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I wanted to post our experience with head pressure and electric costs. We have two Fricks 300hp and one Frick 150hp screws. The tower has three 15hp fans. We lowered the head pressure from 150psi to 120psi. Our savings each month during the summer months were about $5000 USD and the spring/fall months were about $3000 USD. The two 300hp Fricks operate most of the time and tower does have scale build-up.
Hope this helps any curosity...
Good post!
Yes, every 1 deg C lower condensing temp, you use 3 % less energy and you gain 3 % capacity.
You should put in for a bonus;)
Were you able to calculate the extra water usage and costs.
It is good first step. It will be good if can go below 100 psig.Quote:
Originally Posted by That's too cold
It depends. Don't forget that to lower condensing pressure you spend condenser energy. Sometimes you can use more condenser energy than energy saved by compressors.Quote:
Originally Posted by coolstuf
We experienced the same thing with one of our plants.
In the plant are 4 x 185 kw, 2 x 300kw, 1 x 400kw and 1 x 200kw motors on screws running 24/7. i lowered the discharge pressure from 1100kpa to 950kpa by adding another big condenser and the savings where massive. the condenser payed itself in 6 months.
The other thing i did was to install a air purger that removes water and other solid impurity's from the plant (The plant runs at -45'c vacuum) and this also helped the condensers to cycle thus saving water.
The money spend in total were payed back in the firs year.
The capacity of the compressors also improved in such way that some of the compressor were cycling witch never happened in the past.
Nice idea, screw needs to pressure lubrication:eek:
What do u mean Josef , that the screw will pressurize/ pump more oil if the discharge are lower than 1100 kpa?Quote:
Originally Posted by josef
Some screw compressors require minimum pressure differential for oil lubrication. Typically it is 55 psig or 370 kpa. If your plant has these compressors, you have a lot of room for improvement.Quote:
Originally Posted by Lodiev
Segei Yes, not all compressors have an oil pump. Thaťs says Frick - an important oil pressure. It can happen when your savings goes to the repair of compressors, in some cases even less.
I think that if Frick gives 55 psig , it is safe number. You don't need to double or triple this number to make sure safe operation. I found that compressor manufacturer usually give overprotected numbers(I can understand them).Quote:
Originally Posted by josef
One example. Large dairy has Frick compressors. Manufacturer recommended minimum condensing pressure 140 psig due to oil carry-over. It took a while to convince them gradually to lower this pressure to 125 psig. It saves them $100,000 annually. They can buy a new compressor every year (if they need it).:)
Yes Segei while saving you energy and does not invest in early repair, it is saving.
Another example. Recently I visited a plant that has Frick compressors RWB II - 134E. In winter time they have pressure differential 70-80 psig. One compressor run 110,000 Hrs without overhaul.:eek:Quote:
Originally Posted by josef
110,000 hours? Never met so long.
The one thing you have to be careful about is oil carryover. At lower condensing pressures, discharge gas is less dense so the volumetric flow through the oil separator is much higher. This can lead to too high velocity in the separator and then oil carryover because of less efficient separation. On new plants, specify an oil separator selected for low condensing pressure. It may be a little bigger and a little more expensive than the base design but the money comes back in energy saving and reduced oil carryover.
Definitely it can be an issue. However, many plants operate at unnecessary high condensing pressure. Very often cond. pressure can be reduced by 20-25 psig and oil carry-over will not increase. Just do it gradually and a lot of energy will be saved without harming the plant.:)Quote:
Originally Posted by charlie n
I agree completely with you Segei.
I was wondering where you got this 3% from.Quote:
Originally Posted by coolstuf
A friend recently gave me a pile of articles distributed by the Institute of Refrigeration Engineers Service Engineers Section. I started reading them last night - got through about 6 of them - all talk a **** load of nonsense - I think I'm going to ask a magazine if they can publish a series of articles for me so that I can critique all the nonsense they're distributing out to the industry who I would say they are effectively dumbing down - not educating.
I see an IOR Service Engineer article written by a Jim Rusling of Mitsubishi Electric (Dec 2009) claims a magical number of 3% (rule of thumb) to changes in both capacity and power consumption. Actually - the power consumption goes up 1.5% for a 1K rise in saturated condensing temperature.
And it is actually nearer 1% capacity loss per K rise in ambient temperature thus 1% per K rise in saturated condensing temperature - which tends to follow the rising ambient for a given evaporator load.
Just double checked Frick data. Power use of the compressor change 2-3% per 1 degC. Capacity change depends of the cycle. However, it is not significant for ammonia. This is for screw compressors and it can be different for recip.
