Turbocor 1100kw water cooled

Hi Guys,
looking to design a Turbocor installation but got a few questions.
To take advantage of the Turbocor high COP, we would need to run the chillers at 25% capacity where possible. If I run 5 machines at 25% I am possibly running a lot of primary pump volume. Can I easily go chilled water variable volume on the turbocor chillers so we can get saving on the pump power using VSDs?
Also, how would you sequence the machines to best take account of the 25% load high COP without getting over complicated?
Not sure if there is a lot of hype going on here with Turbocors. Screws seem comparable at 75% load. We are achieving COP of 7.6 at 75%. Not sure what we would get on Turbocors but I suspect not a lot more.

Have you got 5x 1100kW or 5x 220kw?

Anyway, your evaporator must be able to achieve a good exchange at low flow rate, if you think to reduce the water flow to 1/5th.

And how do you plan to connect the 5 chillers? Did you consider the bypass rate into the standby chillers?

Finally, in my opinion, the COP of the turbocor doesn't vary much, if evaporating water temperature and condensing water temperature are constant. In fact, the coefficient of performance of turbocor compressor is (grossly said) independent of the speed. I don't think you will experience much difference from 5 20% chillers to 1 100% chiller in terms of power input. Of course, you will benefit from a greater exchange surface when refrigerant flow rate reduces, but I think not much more than few K, provided you will still be in the envelope of the compressor.

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Originally Posted by Gun

Hi Guys,
looking to design a Turbocor installation but got a few questions.
To take advantage of the Turbocor high COP, we would need to run the chillers at 25% capacity where possible. If I run 5 machines at 25% I am possibly running a lot of primary pump volume. Can I easily go chilled water variable volume on the turbocor chillers so we can get saving on the pump power using VSDs?
Also, how would you sequence the machines to best take account of the 25% load high COP without getting over complicated?
Not sure if there is a lot of hype going on here with Turbocors. Screws seem comparable at 75% load. We are achieving COP of 7.6 at 75%. Not sure what we would get on Turbocors but I suspect not a lot more.

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COP's will be in the region of 11/1

You will have some room for play on the volume flow - fit an inverter and experiment.

The energy saved with speed reduction is enormous.

dKE = 1/2 m (vmax² - vmin²)

The energy saved is proportional to impeller tip speed squared.

Gas volume flow is proportional to impeller speed. But the energy per volume of gas and thus the gas's final pressure is proportional to tip speed squared

So even if you have a low load requiring just 75% speed which gives 56% pressure lift - still, if you have warm condensing conditions then you are in trouble because at 76% speed you can only achieve 56% lift and so you have to speed the impeller up anyway to overcome the high head pressures. So you cycle one other compressor and then raise the speed of the remaining and so on until a balance is achieved of speed (for lift) versus speed (for duty).

The best COP I have seen being acheived, because of speed reduction, is 15/1 - though I hear better has been achieved elsewhere.

Hi Gun
You need to check manufacturers curves. I have seen COP of 18 under optimal conditions. They are most efficient at around 67% load. they are much better at lower ambients due to elevated condensing temps on other compressor types for oil return. best to look at total system efficiency as opposed to COP which includes pumps and fans. Load matching and rough calculation/measurement of building U value at varying ambient conditions will achieve lowest power consumption.

I don't know how COP in the range of 15 or 18 is even possible, but just to be sure we are speaking the same language, attached a graphic obtained by Danfoss online selection software for Turbocor compressors for a saturated suction temperature of 8°C.

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Originally Posted by NoNickName

I don't know how COP in the range of 15 or 18 is even possible, but just to be sure we are speaking the same language, attached a graphic obtained by Danfoss online selection software for Turbocor compressors for a saturated suction temperature of 8°C.

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Floating suction and discharge pressures from floating chilled set points along with minimising condenser water temperatures.

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Originally Posted by DTLarca

Floating suction and discharge pressures from floating chilled set points along with minimising condenser water temperatures.

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Would you please make an example, with printouts from Turbocor selection software? Air conditioning application, please, as it seems the case here, nominal capacity 12 to 7°C and Summer in southern europe (38°C) or continental asia. Feel free to choose air or water cooled.
.

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Originally Posted by NoNickName

Would you please make an example, with printouts from Turbocor selection software? Air conditioning application, please, as it seems the case here, nominal capacity 12 to 7°C and Summer in southern europe (38°C) or continental asia. Feel free to choose air or water cooled.
.

