Minimum discharge pressure for ammonia systems

[US Iceman]

To all ammonia guys, this ones for you...

I am interested in finding out how low you can operate your discharge pressures in the winter. What is the minimum pressure you can run?

Any problems with liquid supply to the system?

How many of you are using liquid injection for oil cooling for screw compressors?

Or even, what type of oil cooling are you using? What have been the problems you have had with your oil cooling?

How about hot gas defrost problems?

Has anyone had experience running at 10C condensing with ammonia systems?

[I LUV NH3]

I don't like to see the head pressure drop below 110 psig.

You loose the ability to hot gas defrost. Especially on evaporators that are at the farthest from the hot gas source. Typically defrost back pressure regulators are set to 70-75 psig. You can have up to 20 psig in pressure drop through the piping and valves not to mention the evaporator. This can add up to no defrosts.
Liquid injection? I design my systems with thermosyphon oil cooling. Most compressor manufactures agree that you can loose up to 11% of the compressor's capacity by cooling via liquid injection. Not to mention the reduced life of the bearings.

[NH3LVR]

I am a bit old fashioned and like to see the Head up a bit.
But that is not the way we do things anymore.
The head required to operate the Liquid Injection varies from Compressor to Compressor. Even in the same engine room between similar machines. Older FES and Fricks seem to have trouble lower than lbs (9BAR) In some cases even that is not enough.
One customer found he could keep his head down to lbs (11.8BAR) in the Winter. Of course his coils would not defrost and his TX valves could not keep up.
Another customer had a abundance of Condenser. He kept the head as low as it he possibly could.
I changed liquid line solenoids on a regular basis for him. He finally overfed a Intercooler and defeated the high level safety one night, to the tune of $10,000. Vilters do not pump liquid well.
Low head pressures are fine if the system is designed with that in mind.
We are quoting a Liquid Pump for one plant now, to be installed in the Liquid Injection line for the screws.
I have seen the Head as low as 96lbs (12.2 BAR) during a defrost this fall. This occasionaly shuts down on of his high stage machines.

[aawood1]

Hi, This year under orders from the people who pay the running bills we had to lower the Fan starts on the 4 water/ air condensers we us on one plant. The new start points are now set at 1st. condenser 8.5 Bar from 9.5 Bar, 2nd. condenser 9 Bar from 10 Bar, 3ed. condenser 9.5 Bar from 10.5 Bar. and the last one to 10 Bar from 11 Bar. The old settings have been used for the last 24 years with no problems, and we had a hot gas defrost system fitted to a freezer 8 years ago.

Will have to get over this Winter to find out what problems we will get from the new settings. The only one at this time is a bad hot gas defrost when we have a low load on the plant.
Arthur

[US Iceman]

I originally posted this over a year ago.:rolleyes:

There are some good ideas and concerns being listed so far. What I think this comes down to is, the system has to be designed to operate at low discharge pressures.

Liquid feed systems and hot gas defrost are the two big items. TXV's tend to crap out below about 125 psig (8.6 bar g), hand expansion valves on liquid make-up lines have to be opened further.

The hot gas systems require some thought. As mentioned earlier, it is common to have a high pressure loss in the defrost header. Why not make the pipe larger?

Use a drainer to keep condensate out of the header, so that it heats up faster. Besides, if you drain the condensate you have less opportunity to have liquid hammer during the start of the defrost cycle.

These can be done very easily on new systems.

On old systems, use a dedicated recip. compressor to supply the hot gas. If you operate a single compressor to supply the hot gas, you can operate it with higher discharge pressure, while the other compressors are operating at a much lower discharge pressure.

I've seen some owners install an abundance of condenser capacity to run 150 psig (10.3 bar g) in the summer.

I have also worked on some jobs where we were running at 80 psig (5.5 bar g) in the winter.

Another item to watch carefully is the minimum discharge pressure before it gets too low. Oil separators tend to get fickle around 140 psig (9.6 bar g).

[NH3LVR]

USIceman
A lot of good thought in this thread.
We size our HG Lines adequately, but there is always the problem of pressure difference between the Head and Evap. We sometimes go to 90lbs (6.2BAR) on defrost regulates. 100lbs (6.9BAR is not a lot of difference.
We almost always install drainers on the lines except when the runs are very short. It helps a lot.
Are you suggesting we run two head pressures/Condensers on older systems? I think I misunderstood.
We have a new Screw Machine waiting to be installed.
It is a Variable Speed and I understand it has a new type of valve (Not TX) for the Liquid Injection. It is supposed to overcome the Head Pressure Requirement for Liquid Injection. It is waiting for some Seismic Issues to be resolved before installation.

I had not considered a Oil separator problem before. I would have assumed they would work better under the lower Head/Temp conditions.

[US Iceman]

Hi NH3LVR,

It seems there are getting to be more of us ammonia guys here. That's great!

I think part of the problem of sizing hot gas defrost lines is; how do we estimate the total mass flow of refrigerant needed for defrosts? I've seen some systems where a 2" pipe is used, others install a 4" pipe. I would like to know how they sized this.

Small pipes make the problem more critical. Larger pipes help a lot.



Are you suggesting we run two head pressures/Condensers on older systems?


Not quite... What I'm suggesting is to have a spare/other recip. compressor piped into the hot gas header (instead of the main discharge line) to supply the evaporator coils with hot gas during defrost.

The other compressors could operate at the lower discharge pressure, while the dedicated recip. compressor (for hot gas) could operate at 150 psig (10.3 bar g) just to supply the hot gas for defrost.

The discharge temperatures are a lot higher than on screws, so the extra temperature helps to keep the hot gas header warm. If we use a drainer to keep the liquid out of the header, then it starts warm. You also don't have to wait for the header push all of the liquid out of it during defrost. This liquid normally flows into the coils at the start of defrost, which does no good either.



...I understand it has a new type of valve (Not TX) for the Liquid Injection. It is supposed to overcome the Head Pressure Requirement for Liquid Injection.


Is it a Jordan valve?

Low discharge pressures help to create higher gas velocities in the separator. Due to this, the carryover rate can increase, which is not good.:(

[Andy]

Hi:)

I set up all my plants for a 10.5bar head and when I am designing I use a mulitplier of 1.5 time coil duty for each hot gas line:)

Also it is better to fit a pump out system when not running the hot gas line, saves having to push large quantities of condensed liquid through valves not designed for it slowing up the defrost.

