rating of the heat-pump is actually what the pool requires. Heat-pump sizing needs to be correct.
3°C across heat exchanger is not TD. That is condenser secondary side delta t (Δt).
TD is difference between SCT (saturation condensation temperature) and water in temperature and for water condenser should be around 15K.
Condenser (Δt) should be around 5K for best efficiency and it is adjusted by water flow.
While I agree with the above, there is one more factor to look at: Approach temperature.
The approach temp is the difference between SCT and leaving water, and is an indicator of heat transfer from the refrigerant to the water. Approach temp should be no more than 20F/11K. In this case, the approach is 47 - 30 = 17K/30.6F. This is much too high.
In general, a high approach temp means the interior surfaces of the condenser need to be acid cleaned. But it can also mean that the condenser has been installed incorrectly.
Maximum heat transfer is achieved by running the two fluids (refrigerant and water) through the condenser in opposite directions (counterflow) or across each other (crossflow).
It is possible that the water in and water out connections have been swapped, resulting in the two fluids running in the same direction (parallel flow). Parallel flow is extremely inefficient.
Here is one more thing to think about:
If the water connections are top and bottom, it is good practice to fill the water side of the condenser from the bottom up, pumping water into the bottom and exiting the top. This ensures that water fills every little nook and cranny eliminating any air pockets.
I'm thinking pics of the system might be very helpful here.
nike123
16-05-2009, 07:53 PM
I think he should first measure water flow and calculate achieved duty, and then compare with rated data. That will tell a lot.
laf100
16-05-2009, 09:43 PM
You guys appear to way ahead of me on this one, but surely if we continuously add a certain amount of heat (however small) to a body of water (however large), as long as the heat losses are less that the heat added, it must continue to rise in temp.
However slowly, would eventualy reach setpoint, would it not? Perhaps I am oversimplifying things a bit, but retetion of the added heat must be a factor in this problem?
LAF
:cool:
OK- only after writing this drivel do I notice the OP is from Aus'!!! Maybe my heat loss theory is "blown out of the water" after all!!
:confused:
Thanks guys. I will be using the above info. The flow direction of the water is an interesting one which I havent checked. The unit is new and therefore cannot have scale or a dirty condenser. Thanks for correcting me on the terminology. As a technician and not an engineer i try and keep it simple.
The suppliers of the unit say we should obtain a 3 DegreeC water temperture difference across the coil. When I obtain this the HP is at 47 Degrees. Whether this is a poor design issue or a fault is yet to be proven.The suppliers unfortunately doent seem to know much about their product either.
The way I understand it we can measure the capacity of the unit by the following :
KW= l/s x 4.19 x water temp change
My compressor delivers say 20kw @ (47SCT; -2SST)
If i lower my HP and maintain my suction it then ,on paper delivers 25kw (40 SCT ; -2SST )
If I increase my water flow rate to achieve this lower hp will their be an increase in capacity on the system in theory?
We could therefore go from:??
20kw(47SCT; -2SST) = Lower l/s x 4.19 x (30-27)
25kw (40SCT; -2SST) = Higher l/s x 4.19 x (30-28)
In other words would 25kw @ 40degreesC be more effective than 20kw @ 47degreesC?? with their adjusted flow rates.
At the end of the day we can play with flow rates and temerature difference, but they all balance out unless we increase our KW input.We cannot make energy out of nothing?
If this was an electric heat system and i added 25kw @ 47degreeesC there would be an improvement. Due to the fact that it is a refrigeration system the higher the HP the lower the capacity offered from a heating point of view.
The way i see it is that I have made the condenser more efffeicient by increasing the flow rate, which has lowered the hp which has increased the capacity of the compressor and the system overall?
If this was an air cooled condenser and I increased the air flow over the coil my HP would drop and as long as my TEXV is sized correctley for the pressure drop , I would increase the capacity of the compressor and the system overall.
I understand this might be an increase of only 10% or so , but im tying to make a point with the suppliers that decreasing the flow rate, which increases the HP and has a larger temp difference across the coil is not more effecient than a lower HP , increased flow rate and decreased temp difference.
Understanding this is a pool application which recirculates the water over a few days so the water off temp is not really that critical.
OK- only after writing this drivel do I notice the OP is from Aus'!!! Maybe my heat loss theory is "blown out of the water" after all!!
:confused:
I live in Tasmania which has an ambient average of about 8 DegreesC at the moment.
I do believe the sytem is slightly undersized and is managing to achieve 30degreeC, but no higher. At 47(SCT) the compressor and the pool have reached the
Balance point with heat in equalling heat out. The client only needs 31 Degrees pool temp and that is why I jusy need a little more power and i thought i could achieve this by lowering the HP by increasing water flow rate, increasing the condenser capacity, gaining more Compressor capacity and overall sytem performance.
This as Ive said often, in contrary to the suppliers technical support advice of limiting the flow rate further to increase the HP?
Magoo
17-05-2009, 01:30 AM
Tassi is on a similar latitude as NZ. So if ambient is at 8 'C , it is time to shut down and winterize system. Even with a thermal pool cover the heat loss from pool , add windage factor, will be beyond system to maintain 31' C the Heat pump will not cope.
The heat pump salesman will say anything to sell product and cloud the real issue.
magoo
http://www.aquaheat.co.za/pages/calculate
Please use the above pool heat calculator to determine the 'correct' size of heat-pump required.
