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Heat Pump Defrost Cycle Design

Heat Pump Defrost Cycle Design

Heat Pump Defrost Cycle Design

(OP)
My job is to interface various HVAC and other equipment to Building Automation Systems.  I have just (about) completed the interface of a Carrier Infinity Heat Pump to a standard Building Automation Controller.  This involved complete replacement of the existing Infinity HP controller so as to allow detailed control and monitoring of the HP by the building automation system.  

Every thing is finished and working fine, but I am thinking that the DEFROST CYCLE for the HP is more complex than it needs to be.  Maybe someone can educate me.  

What I implemented (in accord with Carrier's Manual) is:
1) Set minimum defrost interval to 30 minutes at power up.
2) If minimum defrost interval timer has expired AND OCT (outside expansion coil exit temp) is less than 32F for 5 minutes, then initiate defrost cycle.
3) If the prior defrost cycle took less than 3 minutes (for OCT to reach 65F which signifies "defrost complete", set minimum defrost interval to 90 minutes.  If it took 3>6 minutes set next interval to 60 minutes.  If more than 6 minutes set next interval to 30 minutes.  (This 30/60/90 selection is the continuous running time UNTIL the next defrost is ALLOWED to occur. The defrost cycle is not run unless #2 above is satisfied.

I have noticed that the typical OCT coil exit temp to Outside Air temp when running in HIGH HEAT MODE is about 8 degrees F after stabilizing for 5 minutes.  I also have noticed that the OCT temp difference to outside air temp increases as the outside coil ices up as one would expect.  When the coil to Outside Air temp is about 12F, the coil is getting "fuzzy" with ice.  

It seems to me that if I keep a long term trend of the delta-T (temp difference, Outside Coil to Outside Air in HEAT MODE) and use that as a baseline, I should enter the defrost cycle about when the delta-T gets to 12F or so.

Is there anything I am missing with the above protocol?

Also,  Is there a good rule of thumb criteria for a temperature below which a heat pump should not be run?  Some have told me 40F, others 37F and others 32F.  It seems to me that as long as the compressor AMPS are within specs and the defrost cycle is working the machine should be capable of operating independent of any set minimum temperature.  

My criteria for terminating heat pump operation might be if the defrost cycle took more than (say) 20% of the "between defrost cycle" HP running time then stop using the heat pump until the temperature rises.  Anyone have any thoughts here?

Many thanks for your helpul comment.  

Joe

RE: Heat Pump Defrost Cycle Design

w2jo
I am going to preface this by stating that I have no experience with heat pumps except my little through the wall Carrier unit which cools and heats my barn which I converted to a home office.  However heat pumps are basically air conditioners so I will give you my input for what it is worth since no one else with more experience has replied.

Based upon my defrost experience you can use any reliable coil frosting criteria that you can find to initiate defrost.  This is difficult for some applications (like -45 deg. F frozen food processing) but easier with yours.  I am not sure if you mean that OCT is the outside coil leaving air temperature or the outside coil refrigerant temperature as it leaves the coil with superheat.  I suspect the latter since as the coil ices there would be less load and the suction pressure would decrease lowering the temperature.

The minimum temperature for heat pump operation depends on the compressor, the refrigerant, and the discharge conditions.  You won't have a problem with amps since as the suction pressure decreases you have less mass flow of refrigerant due to the decreased density of the refrigerant (check typical compressor curves or tables to see the power trend).  You will probably have a problem with compressor overheating since the compression ratio will be increasing and the mass flow of refrigerant (for compressor cooling) will be decreasing.  You need to monitor compressor head temperatures.  You also do not want to operate beyond maximum compression ratios (usually due to the overheating already mentioned but sometimes due to compressor max pressure rating) and there is no reason to operate once the COP at your suction and discharge conditions is less than one (since electric heat will be more efficient then).  Your unit probably has a Carrier or Carlyle compressor so it will be a little more difficult to get operating information on it (unless they OEM it out to other manufacturers).  If it was a Copeland, Danfoss, Bitzer, or other OEM unit then you could get full information on its operating characteristics.  