These are the numbers for the compressors. What about the refrigeration plant? To lower condensing pressure, we should use additional condenser energy. To save energy, plant(compressors, condensers, evaporators) energy use should be evaluated. The same story on suction side. Higher suction pressure will improve efficiency of the compressors. However, it will increase energy use of the evaporator fans. System(plant) performance should be evaluated.
There is no energy saved when for a given ambient or room condition you increase the cooling or cooled medium mass flows to suspend LMTD's. Energy savings are only possible when head pressures are permitted to drop with ambients and/or suction pressures are permitted to rise with reduced loads by cycling compressor capacity and constant compressor isentropic efficiencies prevail.Quote:
Originally Posted by Segei
Tech's in the field do not work with stand alone compressors - they work with systems. System capacity drops about 1% for every 1K rise is saturated condensing temperature.
I have seen the the 1% duty change and the 1.5% power change many many many times at chiller factories across Europe during thermal and acoustic tests.
I have also attached a graph from a Daikin Service Manual.
Hi DTlarca,
"Tech's in the field do not work with stand alone compressors - they work with systems. System capacity drops about 1% for every 1K rise is saturated condensing temperature."
Sorry, don't agree with above statement. When you lower your condensing temp, you increase your refrigeration effect. Your capacity is RE X Mas flow rate, not RE + mas flow rate, therefore you will not get a 1:1 change either way.
The energy savings are also work done X mas flow rate, again no 1:1 change.
Another energy saving benefit is greater mechanical efficiency of your compressor. The closer your suction and discharge pressures are, the more efficient your system. The 3 % is a rule off thumb, and will vary from refrigerant and type of compressor, with reciprocating compressors benefiting the most
Coolstuf, I'm going to be nice to you and simply say that you have some clue as to what you are talking about but like the Mr Jim Rusling of Mitsubishi Electric who wrote that article for the IOR Service Section you know just enough to be dangerous.Quote:
Originally Posted by coolstuf
Here's the first of a series of articles I wrote on the matter from the angle of practicalities impacting on the service engineer. It might be a good start for your ventures on the topic. This series was published almost a decade back now. Subsequent to this article series BSRIA did some tests on a chiller rig and published their findings - not about the extended practicalities - but about the fundamental principles. I have a copy of their publication on the experiments they did but firstly it is quite simplistic and secondly you have to pay them for copies.
BTW - I have not read this article before posting it here and as I say I wrote it nearly 10 years ago so I can't remember exactly what it was saying - but if you have questions then fire away.
The article --->
This is the introductory article of series I intend to use for the discussion of the practicalities and their difficulties implicated in the quest for the application of more efficient refrigeration systems. Especially with regards to the application and commissioning of reduced head pressure direct expansion configurations.
There is a family of trouble symptoms listed toward the end of this article that are repeatedly seen and it is my hope that this article series will assist others somewhat in their attempts to eliminate these troubles and provide greater operating stability to these reduced head pressure systems that are still often bringing more grief than savings.
For energy savings purposes it has become quite commonplace for refrigeration system designers to select their system components, controls and control settings to permit reduced head pressure operation during lower than design ambient temperatures. Ordinarily, peak design ambient temperatures of 28°C are used. However, it is well known that for the vast majority of the year ambient temperatures in the UK are well below 28°C. In fact, for more than half the year the UK’s ambient temperatures are around the 10°C mark. The average UK ambient temperature might even be that same 10°C.
When we are talking about reducing system head pressures with ambient temperatures for energy savings we are talking about optimising physical principles focused mostly around the compressor and condenser area. The combination of compressor and condenser is often called a condensing unit or with larger systems with multiple compressors and multiple fans the combination might be called a pack system. Not that a water-cooled condenser and compressor combination isn’t called a condensing unit, it’s just that I will be mostly discussing air-cooled systems in these articles and want to make it clear that the energy savings we are attempting to make are associated primarily with the compressors and condenser fans.
It is easily understood that the energy consumed by a condensing unit or pack system is chiefly by the compressor and fan motors. Theoretically, both are consuming electrical energy to provide the mechanical energy required to process gases. In these articles the gases being processed will be R22 by the compressors and then air by the fans until it’s time to move of the zoetrope’s. The compressors are of course both displacing and compressing the R22 whereas for simplicity we assume the fans are merely displacing the air and so in calculations we treat the air as non-compressed.
To calculate the power consumed when either compressing the refrigerant vapour or displacing the air, two slightly different formulae are used but which are both based on the same fundamental principle that if:
Force = Mass x Acceleration,
and,
Work = Force x Distance,
but,
Force = Pressure x Area,
then,
Work = Pressure x Area x Distance,
and since,
Area x Distance = Volume,
we can conclude that,
Work = Pressure x Volume.
For fans then, the formula used is:
kW = (P x V)/h
where h represents the fan’s efficiency factor which is commonly estimated to be 0.63.