If your current heat-pump is new & won't reach duty, please have your client purchase additional capacity, or trade it in for the correct size unit.
Another think we need to check is the subcooling. Subcooling is the difference between SCT and liquid line temp at the receiver outlet (or condenser outlet if there is no receiver). The subcooling should be no more than 15F/8.5K.
High subcooling means excess liquid is backing up into the condenser, reducing the area that is available for active heat transfer. High subcooling is usually caused by overcharge.
This system definitely has a heat transfer problem. We just need to figure out why.
I will check the sub cooling. If the system is designed with a slightly small heat exchanger the head pressure will balance at a higher pressure limiting the capacity of the system. If i increase the flow rate, which i can, by adjusting the bypass valve, i will increase the capacity of the system. ?
To put things a little bluntly - if you don't first determine the correct heat-load requirements & confirm that the heat-pump is the correct size, then you are really just wasting your time. Honestly.
Whether the dT,w is low/high, with water flow high/low makes little real difference - it is fine-tuning, where you're playing around within the condenser itself.
The thing you have to focus on is what is happening to the actual water in the pool - not the heat-pump. If the water is not getting hot (in new condition), it is in all likelihood (95%), undersized. If that is the case, then you are wasting your time searching within the heat-pump itself for your answers.
If the thing is not working in new condition, what on earth will it do when it is fouled?
Given the current heat transfer, coupled with the low delta-T, it appears that you already have a heavy flow rate and have reached a point of diminishing returns. Little can be gained by increasing the flow rate.
Once the heat transfer is increased, however, this will raise the leaving water temp, which will raise the delta-T. At that point, increasing the flow will gain more capacity.
I think you are right. What im trying to do is fine tune a system that doesnt really have the power no matter what i do. It is just the theory that interests me and the annoying fact that the suppliers reacon that it is a nominal 30kw , but no where on the compressors capacity curves can i find that sort of capacity. The most i can find is 25kw @ 0 SST/40SCT. ( This includes the power consumption)
I'll never get my suction above 0 with ambients like what we get in Tassie. I could probably get a little extra if i lower my HP , but i dont think it will be enough. I will check the water side for contra flow and ill check my subcooling. Thanks for the link to the heat load calc for pools.
The pool is an indoor pool for a swimming class and so it is used all year round.
nike123
17-05-2009, 07:26 AM
I think, before anything, you should establish that this heat pump delivers its rated capacity and works as it should be working in given conditions. If this heat pump doesn't deliver its capacity than it must be found what is wrong and then act accordingly. What is the point in flow adjustment and pool capacity calculation if heat pump delivers only 75% of its rated capacity.
So measuring flow and temperature difference should be first thing to do. Otherwise, it is walk in the dark.
Fact that something is new doesn't mean that works as it should. I have seen many equipment coming faulty right from factory (Italian especially) with all quality check tags signed.
I still say that 47°C SCT with 27°C water in is not good and your unit has some fault.
Im not sure of the flow rate, but what i know is that our HP is at 47DegreesC , Suction at -2DegreesC which offers 20 kw according to the compressor manufacturers curves. The temperature difference across the coil is 3 degrees.
Therefore is the following correct:
kw=l/s x 4.19 x Temp difference
Therefore 20 = l/s x 4.19 x 3
l/s = 1,6.
So i can use the compressor capacity charts to obtain capacity of the system or use the above formula.
I was hoping that i could solve this issue by lowering the hp and gaining 5kw which could have saved the day. I will make sure the mechanical side is working correctley . The correct heat load for the pool is 30 kw and the company that sold her this unit stated that it is 30 kw. Unfortunately i can only find 20 and was hoping to squeeze out another 5 by fine tuning.
One of the main problems with heat-pump marketing, is at what temperature the performance & COP figures are quoted.
Heat-pumps start off with tremendous performance & COP values in the startup (cold) condition, but these both decrease as the pool-water temperature rises.
It could be that the 30kW cited is a start-up condition & not the terminal (hot) condition, or even the range midpoint.
What compressor is used?
With regard to the installation the unit is raised about a meter off the ground . I need to check whether the water fills from the top or the bottom. Ideally we should be filling from the bottom with contra flow refrigerant. If it filled from the top or some how has an air pocket in the heat exchanger i could get the high hp with poor temp difference.
Once again if I do the KW = l/s x 4.19 x temp diff
This will tell me what the unit is delivering. We can compare this to the compressor capacity curves and then confirm whether it is working to the specs. if not find the fault. if it is then the suppliers need to pass over a true 30 kw as spec,
nike123
17-05-2009, 09:19 AM
Once again if I do the KW = l/s x 4.19 x temp diff
This will tell me what the unit is delivering. We can compare this to the compressor capacity curves and then confirm whether it is working to the specs. if not find the fault. if it is then the suppliers need to pass over a true 30 kw as spec,
Yep, that is way I preach from post# 10.;)
^ & ^^ Big problem then...
What if your water temperature measurements in/out followed a typical accuracy of around 1'C?
Say your cold in read 1'C high & the hot out read 1'C low? You'll end up fighting with the heat-pump supplier until the cows come home. It is very difficult to measure water temps accurately unless you have a well-calibrated test system.
If you're going to measure anything at all, do it on an average pool water value. Plot this against time, to see what the average heat input is over the whole period. You can then trace backwards to see what the heat pump is delivering - on average.
Thanks everyone for your help on this. I will let you know indue course of what i find. One thing is for sure ive learnt a lot !