Liquid injection is a poor way of cooling, better with an oil cooler:) Oil coolers reject the heat out of the refrigeration circuit increasing COP, liquid injection adds to the compressor load. Oil coolers work better with low ambient, injection doesnot. Plants designed for economy use low heads and oil cooling:)

Kind Regards Andy:)

[US Iceman]

...I use a mulitplier of 1.5 time coil duty for each hot gas line.


I use a similar logic, except I tend to use a larger multiplier.

The way I look at this follows...

When the coil is ready for defrost there is a very large temperature difference between the cold coil and the saturation temperature of the hot gas defrost supply.

When defrost starts, there is a large quantity of condensate being formed. This means you need to be able to supply a large quantity (high mass flow) to the coil to defrost it quickly. To do this with low discharge pressures means you need a larger pipe ID, for the high mass flow with a low pressure drop.

The mass flow is very high when defrost starts, and decreases as the temperature difference (between coil & sat. temperature of gas supply) decreases.

Basically, I use a multiplier of 3 to account for this spread in mass flow requirements.

This results in a larger pipe ID, lower velocity in the hot gas header, and lower pressure drop. It works pretty nice.;)



Also it is better to fit a pump out system when not running the hot gas line, saves having to push large quantities of condensed liquid through valves not designed for it slowing up the defrost.


I prefer to use a high-side float valve to just drain the condensate. This keeps the liquid out, and also helps to keep the pipe warm.

In principle we are both doing the same thing. The condensate in the headers is bad!

[Andy]

Hi Iceman:)

what about the main hot gas line, by that I mean before you branch off to the coolers/valve stations, how do you determine the size for this:confused:

I use a 25% rule, by that I mean 25% of the coolers will be on defrost at anytime.

Kind Regards Andy:)

[US Iceman]

Hi Andy,

The main discharge line (to the condensers) would be sized for all compressors in normal operation as if the hot gas is not being used.

The hot gas defrost header would be sized for 3X the total coil capacity that could be in defrost at any one time. I think your 25% value is pretty close.

So, in effect I would use 25% of the total coil capacity (that this defrost header serves) and multiply that by 3. This makes the hot gas defrost header large enough to have a high capacity at low pressure loss.

You would also have to size the defrost header for a low pressure loss for whatever the total equivalent length is required for the installation.

The individual branch lines to each coil would be sized for 3X the individual coil capacity then also.

I would put a high-side float valve at the end of the header to drain condensate. That way, the header is dry and you would not have to purge liquid through the evaporator coils during defrost.

[Andy]

Hi Iceman:)

much as I thought you would say:D either I an under sizing the branch lines or I am selecting the hot gas lines on lower discharge pressures:)

Thank you:)

Kind Regards Andy:)

[US Iceman]

Hi Andy,

My normal preference is to slightly oversize the piping anyway. However, on hot gas defrost headers with my method it seems I usually end up with twice the diameter of what everyone else uses.

The smaller pipes work OK as long as you keep the discharge pressure up. The smaller pipes also let the pipes jump around somewhat when the hot gas solenoids open!:rolleyes:

If you sit down and look through some pipe selections for the higher mass flows and low pressure drops you will find the following:

Add the defrost pressure, plus the pipe pressure loss, plus the total valve station pressure loss for the hot gas and you end up with the minimum approximate hot gas supply pressure. If the pipe pressure loss is high, the minimum discharge pressure will be higher.

This is something I developed on my own, since I could not get any reasonable answer that was based on some fact or a little science.

[Andy]

Hi Iceman:)

The 1.5 comes from so selections Danfoss did years ago for me:)

The last company I worked for used 3-4 times:)

I have to say there design on most things was quite good;)

Some food for thought, we are going to start up a R404a system on hot gas defrost, it should be OK as it is DX (electronic) the suction is the hot gas line line to each cooler, I have already up the mains hot gas line size by twice what I had originally decided to use:D

I will let you know how it goes:)

Kind Regards Andy:)

[US Iceman]

The last company I worked for used 3-4 times:)


At least we have a consensus! I think the 3-4X number is a good average to size the branch and main headers. I also believe the value is considerably higher when defrost initializes, but drops down to a very low number when defrost is almost complete.

If anyone has a paper written on this I would love to see it.

The next trick is trying to get the defrost cycle to terminate to keep from blowing hot gas down the suction line.:D

[Josip]

Hi, all :)

Very nice I have to agree about almost all said, but isolation of hot gas piping is not mentioned;)

In past times we used one vessel named defrosting vessel to return liquid and gas after defrosting - no problem with return of high pressure gas into suction separator, but now day nobody use it.

Best regards, Josip :)

[US Iceman]

Hi Josip,



In past times we used one vessel named defrosting vessel to return liquid and gas after defrosting...


Would this vessel be operating at a high suction pressure close to the defrost pressure? Were you using a separate compressor operating at this high suction pressure?

If so, that would be a very efficient way to handle the defrost loads, but it would definitely cost more to install.

That might explain why it is not being used more.;)

[Josip]

Hi, US Iceman :)


Hi Josip,
Would this vessel be operating at a high suction pressure close to the defrost pressure? Were you using a separate compressor operating at this high suction pressure?

If so, that would be a very efficient way to handle the defrost loads, but it would definitely cost more to install.

That might explain why it is not being used more.;)

Defrosting vessel was one high pressure vessel above receiver (and little smaller then receiver) to drain all liquid after defrosting down to receiver. First we close the valve to low pressure separator and then equalize pressures by the valve on gas side between receiver and defrosting vessel, then we drain the liquid through the valve at the bottom of defrosting vessel.

Of course after drain it was necessary to close both valves on gas and liquid pipes between receiver and defrosting vessel. Inside of defrosting vessel remain only high pressure gas what was easy to remove and maintain low pressure by opening one small valve to low pressure separator -10C(14F) which remain open all the time until we refill the vessel.

Defrosting vessel was equipped with sight glass and high level switch to give you signal for drain liquid into receiver. Then we use that sub cooled liquid to charge separators again.;)

This system is good when you have a lot of low temp rooms, doors are opened often and you need to defrost them regular.

Yes, that was in past time when you have a lot of operators to run complete plant manually but now days all owners like to run their plants without operators because they are too expensive:D

Best regards, Josip :)

[US Iceman]

Hi Josip,

OK. I see what you are describing. I've seen something like this before. It's a gravity drain transfer vessel.

I had not thought of using this for defrost condensate.

Where I've seen these, the vessel at a higher elevation is used only for an accumulator. When the liquid level rises to a certain height, the level switches open and close solenoid valves for equalizing and venting so the liquid runs into the receiver.