There should be some application engineers for some of the compressor manufacturers that can give better comments.  I am just an ignorant end user of the equipment.

RE: Heat Pump Defrost Cycle Design

(OP)
THANKS for your post..  The information about the compressor characteristics was new.  The compressor in this particular unit is a Copeland so I can likely get data on it.  

In the absence of more definitive information, (and remembering one of my professor's comments to the effect that a well designed and executed experiment is worth 10,000 expert opinions) I have the following experiment operating on this unit.

I have instrumented:
1) Outside Air Temp at the Heat Pump.
2) OCT (temp at the discharge end of the Outside heat exchanger.
3) SUCT (temp of the  suction line (larger copper tube) as it enters the outside enclosure.)
4) LIQT (temp of the liquid line (smaller copper tube) as it enters the outside ecnlosure.
5) AMPS (compressor running amps.
6) FANA  (Heat Pump Fan Amps.)
7) DIFF (differential air pressure between the outside and inside of the heat pump enclosure box.)  I am thinking that this (cheap and easy to obtain) parameter may well be a key measurement in determining when the coil is freezing up.

I would like to have access to compressor suction pressure, but since this parameter is not easily available in the general case, I am not instrumenting that for now.

Now I am awaiting a "big freeze" here in the Atlanta area.  I am logging each point once per minute into the controller memory and I will watch the coil itself and the parameters and over time, I think I will be able to establish a criteria for some defrost cycle criteria that can be of general use with this sort of equipment.  My goal is to design an algorithm for the controller that will automatically adapt to "most any" rooftop or even residential heat pump so as to (intelligently and with efficiency) allow the unit to operate to the lowest efficient temperature before switching to auxiliary heat.  

I have also implemented a COP calculation for the heat pump and have included an entry for the "cost per therm" for the auxiliary heat so as not to accidently run the heat pump below its cost effective temperature even if it is able to do so.

If anyone has any suggestions, comments, helpful hints, please DO.  I am an EE stumbling around in ME territory here and can use all of the help I can get!

Thanks!
Joe

RE: Heat Pump Defrost Cycle Design

w2jo
if you can't get the information on the compressor then someone can run it for you through the Copeland selection software.  I would measure the head temperature if you plan on running at lower ambients than recommended.  Copeland has a recommended location to measure the temperature if you research their website (if you have an OEM password).

RE: Heat Pump Defrost Cycle Design


...

the temperature below which the heat pump should not run is called the balance point. At this point the heat pump dose not produce heat economically cheaper than your supplemental heat source. A compressor runs with approx. 30 moving parts (UNLESS A SCROLL).
When you have your compressor capacity curve and your structure heat loss, the balance point is a straight line to the point at which you don't save money anymore...hence activate the supplemental heat.

t  

RE: Heat Pump Defrost Cycle Design

w2jo
send me the Copeland model number, refrigerant, and the condensing temperature (which will be the air leaving coil temperature plus the coil degree of approach (or temperature difference).  I will let you know which is lower, the point at which the COP=1 or where the compressor either overheats or exceeds the lower limits on suction pressure that Copeland has put on its operation.  I will post the compressor performance table so you can see it.

RE: Heat Pump Defrost Cycle Design

(OP)
Thanks Gepman!
I will get the model number for you.  I do have the "balance point" curves for the compressor but frankly, I did not know how to put it to use.  Would a copy of the graph be useful to you?  