Thanks again. Refrigeration is truely an international language and a challenge.
^ All the very best. I'm sure you'll get it all sorted out.
Hey desA
I sent you a private message. i hope you get it because im not sure if i sent it right.
^ Thanks very much, Drew. Received & answered.
Alles van die beste, kerel.
Therefore is the following correct:
kw=l/s x 4.19 x Temp difference
Therefore 20 = l/s x 4.19 x 3
l/s = 1,6.
Nike stated earlier that the SCT should be 42C with a delta-T of 5K, so let's give those numbers a try:
1.6 x 4.19 x 5 = 33.52 KW (as opposed to the 20 KW we are currently getting)
Are we wasting our time fine tuning the system?... or are we looking for a serious capacity-killing heat transfer problem, which the approach temperature tells us exists?
Thats what i will try and find out today Gary. Thyanks for the input.I'll chech the sub cooling , superheat, reclaim and recharge, chech flow direction, all after confirming what capacity the unit is delivering
Just to keep those interested in whats happening up to scratch:
The suppliers of the unit went to site today with their own technician (without me) and lowered the water flow rate through the heat exchanger, lifting the temperature difference across the coil to 6 C(36C off and 30C on) which lifted the SCT to 52 DegreeC.
Coincedentley after a warmish day today the unit then reached its 30.5C:mad:
Their was lots of back slapping and congratulations all around with a happy client.
However tonight temperatures drop to 2 C and we will see what the pool temp is tommorrow.
Raising the SCT by 1C lowers compressor performance by 1% (so Ive read) so by doing what they have has already lowered system performance by 2%.
I will have to wait a few days to prove this as the water temp in the pool takes a while to change.
Be intrigued to see what has happened to the unit COP.
The suppliers of the unit went to site today with their own technician (without me) and lowered the water flow rate through the heat exchanger, lifting the temperature difference across the coil to 6 C(36C off and 30C on) which lifted the SCT to 52 DegreeC.
So, let's see where we are:
Mean water temp = (30+36)/2 = 33'C
SCT = 52'C
Mean approach temp = 52 - 33 = 19'C !!!
I'm starting to lean towards Gary & Nike's views. Something does not add up nicely wrt the condenser.
Is the system overcharged?
nike123
18-05-2009, 09:58 AM
Just to keep those interested in whats happening up to scratch:
The suppliers of the unit went to site today with their own technician (without me) and lowered the water flow rate through the heat exchanger, lifting the temperature difference across the coil to 6 C(36C off and 30C on) which lifted the SCT to 52 DegreeC.
What type of heat exchanger is in that unit?
Plate or shell and tube or else?
What is unit make and model?
Since you where there, did you checked that flow and return pipe of pool water are correctly connected at unit at their respective connections. Did you checked that maybe connections on unit are tagged wrongly in regard of counter flow of two media at heat exchanger?
We are all going to be very disappointed if you didn't check the water flow direction and the subcooling.
I will get to the bottom of this. (and let you all know)
The unit is labelled correctley with regard to water flow and the water pipe connections are correct.
What I noticed on the evaporator is that their seems to be a few heat pump model numbers on it which indicates 'one size fits all' models? I hope it isnt the same with the titanium heat exchanger. There is no point in having a compressor that delivers 25kw on paper that is dragged down with a small evap and condenser. Still need to get to site to check subcooling.
Hi Guys,
I still havnt been to site , but as expected , with the low ambients, the pool lost temp overnight.
My boss, did however, notice what i think could save the day. I realise we wont get the 30 kw as spec, but could increase it by quite a bit. Also the undersized, inefficient (or what ever) heat exchager isnt great. The manufacturers arent interested.
This unit was installed in place of an older one. The water pipes run into a wall and then undergound!!!
The water leaves the exchanger at say 30C , but only arrives into the pool at 28.5C. We have installed a carel thermometer in the water in and out side of the heat exchanger so the reading leaving the heat exchanger should be accurate and I hope to confirm this myself. This is also why the actual reading into the pool hasnt been worried about. We've been to focused on off and on coil temp.
Would this be true:
l/s = maybe worst @ 1 ( due to suppliers throttling water to obtain 6C across coil)
Temp difference down water pipe = 1.5C
Kw=l/s x 4.19 x temp difference down pipe.
= 1 x 4.19 x 1.5
= 6.2kw
Now, if I can assume, the heat loss to be equal on the return pipe.
Therefore 12kw???
Can this be right?
That would be a big chunk of capacity out of the unit that is already handicapped?
I'm thinking we need to base our conclusions upon measurements and not assumptions.
Another thing to think about: If the water is losing heat underground, then the slower it moves through the pipe the more heat will be lost. Another reason to increase the flow.
We haven't even looked at the low side of the system yet. The low side has similar indicators: delta-T, SST (saturated suction temp), TD, evap approach, superheat, etc.
The more heat in (low side), the more heat out (high side).
This has been a fustrating job. The client has had 2 x 6kw inline electric heaters installed. :mad:
Therefore with one underspec unit with components that have dragged the compressors capacity lower and possibly heat loss underground the poor client has had to make this call. The suppliers of the unit still say that it is a 30 kw(nom) , but the client is loosing business and had to do what she did.
I might never know what the fault(s) is (are).
Thanks again for all your help and sorry to have wasted your time. I have learnt a lot which hopefully I can apply to the next job.
If I were to state that at least one of the more well-known names in air-to-water heat-pumps has a make-up electrical element heater installed 'inside the heat-pump unit' itself - would that surprise you?