Yes, that was in past time when you have a lot of operators to run complete plant manually but now days all owners like to run their plants without operators because they are too expensive:D


Yep, you are correct. I had a discussion with someone the other night about this. After everyone gets dependent on control systems to operate their plant, they will forget how to run it manually.:o

Thanks for the description on the defrost vessel.;)

[TXiceman]

Some of the newer liquid injection valves are electronic Danfoss valves. Much wider range of control than a TXV.

Ken

[Mark Sanchez]

Gentlemen,

Great information!

Typically what I've found is that the system should be designed for the lower head pressure.
ie:- hot gas line sizing for lower pressure differentials.
- Drainers should definitely be incorporated.
-Oil seperator sized for lower head/ if the seperators are not oversized for lower disch press. the coalescers will blow out due to the higher density of the lower pressure gas.

Another thing to look for is that the customer may be using more electricity to run the lower head , than the savings being recognized by the compressors. Check the HP of the fans and pumps being utilized vs the amp draw of the compressors at the lower heads vs say 125# disch. Longer defrost times will also complicate this as the hot gas relief has to be recompressed. Ineficiencies are worse on systems that relieve hg def to the low stage/temp vs relieving to the hig temp.


Hope this adds some food for thought

Mark

[US Iceman]

Some of the newer liquid injection valves are electronic Danfoss valves. Much wider range of control than a TXV.


I have not tried any of those, although an opportunity to may arise shortly.

Ken, Have you used any of the Jordan temperature control valves? These are the same ones Sullair used for liquid injection on their screws. These seem to work pretty well and are simple.

[TXiceman]

I have used a few Jordan valves. They do Ok in a system that pretty stable and the compressor unloads and loads slowly. The valves are pretty slow acting and can over and under feed with a rapid load swing for the compressor. In that case, the oil temp may swing out of range.

Ken

[nh3wizard]

Frick installed the Danfoss valves on the liquid injection on thier new packages, as long as you get the Quantum paraimeters set right the valve works great.

[NH3LVR]

Just looked at a new FES Variable Speed Screw a couple of days ago. They are also using the Danfoss Valve in place of the TX on the Liquid Injection.

[Andy]

Hi:)

the AKVA valve as Danfoss calls it is the only really good way of close control of superheat on DX ammonia that I know of:)

Kind Regards Andy:)

[NH3LVR]

Andy;
Sorry, I may have left the wrong impression. The Danfoss valve that FES is using is a motorized Valve.
I was in the plant to take some measurements for the installation and unwrapped the valve.
I noted the Model Number down carefully. Only when I got home did I realize I only had the Actuator Number.
Will punish myself with some Real Ale.

[Andy]

Hi NH3LVR:)

Thats the new ICM motorized vlave module for the ICV range that came out to replace the PM valves. V seats on the ICV make for real control. I have seen these same valves fitted as Co2 expansion valves on a chiller at the IKK show. Should be a good valve:)

Kind Regards Andy:)

[NH3LVR]

Thank you for the Info, Andy.
Although Danfoss has been selling TX Valves here for years, the Industrial use is only now becoming common. RS is our most used supplier.

[Andy]

Thank you for the Info, Andy.
Although Danfoss has been selling TX Valves here for years, the Industrial use is only now becoming common. RS is our most used supplier.

No Problem:)

Danfoss have also a new valve system for NH3 control, a bit like a hydralic block, with all the controls on one main block. This would be good for packaged NH3 equipment, I'm thinking scraped surface systems and possibly chillers
http://www.danfoss.com/NewsAndEvents/News/RA+Products/Compact+control+solutions+for+industrial+refrigeration.htm

Kind Regards Andy:)

[Bradford]

Another thing to look for is that the customer may be using more electricity to run the lower head , than the savings being recognized by the compressors. Check the HP of the fans and pumps being utilized vs the amp draw of the compressors at the lower heads vs say 125# disch.

Good point. We applied floating head pressure to one plant that runs about 5,000 hp during the week for process and 500 hp on the weekends for cold storage. Because the head pressure setpoint was fixed, every condenser device could run trying to achieve setpoint even though only 10% of the compressors were on. During our tests, our control system caused the compressors to consume about 8 kva more but the condensers went down by about 300 kva! All by raising the head pressure setpoint slightly.

Brad

[Sergei]

Hi, everybody. I've just signed in.
Bradford, you need PLC with wet wulb approach. This feature will help you to run this plant efficiently at different ouside conditions and at different refrigeration loads.

Regards

Sergei

[US Iceman]

Hi Sergei,

I noticed this was your first post and thought I would say welcome. Please feel free to participate in the on-going discussions.

As a matter of curiosity, how are you measuring wet bulb temperature for the approach?

[Sergei]

Hi Sergei,

I noticed this was your first post and thought I would say welcome. Please feel free to participate in the on-going discussions.

As a matter of curiosity, how are you measuring wet bulb temperature for the approach?

Hi, US Iceman.
I know 2 approaches to measure wet bulb temperature.
1. Measure dry bulb temperature and relative humidity. Calculate the wet bulb temperature.
2. Use wet bulb sensor for measure this temperature.

Sergei

[US Iceman]

I know 2 approaches to measure wet bulb temperature.
1. Measure dry bulb temperature and relative humidity. Calculate the wet bulb temperature.
2. Use wet bulb sensor for measure this temperature.


I have heard the RH sensors are cheaper than wet bulb sensors. Which one are you using? Have you noticed any problems with the type you use?

Thanks.
US Iceman

[Sergei]

I have heard the RH sensors are cheaper than wet bulb sensors. Which one are you using? Have you noticed any problems with the type you use?

Thanks.
US Iceman
I don't have information about the prices. As far as I know Frick is using wet bulb sensor, but Hench is using RH sensor. These sensors should be calibrated regularly, bacause wet bulb temperature is base temperature for tunig up of optimum head pressure.

Sergei

[US Iceman]

OK, thanks for the information Sergei.

[Bradford]

Bradford, you need PLC with wet wulb approach. This feature will help you to run this plant efficiently at different ouside conditions and at different refrigeration loads.

Hi all, been out for a while, good to be back.

For measuring wetbulb, we use a combination temperature / humidity unit. Wetbulb temperature is calculated using the Raphson-Newton method for best accuracy. The current units that we are having built have been in operation for about 5 years with no failures and no need for calibration.

[Sergei]

Good point. We applied floating head pressure to one plant that runs about 5,000 hp during the week for process and 500 hp on the weekends for cold storage. Because the head pressure setpoint was fixed, every condenser device could run trying to achieve setpoint even though only 10% of the compressors were on. During our tests, our control system caused the compressors to consume about 8 kva more but the condensers went down by about 300 kva! All by raising the head pressure setpoint slightly.

