Capacity achievement

We have a new refrigeration system under commissioning. There are 2 x 63.9 TR compressors Mycom F8WA for R 22 rated at Te -5C and Tc 45 C. The Evaporative condenser is Baltimore VXC125 has a capacity of 140 TR. The expansion valve is PHT 85 from Danfoss and the supplier says it is orifice 4 giving 130 TR. The PHE after the expansion valve is a Thermowave make TL250 designed to run 6821 kg/hr R22 at -5C and water at 30 m3/hr to cool from 14 C to 2 C. There is a chilled water buffer tank 108 m3 capacity.
There observations are as follows:
1. The supplier mentioned about antifreeze upto 25% in water which was not discussed at the order stage. To get over this issue I have asked the system to be commissioned at 4C instead of 2C as stage I. The antifreeze issue will be taken up after the system runs close to its design load.
2. The return water temperature varies from 8-18 C. The Plate heat exchanger gives a steady 3C temperature drop on the chilled water side at 30 m3/hr at all temperature between 8-18C. At this time, one compressor operates at 13 bar on the hp side and 3.6 bar on the lp side. The load % on the control panel varies between 50-75%. This goes to 100% for a few minutes and gets back to 75% even if the water outlet temperature is 6-9 C. The other compressor will not start unless the first runs on 100% capacity.
3. I suspect the expansion valve to be limiting the quantity of the refrigerant. Is there a way to check the performance of this valve? There is a tiny tag at the expansion valve saying orifice no 1. This corresponds to 30TR according to Danfoss catalogue- which is incidentally the capacity delivered by the plant..

My queries are:
1. What will be the change in capacity of the system if we run it at 4C instead of 2C using same equipment?
2. Is there a way to verify the doubt about the expansion valve?
3. If I insist on changing the expansion valve so I get closer to the designed capacity of the plant what are the pitfalls that I must avoid?


Looking for some enlightenment on this matter.:confused:

Regards,

Hemant Anand

Hi,

Can you provide some additional information please?

Are the high and low pressures in gauge or absolute pressure?

What is the design air wet bulb temperature in your area?

In my explanation below I am assuming the pressures are in gauge and the expansion valve is used for direct expansion feed into the PHE. I am not familiar with the Danfoss PHT 85 valve.

The PHE data specified should provide a capacity of about 119 Tons. The two compressors equal about 128 Tons at the specified operating conditions for R-22. That looks OK.

What are the compressor motor sizes? Do they match the required power input for the specified operating conditions.

I question the use of an evaporative condenser operating at 45C. Most of the evaporative condensers I have seen are selected for no more than 35C. The problem with using this high of a condensing temperature can contribute to excessive fouling on the condenser tubes. A very good water treatment system will be required to control the fouling and scale formation.

The refrigerant mass flow rate of 6,821 kg/hr only produces about 79 Tons of cooling if we assume the refrigerant liquid temperature feeding the TXV is at 45C.
(direct expansion feed with -5C evaporating temperature)

If the system capacity is equal to 118.9 Tons, the mass flow rate should be about 10,233 kg/hr. (This is based on the -5C evaporating temperature and 45C condensing temperature. I'm not sure why the PHE refrigerant flow rate does not balance with the 10,233 kg/hr to match the equivalent water cooling capacity.:confused:

The operating compressor at 50-75% capacity appears to be loading and unloading on suction pressure. This also seems to confirm the the high water outlet temperature. (not enough cooling, or low refrigeration capacity)

Does the control system have a unloading capability if the discharge pressure is too high? This does not appear to be related to the compressor capacity since the discharge pressure is only 13 bar.

If the expansion valve orifice is too small, this can contribute to the operation as you describe it. Not enough liquid R-22 is flowing into the PHE. It also appears that this is confirmed by the number 1 orifice tag you stated is present, when the supplier said the orifice should be a number 4.

One way to confirm the TXV performance is to check the evaporator superheat. My guess is the superheat is higher than it should be.

If you can confirm these points perhaps I can provide additional information.

My first impression is the TXV orifice is too small (low cooling capacity with higher than desired water outlet temperatures). The compressor is only loading to maintain the suction pressure. Therefore, the low capacity is due to low R-22 liquid supply to the TXV.

Any change in outlet water temperature will affect the compressor capacity after the TXV orifice size is corrected. The supplier should be able to tell you the new capacity at the new operating conditions.

Since this is a new installation, I would make the supplier correct any and all deficiencies in the system.

Hope this is useful. I am making some assumptions, but it does seem like the TXV capacity is the problem.