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What happens when your condensing water is 6°C and your chilled water has been reset to 10°C?

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Originally Posted by DTLarca

What happens when your condensing water is 6°C and your chilled water has been reset to 10°C?

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I would divert the condensing water into the chilled water circuit and stop the compressor. EER = +infinite

Anyway EER is evaluated at design conditions for the application. Not in Finland in January during christmas holidays.

I assume you dropped the argument, DTLarca?

We have 5 machines at 1100kw. chilled water at 12/18 degC. condensing 29/35 degC

Hi Gun
Look at the life cycle costs capitol and running, allow for energy cost increase. If you get some degree day cooling data for the location and work out the ratios of condensing temps you may end up with a combo of chillers to allow for base load and low load conditions. I could imagine a 1/3 low load being Turbocor and 2/3 for base load being big centrif like trane. Or you could go for mini tri gen with tiny absorption. Tri gen keeps the operating cost low - the ones I have seen use screws.

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Originally Posted by NoNickName

I assume you dropped the argument, DTLarca?

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No, you did not come back to me with the efficiencies you found for those temperatures I gave you. Once you do that I will ask you to then consider the unloaded efficiency data too.

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Originally Posted by DTLarca

No, you did not come back to me with the efficiencies you found for those temperatures I gave you. Once you do that I will ask you to then consider the unloaded efficiency data too.

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I answered. Your scenario is unlikely in any real life plant. It is stupid to start a compressor for simply separating to water circuits where the heat sink is colder than the heat source.
The whole setup is nonsense and again, comparing apples with apples, the performance data should be declared at nominal conditions, not at meaningless or marginal situations. I will not even bother to check the EER at your conditions.

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Originally Posted by NoNickName

I answered. Your scenario is unlikely in any real life plant. It is stupid to start a compressor for simply separating to water circuits where the heat sink is colder than the heat source.

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Oh, we do it often.

The question is about the conditions under which high efficiencies are achieved with turbocore compressors. The question is not "of all possible systems, including turbocore and free cooling and any multitude of other methods, which is most efficient and which is less nonsensical?" The question is simply a matter of what efficiencies are practicably possible with turbocore compressors.

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Originally Posted by DTLarca

Oh, we do it often.

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Shame on you. It's bad for your reputation.

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Originally Posted by DTLarca

The question is simply a matter of what efficiencies are practicably possible with turbocore compressors.

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You're belatedly avoiding the question once again.
At the unlikely scenario of yours, efficiencies are not an exclusive matter of turbocor compressors. Any other type of compressor would have similar EERs. Any other smarter designer, would even shut the compressor off and exchange heat between the sink and the source, or even divert one on the other, if the fluids are compatible.

You asked if I had dropped out of the discussion. I have not - I'm just going to wait for your figures now on turbocore system efficiency at those conditions. Sure, if you don't like those conditions, which can happen for a day or two each year, then choose conditions of condenser water temperatures just above that which occur much more than just a day or two each year. Just choose the lowest condenser water conditions you see fit for UK turbocore application and then do the calcs - lets see what COP's we get.

I already gave you the conditions. If you prefer UK conditions, fine for me. 12/7 or 12/6°C with air cooled condenser, flooded evaporator and microchannel condenser, for your convenience.
I'm afraid you don't access to online turbocor selection software so you're just pretending being an expert, shouting EER numbers with little or no adherence to actual world.

I had a request today from the UK for a water cooled chiller 12/6°C, 43/50°C 30%EG. The actual turbocor EER is 2.98, very near a standard multiscroll water cooled chiller optimised for dry cooler operation, for a much different capital cost.

Oh yes, the ESEER is 7.98, but the multiscroll wins at part load, because with that condenser temperature we are on the left top edge of the turbocor envelope, with full speed, part vane required and no possibility of part load operation.

Let me know when you've done the calcs for 6°C condensing water with 10°C leaving chiller water temperatures :)

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Originally Posted by DTLarca

Let me know when you've done the calcs for 6°C condensing water with 10°C leaving chiller water temperatures :)

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I'm not going to calculate that. Stop whining.

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Originally Posted by NoNickName

I'm not going to calculate that. Stop whining.

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Too funny :)

I'm the one who, speaking from experience, says the multi compressor turbocore chillers have achieved COP's of 11 - you see, for majority of the year, the UK ambient tempertures are below 10°C and lower the further north we go.

You, in return, are the one who is moaning or whining.