Brad
Hi, Brad.
How you control your head pressure? By wet bulb approach or by fixed head pressure set point.

Sergei

[Josip]

Hi all :)

Check this:

http://www.sabroe.com/fileadmin/filer/pdf/Controls_CPO_for_R717_0410.pdf
3+ MB

it is Condenser Pressure Optimizer working with ambient temp sensor and RH sensor.....

Best regards, Josip :)

[nh3wizard]

Does anyone have any experience working with these optimizer's? If so how well do they work?

[US Iceman]

Hi nh3wizard,

I have seen this before on the web, but have no experience with them. I suspect the guys from across the pond will have something to say.:)

You might be able to get them from M&M Refrigeration, since they are the Sabroe rep here in the US.

If you try one of these let me know about your experiences.

[Bradford]

Sergei

Wetbulb approach. Also, each evaporator has it's own pressure reset for defrost

[Bradford]

Hi US Iceman,

We include this type of floating headpressure in all of our refrigeration control applications. Current installations are accross Canada and in the US. Europe and GB are next.

As I indicated elsewhere in theis thread, I have seen energy reductions of over 300 KVA.

Regards, Brad

[Sergei]

I've looked at this optimiser. Sorry, I don't believe in magic black boxes or in aluminium boxes even from Sabroe. Operating engineers should have information about settings for their refrigeration plant as well as access to change these settings. Sometimes the right settings for one plant are not working for another one. It looks like this optimiser for refrigeration plant with one condenser and condenser fan VFD. Majority of industrial plants have 2 or more condensers. What is the minimum allowable condensing pressure? Every refrigeration plant has it own minimum allowable condensing pressure.

Sergei

[Sergei]

Hi US Iceman,

We include this type of floating headpressure in all of our refrigeration control applications. Current installations are accross Canada and in the US. Europe and GB are next.

As I indicated elsewhere in theirs thread, I have seen energy reductions of over 300 KVA.

Regards, Brad
Hi, Bradford.
Sophisticated PLC is your tools to save energy. You should have the right settings and the right operating strategies to maximize your energy savings.

Regards, Sergei

[US Iceman]

What is the minimum allowable condensing pressure? Every refrigeration plant has it own minimum allowable condensing pressure.


Now we getting to the main problem as I see it.

Each plant is different and does have it's own minimum dicharge pressure. The real problem is finding what particular part of the refrigeration system determines this.

Now, I do not know much about the Sabroe Optimizer if it is for one condenser or many. But I do know that big systems with many condensers can be a challenge to setup properly.



Operating engineers should have information about settings for their refrigeration plant as well as access to change these settings.


Absolutely right. The operators need to be able to change settings as required and also need to be aware of why the setpoints are important.

I am not too fond of control systems that use "fixed setpoints" that are not adjsutable by anyone other than the programmer.

I have seen some control systems that can reduce the total energy by over 40% or more. So I know what is possible. I also believe the operators need to understand what the control system does and what it is capable of, so that the setpoints do not get changed to an old value used before.

Good control systems do not replace good operators.

[Josip]

Hi, all :)

I've looked at this optimiser. Sorry, I don't believe in magic black boxes or in aluminium boxes even from Sabroe. Operating engineers should have information about settings for their refrigeration plant as well as access to change these settings. Sometimes the right settings for one plant are not working for another one.

Of course operating engineer has info and access to those settings. Each plant has own settings even if they are installed in the same factory;) due to different design or duty.



It looks like this optimiser for refrigeration plant with one condenser and condenser fan VFD. Majority of industrial plants have 2 or more condensers.

Agree!

But still we control only one i.e. last one started;) All equipment started earlier we push up to maximum capacity. Trying to control all devices, compressors, condensers, pumps, fans, etc simultaneously we will drive complete system into "hunting" situation and finally to alarm stop.


What is the minimum allowable condensing pressure? Every refrigeration plant has it own minimum allowable condensing pressure.

Having good plant with enough capacity, with good control of each device we can go very low with condensing pressure but then we have to heat something i.e. defrosting cycle, what now?

Switch on manual and come back to 30/35C (speaking about NH3).

It seems to this question is not easy to give simple answer;)

Coming back to "optimizer" its name is telling us we can try to get what we pay. (is that good expression?)

I am assuming also all PCs, PLCs and other black boxes as devices installed to help-optimize our work. Of course they must be programmed for that. Even the best operator is only human thus capable to make mistake;)

Best regards, Josip :)

[US Iceman]

Hi Josip,

You raised some good points in your last reply.



Having good plant with enough capacity, with good control of each device we can go very low with condensing pressure but then we have to heat something i.e. defrosting cycle, what now?

Switch on manual and come back to 30/35C (speaking about NH3).


I think we can design the systems to operate at 10C (50F) as a reasonable condensing temperature, when the wet bulb temperature or ambient conditions allow.

But, we have to determine how not to use components that have a higher minimum inlet pressure (TXV's or liquid feed valves).

Another way to look at the problem is as you suggest; run low discharge pressures when possible and reset to higher pressures for a short time.

Another possibility is to use a separate compressor for the hot gas source, while the other compressors are running low discharge pressures.

I do not think there is a single solution to this, but there are a lot of opportunities!;)

[Sergei]

[quote=US Iceman;57175]
But, we have to determine how not to use components that have a higher minimum inlet pressure (TXV's or liquid feed valves).
Regarding to TXV. Usually, high temperature evaporators are equipped with ammonia TXV. Major refrigeration load for the coolers with these evaporators depends on outside temperature. Lower condensing temperature we can get at cool ambient conditions. Capacities of TXV's are going down, but refrigeration load is going down as well.
Regarding to liquid feed valves. Usually, we have solenoid and expansion(balancing) valve. Expansion valve, usually, set to 50% of operating time for high condensing pressure. Sometimes this setting is working well for all(low and high) condensing pressures. Sometimes 30% of operating time is working fine for all condensing pressures. Sometimes you need 2 setting(summer and winter). I think it is not difficult to readjust this expansion valve twice per year.

Sergei

[US Iceman]

Sergei,

I don't disagree with you. The points you mention are valid and as long as everyone understands there might be some limitations, then they know what they have to design for.

Personally, I prefer to use liquid overfeed systems instead of DX systems. A liquid overfeed system has very few limitations and no requirement for minimum discharge pressure (unless it is a gas powered recirculation system, with pumper drums).