That’s about as far as we would need to consider things regarding air but for processing refrigerant vapours it has to get a little more complicated because as the piston moves to compress the vapour the vapour’s pressure will be continuously increasing rendering further piston travel increasingly more difficult and so the Pressure x Volume formula has to be modified to reflect this by considering a pair of added factors namely pressure ratio and the vapours relative specific heats:
kW = P1 x V x y/(y-1) x [(P2/P1)(y-1)/y – 1]/h
Consider an R22 air-cooled water chiller with a summer design cooling capacity of 100kW, this is a refrigeration system comprised merely of condensing unit and attached water heat exchanger. If we look at this water chiller’s relative compressor and condenser fan power consumptions at two different ambient temperature conditions with both fixed and reduced head pressure philosophies, we get a feel for the energy savings available.
First scenario, the design scenario, during the design summer ambient temperature of 28°C when the compressors and condenser fans are all running fully loaded we might expect a total system power consumption of around 25.7 kW where the compressor is consuming 22.7kW and the fans 3kW. Independent of any head pressure control philosophy, this scenario will always occur at peak summer design temperatures.
Second scenario, during the average 10°C ambient condition we maintain head pressures by slowing the fans. With the fans slowed to maintain the designed condenser log mean temperature difference (LMTD), their power consumption might now be a lower 0.7kW while of course the compressor power consumption would remain at the design summer level of 22.7kW. This system would be saving a mere 2.3kW during the cooler 10°C ambient temperatures in the form of reduced fan energy consumption. There are other reasons yet to be discussed but almost the sole reason for maintaining the designed summer head pressure all year round is so that the systems TEV’s are receiving liquid at their designed inlet pressure otherwise they could not provide for the required capacities.
The third and now more fashionable energy savings scenario is the most efficient and preferred head pressure control philosophy which leaves all the fans running right up until the very minimally required head pressure is reached. Engineers are constantly seeking ways to reduce this minimum required head pressure by reducing the minimum operating pressure drops required by each individual component found lying between the compressors valve plate and the evaporator inlet. If all the fans are running at full speed during a 10°C ambient condition then to affect the design capacity of 100kW the compressor power consumption would drop to around 10.7kW while the fan power consumption would remain near the original 3kW although perhaps increase very slightly with the colder air’s increased density.
The power consumption figures given above were calculated using the formulae given above but when it comes to working with formulae in the analysis of varying scenarios, I prefer to keep them, as best I can, in the proportional expression form. This form doesn’t always allow for the calculation of specific unknowns but does help predict the values that we might want to see as would occur in scenarios different from any known base scenario. By analysing the vapour compression formula above one can see that the bulk contributor to compression power for a particular gas with a fixed suction pressure would the ratio of absolute pressures otherwise known as the compression ratio. Then of course another contributor would be the amount of vapour needed to achieve the duty, which is inversely related to the refrigerant’s net refrigeration effect. And so I concocted the following proportional expression:
KW2 = kW1 x r2/r1 x NRE1/NRE2
If the operating summer design parameters have a saturated suction of 2°C and a saturated discharge of 43°C then the summer design compression ratio would be 3.12. While with a liquid temperature of 35°C and suction vapour temperature of 6°C the refrigerants (R22) net refrigeration effect would be 166.8kJ/kg.
During the lower 10°C ambient conditions with all fans running at full speed the modified parameters would hopefully still have a saturated suction of 2°C but would expectedly have a saturated discharge of 23.5°C which reduces the compression ratio to 1.89. While for the new liquid temperature of 18°C and maintained suction vapour temperature of 6°C the refrigerants net refrigeration effect would have increased to 188.8kJ/kg.
The comparative compression ratios and net refrigeration effect figures above can then be placed into my proportional expression which facilitates easy estimates the modified power consumption:
KW2 = kW1 x r2/r1 x NRE1/NRE2
\ = 25.7 x 1.89/3.12 x 166.8/188.8 kW
\ = 13.8 kW
Interestingly 13.8kW is 0.1kW above the values I already calculated previously when using the relevant formulae but then when doing so I hadn’t considered the air’s increased density at the lower ambient temperature and it’s effect in increasing fan power consumption, which my proportional formula seems to have attempted anyway, by accident of course. None the less, the proportional formula is still sufficiently illustrative in showing that condensing unit energy savings are mostly affected by reduced compression ratios and increased refrigerant net refrigeration effect and that the savings available are substantial.
The difference in energy consumption between the constant head pressure and reduced head pressure control philosophies during 10°C ambient temperatures, 23.4kW - 13.8kW = 9.6kW, is the intended savings.