I have installed a 0-1" differental pressure sensor across the inside to outside of the HP outdoor unit box.  At 100% fan and no icing the pressure drop is about 0.11".  The max icing I have so far developed increased the pressure drop to about .12" and had no measurable effect on the (OAIR - OCT) reading which is nominally 7F in high heat mode.  Just now, the low temps here are in the range of 30F so I am not getting much data.  The outlet temp differential at 48000btu (high) setting is staying about 20 to 22F so that looks fine as well. I have run it down to OAIR temp of 30F and performance was as above.  Compressor current draw was about 11 amps and fan current about 1.8A in the heat pump unit at 30F.

I think the differential air pressure across the coil will prove to be a good indicator of icing.

Thanks for the help.. I will get back with the compressor model.

RE: Heat Pump Defrost Cycle Design

The important thing to remember is the HP's outdoor coil operates 20 to 25 degrees below ambient to absorb heat.

If time & temperature controlled, the HP will initiate defrost with a coil temp of 25 degrees(45-50 deg ambient) by closing the defrost relay contacts.

Increasingly, manufacturers are using an air pressure switch to measure the pressure drop across the coil resulting from ice accumulation on the coil by wiring the air pressure switch with timer & temperature sensor.

Defrost termination normally occurs at 50 degrees at the location of the temperature sensor -- defrost normally should last for 10 minutes.

RE: Heat Pump Defrost Cycle Design

temp of outdoor fan is not required since the outdoor fan should not be running in the defrost cycle.

In the defrost mode, the HP's 4 way valve reverses to cooling mode making the outdoor coil the condensor -- outdoor fan must be off(and auxillary heat running to prevent cold air entering the conditioned space) to insure efficiency.

RE: Heat Pump Defrost Cycle Design

tbedford,

The balance point isn't neccesarily the point that the HP should not run(unless the auxillary heat is gas or oil) -- the balance point is temperature of the structure were the HP needs auxillary heat. If the auxillary heat is electric, the electric can be initiated in stages while the HP is absorbing heat from the outside air.

Typical air-to-air HP's have COP's of 1.5 at an ambient of 0 degrees -- still better than electrics 1.0

RE: Heat Pump Defrost Cycle Design

(OP)
I surely do appreciate your input since my last posting!!  I have been doing quite a bit of experimentation in the past week.  For anyone interested, a draft of my experimentation is posted at http://gpsinformation.org/joe/HeatPump.html

I have to wait for some cold(er) weather before I can fully quantify my recommended defrost algorithm.

RE: Heat Pump Defrost Cycle Design

(OP)
With regard to the comment by EmeraldCoastHVACR, I must agree.  This particular heat pump is cost effective (and fully rated) down to -3F.  Provided of course you can keep the outside evaporator defrosted! It is indeed "cost effective" all the way to -3F.

For the Carrier 25HNA948A30  (a nominal 48,000btu/h rated heat pump) that I am using, the output is rated at 16,000btu/h at -3F with total system KW input of 2.3.  Thus, the per therm cost of this heat at -3F is just $1.15.  This still beats today's NG, Oil and Electric Heat costs.  BUT.. Obviously you are going to have to have supplemental heat on a day like that!

RE: Heat Pump Defrost Cycle Design

(OP)
With regard to temp drop from one end of the evaporator coil to the other:  I am measuring about 6 to 7 degrees F on this Carrier 25HNA9 48,000 btu coil in the steady state in both HIGH and LOW heating modes.  It does go up to about 15F differential when the unit first starts, but within about 5 minutes, the differential is down to the 6 or 7 degrees (with an ice free evaporator).  

RE: Heat Pump Defrost Cycle Design

w2jo
You still haven't given me the model number of the Copeland compressor.

If you are really interested in heat pumps my company has done a lot of research for electric utilities on cold climate heat pumps.  Unfortunately the results are for use by the utilities since they paid for the work.  However here is a link to one of the first in commercial construction.  If I get permission to release our white paper I will.  

http://www.gotohallowell.com/assets/DS_White_PaperLR.pdf

As you can see they designed it how a good refrigeration engineer would do it when faced with a very high compression ratio, they went to a two stage compound system.