Secondly, by the time you add the transfer system to the cost of the accumulator, you are getting close to the cost of a vessel and pumps. I don't think there is a lot of difference in cost between the two systems.

At least not enough to justify selling DX, but I recognize I may be in the minority on this viewpoint.;)

[Sergei]

Sergei,

I don't disagree with you. The points you mention are valid and as long as everyone understands there might be some limitations, then they know what they have to design for.

Personally, I prefer to use liquid overfeed systems instead of DX systems. A liquid overfeed system has very few limitations and no requirement for minimum discharge pressure (unless it is a gas powered recirculation system, with pumper drums).

Secondly, by the time you add the transfer system to the cost of the accumulator, you are getting close to the cost of a vessel and pumps. I don't think there is a lot of difference in cost between the two systems.

At least not enough to justify selling DX, but I recognize I may be in the minority on this viewpoint.;)
Hi, US Iceman.
Reduction of condensing pressure has several barriers. Liquid overfeed will eliminate one of them. Good design is a base for potential energy saving. However, I found that over 50% of potential energy savings can be saved by improving operating set points(one of them is minimum allowable condensing pressure) and operating strategies. This approach doesn't require significant investment and has payback just a few months. As far as know food companies in North America will invest in energy efficiency only if payback is shorter than 2 years.

Sergei

[US Iceman]

Hi Sergei,

It seems we are on the same page.



Reduction of condensing pressure has several barriers.


Indeed. The task is to find those and minimize their effect. The same applies to raising the suction pressure as much as possible.



However, I found that over 50% of potential energy savings can be saved by improving operating set points(one of them is minimum allowable condensing pressure) and operating strategies. This approach doesn't require significant investment and has payback just a few months. As far as know food companies in North America will invest in energy efficiency only if payback is shorter than 2 years.


I believe the 50% savings. There are a lot of systems that just don't run properly. Part of the problem is the initial design, the other part is how the system is operated.

This is where it pays to carefully review the system and electrical tariffs. Significant improvments can be accomplished in almost any system.

You are correct about the 2 year payback. This is something that everyone sticks to. However, I prefer to look at this as 50% ROI.

It seems the thought process with a 2 year payback stops after 2 years. By this I mean, people do not pay too much attention to what happens in the third year.

Putting it into an ROI perspective, they acheive a 50% return every year (based on the original costs). If energy costs increase, the ROI improves. This is something they get each year.

This argument has some semantics, since the above both mean the same thing. The important idea is how to present the data in terms that are meaningful.

[US Iceman]

I also wanted to add some other thoughts. I feel the system should be designed to work efficiently under all of the operating conditons; summer/winter & full load/part load.

All too often I see systems designed for the worse case (summer & full load) with little regard for the other conditions.

The way I see it, the systems should be designed for efficient operation all of the time, not just at one point.

Any control system used should simply provide the means to sustain the optimum operating conditions during those periods, and as loads change.

[Sergei]

Sorry, I don't agree with raising suction pressure as high as possible. I believe in optimum suction pressure.

[US Iceman]

Sorry, I don't agree with raising suction pressure as high as possible. I believe in optimum suction pressure.


OK, I suppose I should ask about your concerns or perhaps clarify my statement.

To me, the optimum suction pressure would be the highest pressure that still meets the temperature requirements of the load.

With this in mind I would attempt to find the temperature requirements for the loads attached to a common suction line. If we know the temperatures required by the load, and the performance of the evaporators, we can find the optimum suction pressure.

In a lot of cases we see back-pressure regulators set to maintain the evaporating temperatures, but the suction pressures are usually less than what the lowest evaporating temperature needs to be.

So if we can increase the operating suction pressure high enough to keep the back-pressure regulators open all of the time, we have reached our optimum suction pressure.

This is the essence of what I meant about getting the suction pressures as high as possible.

[Andy P]

Hello all - interesting debate going on here! Getting back to the head pressure "optimiser" I'm a big fan of keeping things as simple as possible, and it occurs to me that the difference between wet and dry bulb temps in a temperate climate is on average very small. For example in the UK the difference is usually less than 2 deg C, and we only see significant differences for prolonged periods when the dry bulb gets above 15C. Why not save the cost of a fancy black box "optimiser" and use a simple dry bulb measurement offset by 2 deg C when it reads less than 15C (and revert to standard head pressure control when it is higher than 15C - not often round here!)

...or maybe that doesn't sound simple at all!

btw the Australian Fridge Institute (AIRAH) used to provide ambient temperature charts where the frequency of wet/dry bulb combinations was printed on a psychrometric chart, and you could see the effect I'm trying to describe very clearly. Excellent data presentation, but I have never seen it anywhere else - anybody know of the same for other countries?

cheers

Andy P

[US Iceman]

...the Australian Fridge Institute (AIRAH) used to provide ambient temperature charts where the frequency of wet/dry bulb combinations was printed on a psychrometric chart,...


Is this similar to the revised ASHRAE data Andy, or the TMY2 data?

Can't say I have seen this printed on a psych chart.:confused:

[Andy P]

If I can find a page I will scan and post so you can see the idea - it is a very compact way of delivering a large amount of useful info: anybody using evap condensers should have it.

cheers

Andy P

[Mark Sanchez]

Andy,

Great point. I would like to see the chart as well. I'm sure someone has one for the States.

We have installed a few systems with the wetbulb sensors.
It would really only seem to matter if you're operating a large system with a low disch. press. setting, and it was set below the actual saturation corresponding to WB. Mainly in the middle of summer.
Chicago has a design of 95/78 (not if you refer to the latest ASHRAE, but the NH3 guys still use it) so a corresponding saturation = approx 132# Disch.

If you did not have the WB feature, the system would try and run more condenser fans/ pumps w/ no avail. as the saturated Disch press. setting was not attainable. Or you can stop a VFD on a condenser fan from driving higher if in fact you have already achieved the WB/ Saturation.

I'll have to look into the Sabroe system.
This sounds very interesting.
I'll forward the link for this site to Greg @ M&M to see if he knows anything about it.

It's great too see so many technicians & engineers involved with this site!!

Cheers
Mark.