So we know there are savings in running cost to be made with reduced head pressure operation and that they are substantial savings, however, it is quite evident from my troubleshooting ventures that system design and commissioning tends to be a little more complicated and it is the nature of those difficulties that I will attempt to tackle in the next few articles.
There is a family of symptoms related to the improper application or improper commissioning of reduced head pressure systems, they are all related to each other and form a kind of self-perpetuating domino effect. These are the same symptoms I have come across on a number of troublesome sites attended in recent years. The family of symptoms reported comprise mostly:
· Uncontrollable head pressure fluctuations around everywhere but the intended lowest head pressure setting.
· Excessive condenser fan cycling or ramping in response to the above.
· Excessive receiver liquid level fluctuations.
· Inconsistent liquid quality in the liquid line’s common sight glass
· Constant TEV hunting and attempts to adjust them are fruitless.
· Compressors continually loading and unloading with no sign of any settling.
· Compressor and reservoir oil levels constantly low or even tripping.
When deciding what a systems minimum head pressure is going to be one would work their way backwards from the systems highest saturated evaporating temperature, looking at all the pressure drops required for stable operation of each component found between the evaporator inlet and all the way back upstream to the compressor valve assembly. The components have to operate not just with stability but have to maintain their design capacity.
Over the next few months in this article series I will consider the nature of the witnessed problems experienced with each of these components, as I understand them in relation to reduced head pressure operation.
Ok, impressive article, and you are right.:)
I was talking about air cooled condensers (should have said so), where good maintenance and installation practice can lower the condensing temp without using more energy.
By lowering condenser tem-re, the suction tem-re will drop as well. Is any data what will happen for each deg. of condensing tem-re with suction tem-re. My understanding is that the suction tem-re will influence the energy consumption more that condensing tem-re.B”H
“every 1 deg C lower condensing temp, you use 3 % less energy and you gain 3 % capacity” seems as a high number.
:off topic:Quote:
Originally Posted by DTLarca
I thought that we are talking about ammonia industrial refrigeration. Am I wrong?
Daikin. Do they manufacture screw ammonia compressors? I know Frick, FES, Mycom. Daikin?:confused:Quote:
Originally Posted by DTLarca
What's your point? :)Quote:
Originally Posted by Segei
What's your point? :)Quote:
Originally Posted by Segei
Your graph may be good for commercial or domestic refrigeration, but you can not apply it for industrial refrigeration. There is difference between home fridge and ammonia industrial refrigeration plant that has 10+ evaporative condensers, 30+ screw compressors and 100+ evaporators.;)Quote:
Originally Posted by DTLarca
Why are you unable to apply it to industrial refrigeration?Quote:
Originally Posted by Segei
I am a qualified industrial refrigeration mechanic with a good number of years spent on industrial ammonia systems and I find it easy to apply the graph - in fact I find it impossible to try make the graph not fit.
You on the other hand obviously have a problem with the graph. Could you explain to me why you have a problem with the graph?
Perhaps if you show me where you are having a problem I can help you - otherwise I must assume you are just "making it up".
All the data I have applies to systems from 2.5kW to 9MW - I have no reason why it should not also apply to industrial systems.
What is the main reason you think it would not apply also to industrial systems?
The changes would primarily be net refrigeration effect and compressor volumetric efficiency. The head pressure rises then so does the suction on account of a small increase in flash gas, a small reduction in the width of the saturated envelope and a small reduction in the compressors volumetric efficiency. This assumes we remain with the range of design feed device capacities.
Do you really believe that you can apply this graph to industrial refrigeration? Probably, I can not help you but I'll try.Quote:
Originally Posted by DTLarca
Major reason is number of variables. For simple plant when ambient temperature increase heat transmission load increase, condensing pressure will increase, efficiency of compressor will increase.
Industrial plant variables are wet bulb temperature of ambient air, wet bulb approach, condenser sequence, compressor sequence, compressors' part load operation at different compression ratio, optimum suction pressure, optimum intermediate pressure, optimum set points for summer and winter operation, minimum allowable condensing pressure, optimum defosting...... Recently PhD paper was written regarding to optimum defrosting. I don't think that it covered all areas of this complicated issue.
Actually you can help yourself. We have subforume "System optimization". Probably, you can get a lot of information about this issue. I remember that we had good professional discussion with US Iceman. I didn't hear from him for a while. Does anybody have information about US Iceman?
Segei, you are on a losing wicket from the outset.Quote:
Originally Posted by Segei
3 things:
1) When it comes to the science of making comparisons a key philosophical tool must be invoked - look up Ceteris Paribus.
2) The matter must be resolved with demonstrations using figures - not testament.
3) You should go back and look more closely at what I have already said prior to my posting a copy of my old article.
And by the way - because I do not rate US Iceman you are setting up a little war for which I will have to be barred from this board because I never turn away from a technical challenge :)