RE: Heat Pump Defrost Cycle Design

(OP)
Hello Gepman,
I was not able to get a cross to a Copeland number.  Instead, the Carrier Rep sent me a document with all the curves and tables I needed for the 25HNA9048 (48Kbtu/h) unit so I am all set.  I do appreciate your help and offer of more help.  I still have a lot to learn I am SURE!  

I read the Hollowell article.  Most interesting!  I knew that we were not very deep into recovering the maximum efficiency from the heat pump's refrigerant cycle.  I am delighted that the ME guys are hard at work on the problem for colder climates.  I am fortunate that I have both a heat pump and NG backup.  Many citizens do not have NG and propane and oil are not far from the cost of Resistance heating.  

Thanks again for the White Paper.  If you have any more such tutorial info, I would be delighted to read it!
Joe

RE: Heat Pump Defrost Cycle Design

Another point on HP's:

A rule of thumb is usually your heating btu/hr will be around 2 to 2.5 times your cooling btu/hr.

It's kind of interesting that if you buy a HP with variable speed motors & fans and size it for your heating load, you'll save a lot more money over the course of a year than if you sized the HP for cooling(which most contractors do)
and used auxillary heat in the winter.

The 4 ton HP with VS fans & motors will use comparable amperage at similar cooling loads when compared to a regular 2 ton HP, but would have 48,000 btu/hr for heating.

RE: Heat Pump Defrost Cycle Design

(OP)
Hmmm..  I notice that for the 25HNA9 48Kbtu/h rated heat pump:

In Cooling Mode, the cooling rating at 350cfm/ton, 72F inside ambient, 96F condenser ambient air, the rating is 51.16Kbtu/h.  

Then for the Heating mode,  the heating rating at 350cfm/ton, 70F inside ambient, 47F evaporator (outside) ambient is 48Kbtu/h.  

I have never looked at heat pumps in this detail before so other heat pumps may well be quite different.  The Product bulletin on the Carrier 25HNA9 "Infinity Ultra" can be found at http://gpsinformation.info/joe/25hna9-2pd.pdf

There is a lot of useful technical information for engineers in this product bulletin.

RE: Heat Pump Defrost Cycle Design

The capacity ratings will all depend on the condensing and evaporating temperatures for each case.  I took a look at the Carrier product bulletin.  The compressor is almost certainly a Copeland "Ultra" (which for them means it has two capacity points, 67% and 100%).  The model number is most likely a ZPS49K4E.  I didn't go through the performance table of the Carrier and the Copeland to match them (since Carrier does not give the evaporator and condenser degree of approach [or temperature difference] for you HVAC guys) to try to match it.  It could be Copeland's next larger size.  

I have attached both a performance table for the compressor and Copeland's suggested operating envelope.  It shows what I was trying to say at the beginning of the thread that as the suction pressure drops the allowable condensing pressure drops, most likely due to overheating.  Remember the gas cools the motor and the heads.  On their Discus compressors you will sometimes see fans on the heads to cool them.  Sometimes they will liquid inject the scroll to cool it (which of course reduces the efficiency).  Industrially we would either water cool or thermosyphon cool the oil, not that practical for a home unit.

Also Carrier states that this is a "two stage" unit.  It has two capacity steps but it is not a "two stage" unit since it does not have two separate compression steps.  I had this same discussion with Hallowell (see my post above regarding the cold climate heat pump.  Although he has a true two stage system he was calling the high stage compressor with two capacity steps a "two stage" compressor.  I finally gave up on the discussion because it probably is easier for the average consumer to think of two capacity steps as two stages instead of two compression steps.  It may confuse an engineer though.

EmeraldCoastHVAC is right in that most people will size a heat pump on the cooling load but due to the greater spread between the evaporating and condensing pressures during heating (especially when it is very cold) the compressor will have less capacity.  The heating load does not necessarily equal the cooling load.  Both loads should be checked.  That is why the two step or other modulating type compressors are good in a heat pump since it can modulate to the load which may be different between the cooling and the heating case.  You can't use a VFD on a Copeland scroll since it is a compliance type scroll and relies partly on the speed of the orbiting scroll to keep the seal.