I'll introduce myself if I see you at IIAR

[Sergei]

Hi, all.
Regarding to optimum suction pressure. Don't forget about evaporator fans. Their use energy and release this energy in refrigerated space. We have to spend additional energy for refrigeration plant to remove the heat from evaporator fans. Usually, for every 2HP(horse power) of fan energy we have to spend additional 1 HP for refrigeration plant to remove the parasitic heat from evaporator fans. Increasing suction pressure we save energy for compressors,but we use more energy for evaporator fans. At certain conditions, we will save less than we spend. Major criterion of optimum suction pressure is coefficient of evaporator coil HP(fan power) per TR(ton refrigeration) per degF(temperature difference).
Regarding to wet bulb approach. Idea of wet bulb approach is to balance capacities of compressors and evaporator condensers keeping total power(compressors + condensers) at minimum level. This feature is useful when outside temperature over 10C. If ambient temperature is lower than 10C, we usually run the plant at minimum allowable condensing pressure. It is very important to keep wet bulb approach at optimum level. Assume that optimum wet bulb approach is 10F. If we try to keep 6F( 4F or 2C difference), our condenser capacity should double. We can spend additional 200HP for condensers and we save just 50HP for compressors. We overspent 150HP.

Sergei

[US Iceman]

Regarding to wet bulb approach. Idea of wet bulb approach is to balance capacities of compressors and evaporator condensers keeping total power(compressors + condensers) at minimum level. This feature is useful when outside temperature over 10C. If ambient temperature is lower than 10C, we usually run the plant at minimum allowable condensing pressure. It is very important to keep wet bulb approach at optimum level. Assume that optimum wet bulb approach is 10F. If we try to keep 6F( 4F or 2C difference), our condenser capacity should double. We can spend additional 200HP for condensers and we save just 50HP for compressors. We overspent 150HP.


That's the very point I'm talking about with clients right now. In a similar thought as the optimum suction pressure we were discussing, the optimum discharge pressure is not the lowest discharge pressure you can achieve.

It is the lowest discharge pressure which uses the least amount of total energy (condenser fans + compressor motors).

You raise a very good point that all should consider Sergei. Energy savings should be based on the total use, not the individual components.;)

[PHANANHTUAN]

Re: What's PM3 valve? i can't understand it, help me ! please.
thanks !:confused:

[Josip]

Hi, PHANANHTUAN :)

Welcome to RE,


Re: What's PM3 valve? i can't understand it, help me ! please.
thanks !:confused:

your post is in wrong thread but try here:

http://www.ra.danfoss.com/ra/Products/ProductCatalogue.asp?BA=&Division=RC&HLID=510&HL=3&AppID={c2a95dac-5e01-47df-92cd-5972d837cf1a}
http://www.ra.danfoss.com/ra/Products/ProductCatalogue.asp?BA=&Division=RC&HLID=515&HL=3&AppID={c2a95dac-5e01-47df-92cd-5972d837cf1a}
http://www.ra.danfoss.com/TechnicalInfo/Approvals/Files/RAPIDFiles/01%5CDrawing%5CA27f172%5CA27f172A5.jpg

but you can search complete site to learn something more;)
http://www.danfoss.com/Products/Categories/
under Refrigeration and A/C
Hope this will help.

Best regards, Josip :)

[Andy P]

If I can find a page I will scan and post so you can see the idea - it is a very compact way of delivering a large amount of useful info: anybody using evap condensers should have it.

cheers

Andy P
Here is the chart for Adelaide, Australia. The figures in the middle of the chart are the number of hours that the combination of wet & dry bulb temp is achieved. The figures at the bottom of each column are the number of hours per year for that dry bulb temperature and the figures on the left hand side are the number of hours per year that the wet bulb temperature is exceeded. Charts are also available for 6am-6pm and for 6pm-6am in each location. I hope that this is clear enough to give the general idea - the 6 and 8 digits look very similar, so it is a bit difficult to read the detail.

cheers

Andy P

[US Iceman]

Hi Andy,

Thanks for taking the time to scan and post this information. I have never seen the data listed in this manner, but as you stated, the data is a very good source for those using evaporative condensers (or other evaporative processes, such as cooling towers, etc).

Unfortunately, I suspect this format is localized to Adelaide, Australia. Other charts would need to be generated for specific locales with the pertinent weather observations.

In 1997 ASHRAE revised their weather data, which trys to emulate a similar data set. The listed the dry bulb temperatures and mean coincident wet bulb temperatures in various percentages. The percentages are based on the percentage of annual hours in the following format; 0.4%, 1%, & 2%.

This translates into 35 hours, 87.6 hours, & 175 hours. The data provided for each percentage value, would provide a similar, but not as complete, set of values.

The flip side to the ASHRAE data is that there is also another set of data listed in the same format, except, this data lists the wet bulb temperature and mean coincident dry bulb temperature.

In this data set, the wet bulb temperatures are considerably higher for the same frequency values. For evaporative processes I would recommend using the values of wet bulb temperature and mean coincident dry bulb temperature.

I also like to use the TMY2 data for all of the annual hours to develop curves of the maximum, minimum, and average wet bulb temperatures to illustrate a graphical method of displaying the weather data. They end up looking like bell curves.

It's also a good idea to add some safety factor for the wet bulb temperatures and condenser capacity, because as we know, "things happen.".:D

[Sergei]

Hi, all.
I still believe in wet bulb approach, because it is major function to determine capacity of condensers relatively to capacity of compressors. Wet bulb approach is not complicated as it looks.
Measure wet bulb temperature by wet bulb sensor(for example 65F). Set the optimum wet bulb approach(for example 10F). Add these two numbers( 75F). Corresponding pressure to 75F is 126psig(ammonia). This is our set point for head pressure. If wet bulb temperature change to 70F, our set point will change to 138 psig( 80F). This is floating condensing pressure. This feature is working very well for refrigeration plant with fluctuating refrigeration load( up and down). For refrigeration plants with steady refrigeration load we can balance capacities of compressors and condensers just switching on(off) some condensers.
I have question to US Iceman. Why do you think that hot gas lines are undersized?

Sergei

[US Iceman]

I have question to US Iceman. Why do you think that hot gas lines are undersized?


A fair question...

Most of the systems I see usually take at least 30 minutes to defrost a coil (that's an average value for the entire cycle). My question is why does it take 30 minutes to defrost a coil? I have seen other systems where a coil was defrosted in under 10 minutes.

My thought on this is the speed of defrost is determined by the volume of hot gas that can be supplied to the evaporator in question. And, the ability to drain the condensate as fast as it forms.

In the past I have seen large evaporators that also take much longer to defrost (up to one hour).

This got me to thinking... Why does it take so long?

Obviously, the defrost ability is determiend by how much gas you can supply to the evaporator for defrosting.

The capacity of the hot gas header is determined by the available potential, i.e., pressure differential from the engine room to the coil.