Its a shame we won't be able to get to the bottom of this. It could be something as simple as too much refrigerant in the system.
yinmorrison
21-05-2009, 10:08 PM
Or if you go back to desA's original post was it sized correctly in the first place.Best to do a general rule of thumb before going to a whole lot of bother on this type of system.
Yes it was sized correctley. A 24kw unit used to do the job. The client wanted to go up a size so a 30kw was installed. This "30kw" doesnt obtain temps that the 24kw obtained.
Yes it was sized correctley. A 24kw unit used to do the job. The client wanted to go up a size so a 30kw was installed. This "30kw" doesnt obtain temps that the 24kw obtained.
Were these two heat pumps from the same manufacturer?
If so, then something else does seem to be amiss. If same manufacturer, did their technology change in the interim?
Couldnt the HP be to high due to the suppliers insisting on a 3 degree TD across the coil?If I increase the water flow rate the HP would drop( incresaing the compressors capacity) , the TD would decrease , but overall the system would deliver more heat?
After this thread & the ongoing discussions, I decided to develop an equation which describes how a generic condenser operates under given parameter changes.
What seems to come from this is the following:
- decrease water flowrate;
- water outlet temp rises;
- condenser saturation temp required reduces.
- for the SAME heat-transfer rate (kW)
Basically, unless the change in water flowrate drastically alters the water-side heat-transfer coefficient in the condenser, no net additional heat-transfer is effected by tweaking the water flow-rate - it is an illusion.
The equation for water heat-balance is:
q' = m'w*Cpw*dTw
where :
q' = heat-transfered to the water [W]
m'w = water mass flowrate [kg/s]
Cpw = water specific heat [J/kg.K]
dTw = Tw,o - Tw,i
Tw,o = water outlet temp ['C]
Tw,i = water inlet temp ['C]
For a fixed heat-transfer (q'), the m'w & dTw values merely trade off against each other.
So, by having a hotter Tw,o at lower flowrate, no additional heat is actually transferred - it is an illusion, I'm afraid.
Yup, thats the way it is. You cannot create energy out of nothing.:D
If the heat transfer remained the same then the high side pressure would remain the same. The fact that the high side pressure increases says that lowering the water flow decreases the heat transfer.
If the heat transfer remained the same then the high side pressure would remain the same.
The governing equation for a condenser seems to show that under same heat-transfer condition, as stated above, that the Tsat of the condenser will actually drop slightly, when flow is reduced & Tw,out increased.
The system floats on the inlet temperature of the external cooling medium. This also presumes no fundamental heat-transfer change in that process.
I'd be very interested in why the high side pressure should remain the same, if the Tsat for equilibrium is to reduce slightly - I'd expect it, under equilibrium conditions, to reduce slightly in line with the reducing Tsat. This is in the order of a few degrees celsius, but it is predicted.
:)
When an actual (accurate) measurement shows one result and a calculation shows another result, there is a flaw in the calculation.
A measurement beats a calculation every time.
Yuri B.
26-05-2009, 06:50 PM
Drew might have achieved greater "q" due to decrease of "m'v" being less than increase of "dTw".
Yuri B.
26-05-2009, 06:57 PM
Dew might have achived greater "q" due to decrease of "m'w" being less than increase of "dTw".
Drew might have achieved greater "q" due to decrease of "m'v" being less than increase of "dTw".
Since q in fact decreased, as evidenced by the increase in high side pressure, I'm thinking that you have this backwards, but you are on the right track.
What we are missing here is the measured values of mw. A decrease in q tells us that the value of mw decreased faster than the value of dTw increased.
The governing equation for a condenser seems to show that under same heat-transfer condition, as stated above, that the Tsat of the condenser will actually drop slightly, when flow is reduced & Tw,out increased.
The system floats on the inlet temperature of the external cooling medium. This also presumes no fundamental heat-transfer change in that process.
I'd be very interested in why the high side pressure should remain the same, if the Tsat for equilibrium is to reduce slightly - I'd expect it, under equilibrium conditions, to reduce slightly in line with the reducing Tsat. This is in the order of a few degrees celsius, but it is predicted.
:)
The flaw in all this is the assumption of fixed heat transfer. The heat transfer is not fixed.
At this point I would strongly suspect that the system is overcharged, with excess liquid refrigerant backing up into the condenser, limiting it's ability to transfer heat. A simple subcooling measurement would confirm or deny this.
The flaw in all this is the assumption of fixed heat transfer. The heat transfer is not fixed.
Thanks for that. I'm very interested in understanding why the heat-transfer should be variable, & the high-pressure remain constant.
The equation, of course, also allows for a changing heat-transfer. I'm interested in the physics/thermodynamics & experimental observations for the reasoning behind this.
I bow to your many years of experience in this - no contest - & would be interested in why this occurs.
I've resolved the governing equation in terms of heat-transfer as a function of the remaining variables. The water mass flowrate appears in two places:
1. Linear increasing term;
2. (1-exp(-a)) term;
This shows a competition between two scenarios for heat-transfer as a function of water mass-flow. I'll plot it out.
This has got me intrigued. I'll scurry off to my laboratory heat-pump & test out the real system response over a number of scenarios. It's actually an age-old chestnut & I'd like to get completely to the bottom of it.
:eek:
Thanks for that. I'm very interested in understanding why the heat-transfer should be variable, & the high-pressure remain constant.