RE: Heat Pump Defrost Cycle Design

(OP)
Thanks gepman.. You are just FULL of interesting data!  I was fascinated by the description of how they achieve 2 "stages" of capacity in this scroll compressor.  They have a valve that they can open part way up the scroll to "unload" the compressor.  I gather this means they simply do not achieve the full capacity pressure differential and that this unloads the compressor motor and reduces the capacity.  My first thought was that the SEER would be worse in the 67% (low) capacity mode.  But the SEER numbers (IF I assume that the Carrier capacity data tables are correct) are within less than 1 SEER unit if I go back and forth between low and high speed under a constant Outside Air and Inside Air temp and with constant inside duct pressure.  Of course, in this scenario, the inside delivered supply air temperature does show the appropriate change from high to low capacity.

I will read your attachments and see what they are about.. THANKS for sending them along.

Joe

RE: Heat Pump Defrost Cycle Design

(OP)
Hi Gepman,
The charts you linked to are very interesting.  The "Operating Map" in particular is useful.  

The operating map appears to be for a heat pump in heating mode whereas the table appears to be in air conditioning mode.  Do you also have an operating map for a heat pump in cooling mode and a table for the heat pump in heating mode.

Is there more to the document?  Perhaps you can give me a link to the entire document??

Thanks
Joe

RE: Heat Pump Defrost Cycle Design

w2jo
The way that Copeland achieves two steps of capacity is with a "valve" or "port".  It does not affect the suction and discharge pressure which basically are controlled by the evaporator and the condenser but affects the "effective" swept volume of the scrolls.  The less swept volume the less mass flow of refrigerant and therefore the less capacity of the compressor.  It is very similar to unloading a cylinder on a reciprocating compressor.  You can see an animation of this at http://www.copelandscroll.com/index2.html the click on "products", then click on "ultra tech", the click on "view ultra tech animation".

The operating data that I sent you was generated by the Copeland Compressor Selection Software.  The operating envelope is for the compressor whether it runs as a heat pump or an air-conditioner.  Remember the compressor doesn't know if it is rejecting the heat inside the house (heat pump) or outside the house (air conditioner).  It just cares about the suction and discharge pressures.  The performance table was generated around the particular set of operating conditions that I gave it (where the star is on the operating envelope).  I can generate a performance table around any set of operating conditions that you want so if you have a particular set of conditions, let me know.  I can also generate the information for the compressor at 67% capacity.  I can put in any amount of superheat and subcooling that you want.  If you are referring to the 95 deg. F ambient "air over" in the upper left of the performance table I think this is a fixed value.  This does not refer to the temperature of the air over the condenser (remember this table is for the compressor only and does not care what the condenser is), it is for the temperature of the ambient air around the compressor which affects whether the compressor will overheat.  Obviously in heat pump mode the compressor will be cooled more efficiently by 10 deg. F air than by 95 deg. F air.  That is why you could put some temperature sensors in the motor winding and in the discharge gas to see if you are overheating instead of relying solely on the operating envelope.  Copeland has some information on this for OEM's that design around their compressors.

RE: Heat Pump Defrost Cycle Design

(OP)
Hi Gepman,
I did look at the Copeland site as you suggested.  I would like to see a operating map for a heat pump in cooling mode and a table for the heat pump in heating mode for the Copeland ULTRA model, nominal capacity in colling.  My 25HNA9 unit claims a nominal cooling capacity of 48,000btuh at outlet air 72 and Outside Air of 100F.  Superheat is specified at 12F for this system.  

I tried to sign up for access to the site, but I never got a confirmation and my chosen PW does not work.  Maybe I have to phone them to gain access.