In the summer the discharge pressure is much higher, so the available potential is much greater as there will be a greater differential pressure to allow higher mass flows.

In the winter with reduced or floating discharge pressure, the potential is much less, hence you see lower mass flow and potentially longer defrost times.

If the hot gas header is sized sufficiently large for low pressure drop (e.g., larger pipes) the ability to implement fast defrosts is provided.

In other words, the capacity of the hot as header is determined by the pressure loss across the pipe. Higher pressure losses in this pipe will require higher discharge pressures to maintain the same flow capacity.

This is one of those barriers to implementing low discharge pressure operation.;)

[US Iceman]

It is my belief the use of floating discharge pressure is a good method.

The weather data Andy P presented is a reference tool for use during the design process.

I think it's important to distinguish between operational and design parameters.

The floating discharge pressure method would simply be a function of the data Andy P posted and the capacity of the condensers and the load on the refrigeration system.

[Sergei]

I have a different opinion. I think that major reason of poor(slow) defrosting is under supply of hot gas due restriction in balancing valve. We don't have this problem at high condensing pressure, because we have higher pressure difference to supply hot gas to the coil and density of hot gas is higher at higher condensing pressure. Due to these 2 factors the mass flow of ammonia to the coil significantly greater at higher condensing pressure.
Pressure drop in main hot gas header. This header usually designed for defrosting from 25% to 50% of all coils. For example. We have cold storage with 6 penthouses and 4 coils in each penthouse. Total is 24 coils. Main hot gas header designed to defrost 6 coils simuntaniously. Most likely we defrost not more than 2 coils at the same time. Hot gas flow in main line will be 3 times less than design flow and pressure drop will be just 1-2 psig. Lower density of hot gas is another factor to reduce pressure drop.

Sergei

[US Iceman]

We don't have this problem at high condensing pressure, because we have higher pressure difference to supply hot gas to the coil and density of hot gas is higher at higher condensing pressure. Due to these 2 factors the mass flow of ammonia to the coil significantly greater at higher condensing pressure.


I think we saying essentially the same thing. Hot gas systems seem to work better at higher condensing pressures. However, I think they do so because the systems are designed for the summer time, and not quite for the conditions that result in winter operation.

[Sergei]

I think we saying essentially the same thing. Hot gas systems seem to work better at higher condensing pressures. However, I think they do so because the systems are designed for the summer time, and not quite for the conditions that result in winter operation.
Do you have any idea about winter design of hot gas supply?

[US Iceman]

Well let's see about this...

The defrost headers need to be able to be drained of condensate. Liquid drainers work well for this purpose. In the winter time, the colder ambient temperatures tend to condense a lot more gas.

The defrost headers need good DRY insulation. I have seen a fair amount with bad insulation jobs, or others that are not maintained properly. This causes more condensation also. All of the condensate has to flow through the coils (unless drainers are used), which increases the defrost time.

The hot gas header needs to be large enough to supply sufficient hot gas to the coils (at a low pressure drop, since we hopefully will be running low discharge pressures). All too often I see hot gas headers of about 2" or so, which is just too small in my opinion. This is probably the biggest problem.

The headers need to be pitched for gravity drainage. I see some occasionally that have the headers trapped between two different elevations.

Come to think of it, the design process is about the same for summer, except the colder air temperatures increase the condensation in the hot gas headers.

[Sergei]

Drainers are essential to hot gas lines. Certainly, they improve defrosting. These drainers should be installed for the safety reasons as well. They will prevent the coil damage during initial stage of defrosting.
In winter time we loose the superheat of ammonia in hot gas header. This is the major reason of poor defrosting for the farthest coils.

Sergei

[NH3LVR]

Most of the systems I see usually take at least 30 minutes to defrost a coil (that's an average value for the entire cycle). My question is why does it take 30 minutes to defrost a coil? I have seen other systems where a coil was defrosted in under 10 minutes. My thought on this is the speed of defrost is determined by the volume of hot gas that can be supplied to the evaporator in question. And, the ability to drain the condensate as fast as it forms.
This got me to thinking... Why does it take so long?

I have been watching this thread with some interest.
I am speaking mostly of Nh3 (Flooded) Evaps as they are what I have the most recent experience with.
Many Cold Storage Evaps we install today are put in a Penthouse, with the Valving outside. This works well and is easy to service, compared to inside mounting. However many of this Evaps have the Coil Outlet behind panels where it cannot be observed easily. Any buildup on this side of the coil is hidden until it restricts the Airflow. Then it becomes a problem to remove as it can be two inches thick, and embedded in the Coil.

Our Defrost Cycle commonly consists of;

3-5 Minutes: Pumpdown (Not very effective with Ammonia)
(?) Minutes: Fans stopped, and Hot Gas on (May use Soft Gas at first to to avoid Shocking Coil thermally or Mechanically?)
3 Minutes: Equalizing time (Bleed Down)
2-5 Minutes: Liquid back on, Fans off.
Fans on, normal operation

With proper drainage of the Hot Gas Line a Coil operating at -20F (-29C) can be defrosted in 25-30 Minutes. But 8-10 Minutes of that are not in the Hot Gas mode.
I have seen many Operators shorten these times, and pay for it later.

[Sergei]

I have been watching this thread with some interest.
I am speaking mostly of Nh3 (Flooded) Evaps as they are what I have the most recent experience with.
Many Cold Storage Evaps we install today are put in a Penthouse, with the Valving outside. This works well and is easy to service, compared to inside mounting. However many of this Evaps have the Coil Outlet behind panels where it cannot be observed easily. Any buildup on this side of the coil is hidden until it restricts the Airflow. Then it becomes a problem to remove as it can be two inches thick, and embedded in the Coil.

Our Defrost Cycle commonly consists of;

3-5 Minutes: Pumpdown (Not very effective with Ammonia)
(?) Minutes: Fans stopped, and Hot Gas on (May use Soft Gas at first to to avoid Shocking Coil thermally or Mechanically?)
3 Minutes: Equalizing time (Bleed Down)
2-5 Minutes: Liquid back on, Fans off.
Fans on, normal operation

With proper drainage of the Hot Gas Line a Coil operating at -20F (-29C) can be defrosted in 25-30 Minutes. But 8-10 Minutes of that are not in the Hot Gas mode.
I have seen many Operators shorten these times, and pay for it later.
I prefer at least 10 min. pumpdown to evaporate all liquid in the coil. Hot gas supply is in the same range 25-30 min. Hot gas supply should be optimal. Undersupply will lead to poor defrosting. Oversupply will lead to significant parasitic load on compressors, because only certain portion of the gas will condense and the rest will go through the coil, BPR to the suction line. This supply depends of coil size, refrigerated space temperature, amount of frost on the coil and etc.