Possibly I was not clear in this. You were saying that the heat transfer must remain constant... and I was responding that if the heat transfer were constant, then the high side pressure would be constant. In fact, neither is constant.
Possibly I was not clear in this. You were saying that the heat transfer must remain constant... and I was responding that if the heat transfer were constant, then the high side pressure would be constant. In fact, neither is constant.
Fair enough. :D
Can I press a little further in, if I may be so bold...
(1) Why is neither constant?
(2) How are heat-transfer & high-side pressure related?
In a system, some component interaction around the refrigeration loop can sometimes showcase some very interesting effects.
The pressure of refrigerant is directly related to its saturation temperature. As the saturation temp rises, the pressure rises. As the saturation temp falls, the pressure falls.
The purpose of the condenser is to transfer heat from the refrigerant to the water. If heat is transferred faster (increased q), the refrigerant is cooled. In other words its saturation temperature falls, as does its pressure.
If the rate of heat transfer is less (reduced q) the refrigerant retains more heat and its saturation temp rises, as does its pressure.
In this case the suppliers decreased the water flow. As a result, the pressure increased, therefore the saturation temp increased, which tells us there is less heat transfer (lower q).
If you factor lowered q into your formula, you will see that water flow and delta-T cannot be an equal trade-off. In order for q to be lower, the decrease in water flow must exceed the increase in dT.
Decreasing the water flow was counter-productive.
^ Thanks very much for your explanation, Gary. :)
Yes , the fact that the high side pressure increases says that lowering the water flow decreases the efficiency of the heat exchanger forcing the compressor to pump to a higher pressure to condense . With the lowered refrigerant flow rate the compressor and system balance at a new point with adjusted THR and system performance.
And visa versa, increasing flow rate would create a more efficient heat exchanger and would lower the head pressure increasing capacity.
This has got me intrigued. I'll scurry off to my laboratory heat-pump & test out the real system response over a number of scenarios.
:eek:
ha ha thats funny. I can just picture it!:D
Heat absorbed = Heat transfered + heat of compression = heat rejected
If we alter the 'heat transfered + heat of compression' part of the formula (compressor capacity), by altering the operating pressures the heat 'in' would still equal the heat 'out'.
The compressor will deliver a certain capacity based on the conditions that are created by the 2 heat exchangers. As the oncoil temp increases, the head pressure rises. As the head pressure rises the compressors capacity decreases.As the compressors capacity decreases there is less energy to reject . The compressor and condenser balance at a new pressure which equates to the capacity.The three components individually now have this capacity which becomes the changed system capacity. The system is dynamic.
Again
Heat absorbed (kw , evap)= heat transered + heat of compression (kw ,compressor) = heat rejected (kw, cond)
In this case the suppliers decreased the water flow. As a result, the pressure increased, therefore the saturation temp increased, which tells us there is less heat transfer (lower q).
Decreasing the water flow was counter-productive.
Decreasing the flow rate was counter productive because it forced the compressor to a higher SCT and at this higher point the compressor has less capacity so less heat has to be rejected by the condenser. This is why there is less heat transfer?
What seems to come from this is the following:
- decrease water flowrate;
- water outlet temp rises;
- condenser saturation temp required reduces.
- for the SAME heat-transfer rate (kW)
I do apologise - been re-reading my typing here (had a niggling thought today, thought I'd check) & checking my simulation - a typing error had crept in. It should have read as follows:
- decrease water flowrate;
- water outlet temp rises;
- condenser saturation temp required increases.
- for the SAME heat-transfer rate (kW)
This would now seem to agree with what followed my original comment. Sorry about the confusion.
Following on the Tc,sat change should be an appropriate Pc,sat change.
For small flow-rate changes, the change in equilibrium Tc,sat is not too much, but at very low flows, the Tc,sat can rise substantially. I've actually seen this effect in my lab heat-pump, when the water flow is backed off hard, or stopped. The pump then trips on the high-pressure safety switch fairly promptly.
Practically, as Gary stated, the actual heat-transfer in the condenser most probably doesn't remain completely constant, but the analysis offers useful information nevertheless.
- for the SAME heat-transfer rate (kW)
I might be loosing the plot here, but just to make sure I understand: If the head pressure climbs the compressor delivers less energy, the heat exchanger will adjust to the new capacity and therefore exchange less kw?
I might be loosing the plot here, but just to make sure I understand: If the head pressure climbs the compressor delivers less energy, the heat exchanger will adjust to the new capacity and therefore exchange less kw?
Will the compressor deliver less energy, or reset itself to a higher energy level, with consequent higher delivery to the condenser?
If this is correct, then surely the condenser output would rise & the heat-pump folks would have been correct all along?
:confused:
If you look at a compressor capacity chart you will see that the higher the saturated condensing temp the less the capacity . The lower the suction the less the capacity. To obtain the best capacity from a compressor we keep the suction as high as possible and the head as low. These pressures are governed by ambients and design product temp. Within reason we try to stick within these parameters. Its also a matter of cost and rules of thumb.
The reason that the capacity is reduced with high head pressure and low suction is due to the reduced refrigerant flow rate with these pressures. With a reciprocating compressor the piston must never hit the valve plate so we have a gap (clearance volume) as a saftey. On the piston downstroke the refrigerant in the clearance, which is actually the same as the head pressure,expands before new refrigerant can be drawn in. If the pressure is high a lot of piston downstroke is wasted before drawing in new refrigerant. If the suction is low it is even worse. Not only is a lot of the downstroke wasted overcoming the high head pressure , but the suction valve will open even later. These issues reduce refrigerant flow through the system. The lower the flow rate the lower the capacity.