Thanks for your help.  The information you have furnished so far have helped my understanding a LOT!
Joe

RE: Heat Pump Defrost Cycle Design

w2jo
Now we get into the problem that I have every day when I try to back calculate efficiency or capacity of condensing units.  I do this because many times there is no published information on the EER of a condensing unit at various conditions, only the capacity but not the EER.  Manufacturers almost NEVER give the degree of approach, temperature difference, or heat transfer rating in BTU/deg. F of either their indoor or outdoor coils (in the case of you A/C Heatpump you can't say what the function is of each coil since it changes).  

What I have done is assumed a 22 deg. TD on the indoor coil (72-22=50) and a 30 deg. TD on the outdoor coil (100+30=130).  Using 12 deg. F of "useful" superheat (superheat which occurs inside the evaporator) and 3 deg. F of "non-useful" superheat (superheat which occurs in the suction piping after the evaporator).  Useful superheat contributes to the cooling effect of the refrigeration system, non-useful superheat does not.  This gave me very close (within 100 BTU/H) of the capacity which you stated.  I am not much of an expert on these small commercial units but 10-20 degrees on the typical evaporator and 20-30 degrees on the typical condenser is what I have seen.  Industrially on large ammonia systems I would normally design 8-10 degrees on the evaporator and 10-15 degrees on the condenser (usually an evaporative condenser).  Other people may go higher to minimize capital cost.  

I have attached the actual calculation at that those conditions along with the performance table.  As I said previously the performance table and the operating envelope will be the same whether it is a heat pump or air conditioner since the compressor doesn't care.  When calculating the performance tables the software does NOT allow me to enter the ambient air temperature so it always defaults to 95 deg. F.  The ambient air temperature, as I said previously, does NOT affect the compressor capacity but would slightly affect the operating envelope, since I believe that this only affects the cooling of the compressor.  Lowering the ambient air temperature would probably raise slightly the allowable condensing pressure in the operating envelope since more heat could be removed.  You would have to call Carrier or Copeland to find out how much.  

RE: Heat Pump Defrost Cycle Design

gepman,

The evap coil to inlet air temp varies 10 to 20 TD depending on the system, but, in HVACR engineering, the AHU's are sized by the TD of the inlet air before the evap coil and supply air after the coil, which rarely exceeds 20 deg (Q = 1.08 x cfm x TD).

The condensor TD is an actual temp of condensor coil from compressor discharge to end of condensor coil -- 30 deg is a standard efficiency condensor with higher efficient ones condensing closer to 20 deg above ambient.

To actually measure the amount of subcooling at the condensor & superheat from the evap outlet to compressor suction, manifold guages must be installed to correlate refrigerant temp to piping temp -- only a licensed & skilled technician like myself is qualified to do this -- if this procedure isn't done correctly, oxygen could mix with the refrigerant and cause a violent explosion.

superheat is never measured in the evaporator, only from the evap outlet to suction of compressor -- along with subcooling at the end of the compressor, the superheat adds to the total "refrigerant effect" of the system.

RE: Heat Pump Defrost Cycle Design

(OP)
Hello EmeraldCostHVACR.

If you are warning  me not to "mess around" with refrigeration equipment, I am modestly qualified.  I have a HVAC Contractor's License, and a Electrical Contractor's License.  And I have installed and engineered a myriad of smaller Air Conditioning systems.  BUT.. My experience with the engineering details and operation of heat pumps is indeed minimal as you can likely surmise.

Therefore, I <really DO> appreciate any tutorials that you and Gepman may choose to pass my way. I am soaking it all up!  :)

I have added some interesting graphs of my 4 ton Carrier Infinity  Heat Pump (modified with new defrost algorithm) for your possible interest.  They are at the bottom of the article at http://gpsinformation.org/joe/HeatPump.html

The LEGEND is shown in the first graph.  From the graphs you can see that the new algorithm a) cuts the expected defrost cycles by a large percentage.  I still have not had weather much below 30F and I expect the improvement to be a lot less in below freezing operation.  