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Originally Posted by NoNickName

Oh yes, the ESEER is 7.98

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Well now - if the ESEER is 7.98 what was the best COP? It would have to be pretty much higher than 8 - wouldn't it? Mayber 12?

The TECS-HF 2AI has an EER of 11.81 with 18°C entering condenser water and that is for a lower 7°C supply - not the 10°C supply I am talking about. And in the UK we can have a much lower 12°C supply condensing water temperature for 60% of the year - Note - that is 12°C and not the higher 18°C used by Eurovent for the ESEER calculations you are quoting.

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18°C entering condenser water

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It looks like you don't read or not able to.
UK summer temperature of design is either 27°C or 32°C.
Mind recalculating and air cooled chiller with those conditions?

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Originally Posted by NoNickName

It looks like you don't read or not able to.
UK summer temperature of design is either 27°C or 32°C.
Mind recalculating and air cooled chiller with those conditions?

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You are so funny :)

You are not being consistent - you are implicitly contradicting yourself.

If what you say is correct then Eurovent is WRONG.

Eurovents ESEER and ARI's IPLV are either WRONG or YOU are WRONG. But it is impossible for both YOU and EUROVENT to be correct at the same time on the same matter.

In fact you yourself have quoted the ESEER figure for a particular turbocore chiller which is a figure based on a concept (ESEER) which denies the need to fix head pressure at the summer design conditions.

Modern chiller save money by partially yet still substantially floating head pressures - letting them drop with ambients. This is the advantage turbocore has over screw chillers which even with variable Vi have built-in compression ratios.

Here - check out the ESEER's on these chillers
http://hvacr.vn/diendan/_oldforum/files/TECS_HF_GB.pdf

And while doing so keep in mind that COP is cooling power versus compressor power while EER is cooling power versus system power in and then finally ESEER is the seasonal weighted average EER for conditions standardised by Eurovent.

The COP's then, according to Eurovent, are actually going to be in excess of 11/1.

I have seen even 15/1 being acheived with condenser water temperatures at around 15°C - I have not known any screw compressor system, or scroll compressor system come near this.

The big advantage with parallel centrifs is that when the head pressure is low but the suction gas volume needed is higher you run several compressors at low speed which gives you the volume while saving on the head energy and then when you have a small suction gas volume demand but higher head demand you run fewer compressors faster. You do not have any such facility with screws or scroll though you do with recips except you do not get the capacity from recips foot print for foot print.

Blablabla again water cooled condensers. Blablabla ARI, Eurovent, IPLV, ESEER all in a soup of babbles.
Blablabla floating head pressure, while all I'm asking is a single figure at design conditions.
Blablabla shameless Climaveneta advertising twice in two messages in a row.
I'm giving up.
You don't know what you're talking about, but very good at making noise.

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Originally Posted by NoNickName

Blablabla again water cooled condensers. Blablabla ARI, Eurovent, IPLV, ESEER all in a soup of babbles.
Blablabla floating head pressure, while all I'm asking is a single figure at design conditions.
Blablabla shameless Climaveneta advertising twice in two messages in a row.
I'm giving up.
You don't know what you're talking about, but very good at making noise.

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Trust me on this - chillers do not run all year round any more with head pressures fixed according to the design ambient.

When we buy chillers we are not interested in the COP per se but rather the ESEER. For high base load comfort cooling we don't care what the chillers COP is at design full load because the chiller is hardly ever going to run at design full load. Instead we want to know what its ESEER is and we know that the higher the lowest load with lowest head pressure COP is the higher will be the ESEER. So when we are looking for chillers that might have higher ESEER's we are particularly concerned as to what their highest COP will be and that it is not at highest design load but at the condition the chillers will spend most time in operation.

I prefer to use scrolls over screws when I can because with multiple scrolls your isentropic efficiency stays up at 74%+ regardless of how loaded the chiller is whereas with screws the isentropic efficiency drops as the chiller unloads.

I prefer scrolls over turbocore for data centers because with data centers you have a high base load all through the year requiring the turbocore impellers to run full speed all year which means they also create within the impeller and volute the same head pressure all year and thus consume the same energy all year - unless you choose to have twice as many turbocore compressors as you need for summer full load so that in winter you run them all but at lower speeds for better COP's (11/1) and then in summer when you need the head pressure for the higher ambients you run half the compressors at twice the speed - sort of thing :)

I don't think you understand turbocore nor what are the factors we take into consideration when trying to design an energy efficient system.