Ive read that on average 1kg/hr of R22 will absorb about 50w of heat at the evaporator. If we reduce this flow rate the capacity will decrease.
If you look at a compressor capacity chart you will see that the higher the saturated condensing temp the less the capacity . The lower the suction the less the capacity. To obtain the best capacity from a compressor we keep the suction as high as possible and the head as low. These pressures are governed by ambients and design product temp. Within reason we try to stick within these parameters. Its also a matter of cost and rules of thumb.
For condenser heat output, at increased hp temp/pressure, you're correct in that the total heat output should reduce slightly. This is true. So, it should reduce then... fair-enough, on that logic.
Sidebar:
The compressor input is predicted to rise, though - not fall with rising Tsat,cond (at same Tsat,evap).
For a heat-pump, as the Tsat,cond rises, the percentage of the total condenser output, moves from being evaporator-dominated, to compressor-dominated. Take a Copeland chart & try this. Simulators predict the same thing.
Downside:
The problem with only looking at one aspect of the system like this is that we've forgotten about the other parts. Once the condenser stabilises, does the evaporator re-adjust itself to a new equilibrium?
Practical note:
One of the problems with adjusting water flow in such a way as to raise Tsat,cond, is that it becomes a problem to reach water temps of more than 60'C, since this can push compressors above their recommended safe operating range (65'C shown in Copeland US data-sheets, 75'C in European data sheets). For R-134a it becomes a real balancing act up at that temp range, I can tell you. I try, if possible, to keep as low an approach within the condenser, as possible, to compensate for this problem.
In a pool-heater this is generally not an issue, as you still have ample temp margin available - but, for an AWHP, this can become an issue.
Heat absorbed = Heat transfered + heat of compression = heat rejected
Using your formula:
If the heat rejected is less, then the heat absorbed must be less and the heat transferred must be less.
Which of these factors is the controlling factor? None of them and all of them. A change in any part of the system is a change in every part of the system.
Well, this has indeed been a fascinating thread. I've certainly learned a great deal.
In order to properly evaluate a system, we need to view it as a heat transfer chain, with the goal being to identify and strengthen the weakest link.
For example: If we reduce the airflow through an evaporator, the slower moving air has more time to cool as it passes through the coil, thus the delta-T increases. At the same time, the SST has decreased, verifying that less heat is being transferred due to the reduced air volume.
Assuming a fixed speed compressor, this tells us that an increase in delta-T must indicate reduced airflow and that reduced airflow = reduced heat transfer. In other words, high delta-T is an indicator of airflow problems.
There are similar indicators for each link in our heat transfer chain.
Sidebar:
The compressor input is predicted to rise, though - not fall with rising Tsat,cond (at same Tsat,evap).
the rise in input doesnt out do the loss in refrigeration effect and the result is a loss.
In order to properly evaluate a system, we need to view it as a heat transfer chain, with the goal being to identify and strengthen the weakest link.
Well put.
It just takes one undersized component to drag the system down to a less effeicient capacity.
Sidebar:
The compressor input is predicted to rise, though - not fall with rising Tsat,cond (at same Tsat,evap).
In fact, rising Tsat,cond is generally accompanied by rising Tsat,evap... yet we know there is a drop in capacity. The loss must therefore be in refrigerant mass flow.
Ran up a test on the air-to-water heat-pump in my lab today.
I let the system stabilise at a full-flow condition on the water side, for around 30 minutes & recorded system readings. Then reduced the water flowrate substantially, let system stabilise to same tank water temp (allowed 45 minutes) & took new readings.
The trends observed were at follows:
1. Tsat,con rose from 48'C to 53.8'C on gauge (+5.8'C);
2. Psat,con rose from 1150 kPa to 1350 kPa on gauge (+200kPa);
3. Tsat,evap rose 0.5'C (within experimental error)(on gauge);
4. Compressor amperage rose 12.2% (actual recorded, marginally different from Copeland data sheet);
5. dT,water rose from 1'C to 5.7'C (poor sensor placement - indicator only)
6. First test : Tw,in=38.0'C ; Tw,out=39.0'C ; Tc,sup=60.2'C; Tc,exit=43.6'C ; Tw,tank=37.1'C (sensor position & mixing); Tevap,sup=23.4'C;
7. Second test : Tw,in=38.8'C ; Tw,out=44.5'C ; Tc,sup=65.8'C ; Tc,exit=48.7'C ; Tw,tank=38.2'C (sensor position & mixing); Tevap,sup=22.6'C;
When calculations & curve-fits are performed against the Copeland Europe predictions in 'Select 7.1', the following information emerges:
a. Condenser output load reduces (-3.06%);
b. Compressor input power increases (+14.522%);
c. COP,hp reduces substantially (-15.355%)
Air on/off evaporator coil remained similar in both tests (performed in same test run). The slight difference in recorded compressor power increase & Copeland predicted values is attributed to a system refrigerant charge still slightly under the optimum conditions. Optimum charge determination is still in progress for this heat-pump.
The approach temp seems to have remained relatively stable (Tsat and water out temp rising in unison), as I would expect, but not knowing the actual Tsat and leaving water temp I am wondering what your approach temp is.