Sergei

[NH3LVR]

I prefer at least 10 min. pumpdown to evaporate all liquid in the coil. Hot gas supply is in the same range 25-30 min. Hot gas supply should be optimal. Undersupply will lead to poor defrosting. Oversupply will lead to significant parasitic load on compressors, because only certain portion of the gas will condense and the rest will go through the coil, BPR to the suction line. This supply depends of coil size, refrigerated space temperature, amount of frost on the coil and etc.Sergei

I agree, although I have never ran across a situation with an oversupply of Hot Gas to an Evaporator.
I do question pumping out all the Liquid in Evaps (In the case of NH3). Due to the high Latent Heat of Ammonia that could take a very long time. Pump recirculated Coils are the worst in my experience. We now return the condensed Liquid to the Intercooler if possible.

[US Iceman]

Oversupply will lead to significant parasitic load on compressors, because only certain portion of the gas will condense and the rest will go through the coil, BPR to the suction line. This supply depends of coil size, refrigerated space temperature, amount of frost on the coil and etc.


NH3LVR, What Sergei is describing is the samething as the coil being in defrost (with the hot gas still on) after the coil has been defrosted. Once the frost is melted off, additional hot gas continues to heat the coil resulting in steaming and a lot of gas flowing out of the coil.

The gas flowing out of the defrost regulator adds to the compressor load (booster if the hot gas returns to the booster suction, or to the high stage if the condensate line is piped back to the intercooler).

Since the defrost pressure is around 75 to 90 psig ( 5.2 to 6.2 bar g), it would make sense to have the condensate piped back to either, a separate vessel with a separate recip. compressor, or a controlled pressure receiver. Whatever is close enough to work.

Pumping out ammonia evaporators is a pain for the reason stated. It's the same problem with trying to get a flooded out screw started. A little bit of ammonia takes a lot of heat.

[NH3LVR]

it would make sense to have the condensate piped back to either, a separate vessel with a separate recip. compressor, or a controlled pressure receiver. Whatever is close enough to work. Pumping out ammonia evaporators is a pain for the reason stated. It's the same problem with trying to get a flooded out screw started. A little bit of ammonia takes a lot of heat.

US Iceman;
I had never thought about the concept of using a separate Compressor. Has this been done before, could you elaborate?
On the subject of Flooded Screws, one of our Customers managed to fill up a Screw Booster with NH3 the other day. It took 3 days to get it going again. They now believe in the safety Shutdown on the LPR we wanted to install.

[US Iceman]

HI NH3LVR,

What I'm thinking about is to use a separate vessel (accumulator if you will) to collect the gas and liquid from the defrost regulators. If the suction pressure on the compressor attached to this vessel is running higher than any other house suction pressure, it just means the penalties for hot gas defrost would be much less.

In other words, we are taking the condensate up to as high as pressure as we possibily can. From that point, a separate compressor could run the suction pressure with no problem (especially a recip).

I've seen controlled pressure receivers run this way, so it should not be any big jump in logic to say try this for defrost.

The trick would be to size the compressor to handle the defrost loads accurately. This is just me thinking out loud. I have not done this for defrost, but I can't see why it wouldn't work.




On the subject of Flooded Screws, one of our Customers managed to fill up a Screw Booster with NH3 the other day. It took 3 days to get it going again. They now believe in the safety Shutdown on the LPR we wanted to install.


Ohhh, those are always fun jobs to get.:D

[Sergei]

Frequency of defrosting is another important factor of refrigeration plant optimization. Right now, here in Ontario, one cold storage has defrosting once per week, but another one has defrosting twice per day. I found that overdefrosting is common problem for many refrigeration plants. It is difficult to justify additional compressor installation, if you have defrosting once per week.

Sergei

[US Iceman]

I found that overdefrosting is common problem for many refrigeration plants. It is difficult to justify additional compressor installation, if you have defrosting once per week.


I would agree. Each facility has it's own personality in how it operates and what it needs. This also changes with the seasons (summer, winter, etc).

The concept I posted earlier is just one way of dealing with an issue. Just like in troubleshooting, there is always a cause and effect we need to account for in any instance.

[Bradford]

Frequency of defrosting is another important factor of refrigeration plant optimization. Right now, here in Ontario, one cold storage has defrosting once per week, but another one has defrosting twice per day. I found that overdefrosting is common problem for many refrigeration plants. It is difficult to justify additional compressor installation, if you have defrosting once per week. I’ve been reading this post with great interest. What started out as a discussion on lowering head pressure and it’s implications has become a discussion on total system optimization! And it seems to be a topic that a lot of people are interested in. With the ever rising costs of energy and the push for a greener planet, I wonder if this would make a good topic for a forum or sub-forum.

What Sergei mentions about defrosts going from twice per day to once per week has become quite typical in plants that incorporate smarter controls and/or keener operators. Evaporators that require 2 defrosts per day during that one hot week in august can often get away with one or two defrosts during the winter months. This gets back to the problem of fixed setpoints often being the “worst case” setpoint. If two or three defrosts are needed in the summer, that’s where the pins are left in the time clock. If -25 suction kept the rooms cold in August, that’s where the suction is left all year. If one evaporator requires 145# for defrost, that’s where the pressure switches are set. It’s obvious that a lot of systems have a lot of room for energy usage improvements

Another great up-side to long durations between defrosts is you can float the head pressure lower for a longer period of time thus enhancing energy savings.

Brad

[US Iceman]

a discussion on total system optimization! ... I wonder if this would make a good topic for a forum or sub-forum.


That's good idea. Let me run it past the boss of RE for his comments.

[US Iceman]

It’s obvious that a lot of systems have a lot of room for energy usage improvements


That was my idea for starting this thread. Refrigeration systems waste too much energy doing something that can be done for less.

I continue to believe the system can be designed better and more flexible. The control systems actually provide the ability to sustain the original design concepts with the capability of smarter operation.

One other important criteria is the utility bills, i.e., how much energy and demand cost, the various time of use charges, and ratchet clauses can also present some challenges.

[stan1488]

recip plant on vfd ;) ( im giving my tricks away) :P

[TXiceman]

recip plant on vfd ;) ( im giving my tricks away) :P

This is nothing new, I did this years ago and it worked like a champ on lowering energy demand and smoothing out plant operation.

You have to think outside the box all of the time now.

Ken