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Originally Posted by NoNickName

Blablabla again water cooled condensers. Blablabla ARI, Eurovent, IPLV, ESEER all in a soup of babbles.
Blablabla floating head pressure, while all I'm asking is a single figure at design conditions.
Blablabla shameless Climaveneta advertising twice in two messages in a row.
I'm giving up.
You don't know what you're talking about, but very good at making noise.

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Welcome to my world, Happy to meet you.
Advise DT o'brien is added to your ignore list.:mad:

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Originally Posted by slingblade

Welcome to my world, Happy to meet you.
Advise DT o'brien is added to your ignore list.:mad:

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The original poster talks of running multiple turbocore compressors at 25% to gain from the high COP's. What does this mean? And how does it relate to NoNickName's points made about design summer full load conditions?

Answer those questions and you will realise that NoNickName has completely missed the point.

Plonked now

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Originally Posted by Gun

Hi Guys,
looking to design a Turbocor installation but got a few questions.
To take advantage of the Turbocor high COP, we would need to run the chillers at 25% capacity where possible. If I run 5 machines at 25% I am possibly running a lot of primary pump volume. Can I easily go chilled water variable volume on the turbocor chillers so we can get saving on the pump power using VSDs?
Also, how would you sequence the machines to best take account of the 25% load high COP without getting over complicated?
Not sure if there is a lot of hype going on here with Turbocors. Screws seem comparable at 75% load. We are achieving COP of 7.6 at 75%. Not sure what we would get on Turbocors but I suspect not a lot more.

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Here is an image of what I am talking about.

With scrolls you need to run more in summer to meet the duty.

With Turbocore compressors your idea readily becomes apparent. Because in summer you need fewer compressors for the same capacity and more compressors in winter to have the same capacity but save energy on compression by running more of them slower.

So what I have been saying is that you may not want 5 chillers, depending on the redundancy requirements, but rather one chiller with an infinite amount of compressors mounted on it.

Here is a picture on a slide from one of my level 3 lectures on chiller system, capacity control, efficiency and oil return from flooded evaporators. Remember these systems are oil free so there is no minimum load concerns regarding oil return.

Well I think with all your bickering off-topic has impressed the OP who had the decency to reply to a question but has not had a reply as yet.

If you can't keep on topic then please start it elsewhere.

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Originally Posted by Brian_UK

Well I think with all your bickering off-topic has impressed the OP who had the decency to reply to a question but has not had a reply as yet.

If you can't keep on topic then please start it elsewhere.

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It's always worth while examining the relevant principles - we have to unpack the technical assumptions for each other so that a sharing of knowledge by way of explanation can take place. Very often the best ideas are not the ideas already known to the contributors but instead the ideas developed out of a synthesis of ideas formed after everyone has critiqued each other's ideas - the best option could always be the so far not realised idea.

But this nobel process is always ruined when people get personal. If someone is possibly wrong about something they say all this means is that what they say contradicts some other fact or theory already in wide circulation or already agreed prior in the discussion. This means that if we think someone is wrong we should present as evidence for our argument the principles in conflict with their argument - we should demonstrate how their argument is wrong. Not merely call them names.

We can make personal judgments of people only after we have satisfactorily demonstrated them to be in error but not in place of demonstrating their errors.

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Originally Posted by Gun

We have 5 machines at 1100kw. chilled water at 12/18 degC. condensing 29/35 degC

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If the idea is to run the turbocore units at 25% load at these conditions then making savings could be difficult because the high COP's occur when the impeller speeds are reduced and not when the pre-whirl vanes are closing. At these conditions 25% capacity would likely be achieved by pre-whirl vane closure. Savings would best be realised when the vanes can remain fully open and instead the speeds are reduced to unload - this requires reduced ambient temperatures be used to reduce condensing pressures.

so basicly you will get the best COP when you have 5.5 MWatts of cooling capacity, against a rated load of 2.5 MWatts, and the ambient temperature only needs you to deliver 1.5 MWatts.
Bad investment, sorry.

Here are some old MET Office numbers for southern England. Hours per year at varying dry bulb temperatures. I notice the wet bulb temps for below zero are higher than the dry bulb temps. Is that an error in the placement of the data or is there something to psychometrics that I haven't yet realised?

Anyway - the point is that not since the 70's have we held systems, all year round, up at their design summer condensing temps. For most hours in the UK we can have condensing water temperatures of 16/22 °C when we can easily have COP's of 11/1 with trubocore.

Later too with scrolls when they eventually build them with internal Vi.