The trends observed were at follows:
1. Tsat,con rose from 48'C to 53.8'C on gauge (+5.8'C);
2. Psat,con rose from 1150 kPa to 1350 kPa on gauge (+200kPa);
3. Tsat,evap rose 0.5'C (within experimental error)(on gauge);
4. Compressor amperage rose 12.2% (actual recorded, marginally different from Copeland data sheet);
5. dT,water rose from 1'C to 5.7'C (poor sensor placement - indicator only)
6. First test : Tw,in=38.0'C ; Tw,out=39.0'C ; Tc,sup=60.2'C; Tc,exit=43.6'C ; Tw,tank=37.1'C (sensor position & mixing); Tevap,sup=23.4'C;
7. Second test : Tw,in=38.8'C ; Tw,out=44.5'C ; Tc,sup=65.8'C ; Tc,exit=48.7'C ; Tw,tank=38.2'C (sensor position & mixing); Tevap,sup=22.6'C;
There we are, Gary - I've added in the missing data you need.
I'll add in a few more items, if you'd like to crack it open - I'd love to hear your thoughts.
The approach was 9K in the first run and 9.3K in the second run. It should be no more than 11K, so you are getting good heat transfer. Contrast this to the 17K approach in Drew's pool heater and you see how very sick the pool heater is.
Note that adjusting the water flow has very little effect upon the approach temp. It is an entirely different issue.
There we are, Gary - I've added in the missing data you need.
I'll add in a few more items, if you'd like to crack it open - I'd love to hear your thoughts.
First we would need a better description of your system. Is it a cap tube or TXV system? Is it designed for comfort cooling and/or heating or for water heating?
Any further information you can provide about the design/purpose might prove to be helpful.
There we are, Gary - I've added in the missing data you need.
I'll add in a few more items, if you'd like to crack it open - I'd love to hear your thoughts.
Also, if we are to evaluate your system, you should start a new thread for that purpose.
As Gary said ealier a raised head pressure would normally cause an increase in evap pressure due to the reduced refrigerant flow. If they did increase proportionally the capacity loss according to compressor capacity charts wouldnt be much. But as we see with DesA 's experiment under reduced flow rate which forced the hp up the suction didnt climb as much. Therefore kW output would be reduced?
With the sick pool heater that I was working on it eventually had a Hp of 52C and Suction of -5C (due to low ambients)
If you knew the flow rates of the water we could see the two different Kw outputs of your system?
This would explain a lot.
kw = l/s x 4.19 x temp diff.
If we could work out the refrigerant flow rate we could see this transfer taking place through the chain.
We could also follow the heat tranfer chain to the evap where the heat gained in the refrigerant will equal the heat transfered through the heat excahanger and will equal the heat lost from the air.
As Gary said ealier a raised head pressure would normally cause an increase in evap pressure due to the reduced refrigerant flow.
Hmmm... on second thought this would be a result of warmer liquid entering the coil and the refrigerant mass flow should actually increase. Part of the heat is simply recirculating.
Excellent feedback.
The exit approach temp is a very interesting indicator. Thanks for that, Gary.
I'll sort out the water flow-meter in the next few days, so that more accurate water measurements are available (beats bucket & timer).
As Gary suggested, I'll then start a new thread so that we can dissect the system in more detail.
With the sick pool heater that I was working on it eventually had a Hp of 52C and Suction of -5C (due to low ambients)
Ahah... -5'C suction temp. :eek:
What refrigerant was it using? R-22, or R-134a?
Added later:
Currentley the HP is 47 degreesC , LP is -2 degreesC, 3 degreesC across heat exchanger. Ambient is 10 DegreesC.
I missed that in the earlier post. With an ambient like that, you are on a hiding to nothing, with an air-to-water pool heat-pump. If the ambient goes any lower, the defrost mechanism will probably begin to kick in - if it's equipped with one (some aren't in Australia).
Magoo
30-05-2009, 03:26 AM
Drew . Sell the owner a spar pool and throw all the heat into it, you are on a mission to no were with the swimming pool. Given the reduced heat pick up from lower ambient, and losses from pool overnight.
Do the sums, volume of water in a ground pool, loses to ground, add losses to atmosphere even with a cover, thermal separation temps. Add reduced h/p capacity with low ambient.
everyone will theorise till the cows come home, but nothing will change.
magoo
Hey Magoo
Im over the pool heater. It has been sorted out. Not the way I would have liked , but anyway. This thread has gone beyond that. Im enjoying the theory.
Its running on R22. The low suction is a reflection of the ambient at the moment. When the suction lifts in summer our head will also climb. Hopefully not to high. Probably have to lift the water flow rate then to lower the hp.
Ahah... -5'C suction temp. :eek:
What refrigerant was it using? R-22, or R-134a?
Added later:
I missed that in the earlier post. With an ambient like that, you are on a hiding to nothing, with an air-to-water pool heat-pump. If the ambient goes any lower, the defrost mechanism will probably begin to kick in - if it's equipped with one (some aren't in Australia).
It was going into defrost often. In tasmania we will battle to maintain the suction above zero. That has to be taken into account in the heat load. Their is no doubt that their is an issue with the pool heater. My whole issue has been with the fact that the suppliers throttled down the water flow thinking that they increased the capacity. That is why i look forward to seeing your water flow figures so i can appply the formula.
It was going into defrost often. In tasmania we will battle to maintain the suction above zero. That has to be taken into account in the heat load. Their is no doubt that their is an issue with the pool heater.