So far, I have operated the system down to about 28F ambient and the COP has stayed above 2 even at 28F.  Not bad.

RE: Heat Pump Defrost Cycle Design

w2jo
I salute you for the great amount of good work and intelligent thinking that you accomplished.  You might want to see if you can sell your information to Carrier.  You did hit on a difficult problem which is determining when to defrost.  This is a problem in a lot of cold storages and as you noted it wastes a lot of energy.  Heatcraft has a new method of defrost initiation that I need to check out.

My wife would kill me if I put that much time into our HVAC system!

RE: Heat Pump Defrost Cycle Design

(OP)
Gepman:
My motto is:
You spend 40 years of your life working for others so that in your retirement you can run all the experiments you want of your own selection - guilt free!  And sometimes make a $ in the process..  :)  

RE: Heat Pump Defrost Cycle Design

(OP)
I have placed the latest article on my work at
http://www.gpsinformation.org/joe/HeatPump2.html

This system has now allowed my heat pump to operate normally and with FAR FEWER DEFROST CYCLES all the way down to 16F.  I presently have it set to switch over to natural gas at 20F as this particular 48,000btu/h heat pump is down to 30,000btu/h at 20F and that plus the second furnace output of 48,000btu/h is not adequate to go to lower temps.  I then go to the backup 90K btu/h + the 48K btu/h gas furances.  

But above 20F the heat pump runs fine and as  you can see from the graphs in the article, it averages less than 1 defrost cycle per day.

RE: Heat Pump Defrost Cycle Design

The rate at which you frost up will be closely related to the Ws of the air and the dew point depression....as your suction pressure drops and refrigerant mass flow decreases, the heat absorbed at the evaporator and consequentially the mean temp difference at the evaporator decreases.  If the Tube Outside Temperature never reduced below the ambient dewpoint you would not make any frost...

As to Defrost due to (air) pressure difference, there has been LOTS of eperimenation on that topic with lots of published data. much of it for industrial uses but not that much different when you get past the intensity of the vernacular...

Also Sight Eye type devices compacted so they wrapped around and looked at the coating on a single tube....

The neatest one I have seen used emissivity in a similar way to a motion sensor, and the sensor element (tranmitter and receiver combined) moved around on one side of a wall of 6-coils in a very large industrial food freezer.  Don't know if the electronics could be rationalized. but such a "seeing eye", fixed at a short range, wcould have a fkeld of view of enough coil that all the local effects would be averaged out...

Just a thought.

RE: Heat Pump Defrost Cycle Design

(OP)
Thanks  Sterl for your comments.  I think an optical scheme such as you suggest would work, but my experience with such devices outside in the weather has not been good.  Dirt, water, and films are constant environmental factors which cause optical devices in the "dirty outdoors" to be problematical.

As you can see from the latest graphs at http://www.gpsinformation.org/joe/HeatPump2.html
the differential pressure sensor is working out quite well.  We just went through a couple of days of snow and sleet and the unit defrosted every 2 hours or a bit less during the worst of it.  So.. During such times this differential pressure scheme defrosts at about the same rate at the T&T defrost designs.  However, I like to point out the weeks it has gone without ANY defrosts when the temp was in the 30 to 40F range and the humidity was below 85%.  As you noted, Humidity and dewpoint are what determines an icing situation.  Most defrost controllers do not have a clue about that but the differential pressure scheme takes it into account automatically.

RE: Heat Pump Defrost Cycle Design

sterl
Do you have a brand name or a link to the IR defrost initiation device?  It has been a while but the last time I tried any of these devices they didn't last long in the -40 to -70 deg. F. blast freezers that I used in the frozen foods industry.

w2jo
Can you post a sketch of the orientation of your dp sensor to airflow and the coil?

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