A few thoughts - why not preheat the air stream into the evaporator, so that it never sees a very cold ambient & does not need to go into defrost mode?
You could pull off some alternative warm air-stream for this.
In Australia, some folks actually put air-heater elements in front of the evaporator. for winter use. Sounds odd, but seems to be used.
-----------
Heat-transfer computation (local method):
For the water temp values, I reserve judgment on their accuracy. Current probe placement is not trustworthy at this point, in my view, & will probably throw you a few curved balls when you try & compute the instant condenser output. I NEVER trust local water probes (type K has accuracy around 0.4'C at very best, & closer to 1-2'C with poor placement), for heat-transfer calculations. Instead, I measure the end user of the energy e.g. water storage tank (250L), or swimming pool, over time & then compute the average heat absorbed by the water.
Nice logic with regard to applying that formula. How practically do you do it?
You would have to take the heat loss of the water container (pool) into account?
Whereas the other method of measuring the inlet and outlet of the heat exchanger would give you an instant reading of what the unit delivers without having to factor in other variables.? I do understand about the accuracy though.
I always measure what the end user is actually doing.
An estimate can be worked out for the convective & evaporative cooling from the pool if uncovered, or conductive/convective from the pool cover.
This can then be worked backwards to determine what the heat-pump actually delivers.
This method turns out to be more accurate, in most cases than using your equation which is actually:
Q'w = m'w*Cpw*dTw = rw*V'w*Cpw*dTw
where:
rw = water density [kg/m3] ~ 1000 kg/m3
V'w = water volume flowrate [m3/s]
Cpw = water specific heat [J/kg.K] ~ 4186.8 J/kg.K
dTw = Tw,out - Tw,in ['C]
Now, what happens on a small dTw, with probe uncertainty of +- 1'C & Tw,in=35'C; Tw,out=36'C?
dTw = (36+1)-(35+1) = 37-36=1'C
dTw = (36+1)-(35-1) = 37-34=3'C
dTw = (36-1)-(35+1) = 35-36=-1'C
dTw = (36-1)-(35-1) = 35-34=1'C
You'll have a range of interesting heat-transfer values. :)
The problem gets less significant for larger values of dTw, but still represents a significant experimental uncertainty.
That is why i look forward to seeing your water flow figures so i can appply the formula.
I had the impression that you were no longer involved with or had access to the pool heater.
If you still have access then go over there and measure the subcooling (SCT minus liquid line temp). If the subcooling is more than 15F/8.5K, then remove refrigerant until the subcooling is down to 15F/8.5K and let us know the results.
Check system for noncondensables...
How do we check for incondensibles being present in the system?
What is standard practice to remove air, other than through evacuation? Is a high-point bleeder point used e.g. service valve located at highest point in system?
Are you able to check this, Drew?
Hey Gary.
I dont have access to the pool unless i do it in my own time. I was going to use the flow rates on the figues that where obtained during DesA 's experiment to prove my point. maybe in due course i could obtain the subcooling specs and this would explain a lot.
DesA I hear you about heater banks on the evap surface to maintain a higher suction. I cant see this as being as effective as fitting electric elements in the water stream?
Being a tech and not an engineer i pick up formulas as i go and from what i read, so forgive me if I dont know the correct terms.
Condesables can be proved to be in a system by turning it off for a while to cool to ambient temp. We assume that the refrigerant should be saturated in the system, which means that we have vapour and liquid present. We then can read off our guages to see whether the ambient temp lines up with our guage pressure/ temp. If there is air in the system we what have the sum of the refrigerant and the air pressure which would be higher.
We also use this test to confirm what refrigerant is in a cylinder. This has become a bit difficult with the new refrigerants, but it gives us an idea. The only way to remove air in a system is to reclaim and evacuate.
DesA I hear you about heater banks on the evap surface to maintain a higher suction. I cant see this as being as effective as fitting electric elements in the water stream?
I understand that parts (?all?) of Australia were looking at legislation which make it illegal to use direct heating of the water stream - for energy-efficiency purposes. This forces the use of indirect heating mechanisms.
Under this scenario, the heating of the air-stream into the heat-pump could be considered as indirect heating.
Direct electrical heaters in water, for a swimming pool, do scare me somewhat - from a health & safety point-of-view.
Valid point. The ones that have been installed came as a kit from a pool shop. Not quite sure of the law in Tasmania with regard to that.
I recall back in the 1980's there was a heat pump which used an "A" coil in the outdoor unit... and under the A coil was a gas burner. I never actually saw one, don't recall who made it or know what happened with it... but it seemed like a good idea to me.
Condesables can be proved to be in a system by turning it off for a while to cool to ambient temp. We assume that the refrigerant should be saturated in the system, which means that we have vapour and liquid present. We then can read off our guages to see whether the ambient temp lines up with our guage pressure/ temp. If there is air in the system we what have the sum of the refrigerant and the air pressure which would be higher.
We also use this test to confirm what refrigerant is in a cylinder. This has become a bit difficult with the new refrigerants, but it gives us an idea. The only way to remove air in a system is to reclaim and evacuate.
Thanks very much for that information - very useful to know. :)
After my boss sent the suppliers a letter asking them to explain where they got the NOM capacity from they have agreed to replace the pool heater with a unit that is 2 sizes up!!
Now, isn't that what some of us have been saying all along?
Nice to see it's sorted out... :)
But it isn't sorted out. We will probably never know why it wasn't working properly.
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