My comment ,only for reference-:)

1. What will be the change in capacity of the system if we run it at 4C instead of 2C using same equipment?
# run at 4C,you mean from 14C to 4C，the capacity is decrese as the loading of evap decrease;
2. Is there a way to verify the doubt about the expansion valve?
# you can confirm the evap capacity by checking the chiiled water volume and temp difference,if it meet the reqiurement,no reason to doubt it;as Iceman said ,the refrigerant flow volume may be some problem.What you gived 6821kg/hr is not correct,reconfirm it.
3. If I insist on changing the expansion valve so I get closer to the designed capacity of the plant what are the pitfalls that I must avoid?
# carefully match the TEV capacity with the refrigeration system at the same working condition and leave 15&20% extra for buffer

BTW,the evaporative condenser is not working at 45 Tc ,if this ,no obvious advantage than air cooled condenser. i think it's at least 40C Tc which will 0.4 higher COP than 45C Tc.

rgds
LC

Thanks for spending your time to give this information.

The additional information is as follows:
1. The figures are for guage pressures.
2. Wet bulb temperature is 29C. In fact the evap condensor water blowdown is at 29-30C over last 2 weeks of peak summer. This cools the gas to about 35 C ( to be measure at next trial).
3. Each compressor motor is 90 kW.
4. The motor capacity and evaporator superheat will be checked at the next trial.

Thanks a million for your help.

Regards,

HKA.

I reviewed the compressor performance data to verify the information.

The compressors you mentioned have the following performance characteristics with R-22:

At 45C/-5C @ approximately 1380 RPM
Capacity = 63.9 Tons
Power = 66.6 kW

At 35C/-5C @ approximately 1380 RPM
Capacity = 73.1 Tons
Power = 59.3 kW

Based on this information I would say the motor sizes are sufficient for the requirements.

The 29C wet bulb temperature will have a large impact on the condenser capacity. I could not find the BAC model number (in my literature) you specified. However, if the condenser was sized for 29C entering air wet bulb temperature and 35C condensing, I think the condenser will be short of heat rejection capacity.

If the Tc=45C, the condenser may reject the heat with 29C entering WB, if the condenser coil is clean and no discharge air (exhaust air from the condenser) recirculation is occurring. With the entering WB temperature being this high, any exhaust air recirculation will quickly cause a large loss of heat rejection capacity.

If the water blowdown is being controlled by water temperature (I have never seen this done before:confused: ), I suspect the large quantity of make-up water is assisting in keeping the discharge pressure down.

In addition, the fact that only one compressor is running most of the time at 50-75% capacity with the discharge pressure at 13 bar would seem to indicate the condenser capacity will be insufficient when both compressors are operating.

Pertaining to your question about glycol: Can the PHE have additional plates added to it to increase the evaporating temperature? If glycol is added, the heat transfer for the PHE will decrease. This means you would have less capacity available for cooling. The pump input power will also increase.

If glycol is added I normally suggest the percentage of glycol is adjusted to provide a freeze point 3-5C below the evaporating temperature of the refrigerant.

I would recommend a flow switch be installed in the inlet water pipe to the PHE. This would be wired into the control circuit to shut the compressors OFF, if the water flow is stopped for any reason (pump failure, etc.).

I also think you should have a freeze-stat installed. This is a temperature switch that is wired into the control circuit to shut the compressor's OFF, if the water temperature gets down to 0-1C. The bulb for the freeze-stat is placed in the outlet water line from the PHE. Some extra safeties to reduce potential equipment failure.;)

It sounds like you have several issues to deal with. I recommend you get the supplier to assist you in making the necessary corrections for the PHE and expansion valve.

Good luck with your project. Let me know how it develops.

Best Regards,
US Iceman

From what I know of the PHT Danfoss valves they use the pulse width controllers which provide very good turn down ratios (I have also used this technology in furnace burners successfully), therefore raising the size of the expansion valve should not pose too many problems (but this depends on your operating conditions of course).
Hopefully the condensing temperature does not get as high as 45C, I agree with previous that 35C max is more normal.
Supprised that you were not warned about using Glycol in this system at the operating temps you are looking at, especially if 2C was desired, even at 4C I would be careful. Is there a problem with putting Glycol in system and running at your required 2C? Although running at 4C will reduce your electrical consumption significantly.

Also agree with previous, check the superheat.

Ask manufacturers about running at reduced head pressures in winter to save energy, this may be a possibility with this system.

Regards
Darren

Daz brings up a good point on the expansion valve. Since it appears that it will have to be revised, now is a good time to verify the orifice size for two conditions; summer and winter operation.

The pressure difference between the evaporating and condensing pressures will be the maximum differential for the valve orifice.

In a rough approximation, the orifice should be sized for the capacity (mass flow) required at the summer design differential and also the winter design pressure differential.

A pulse width valve should be able to handle this, if everything is sized well.

More things to consider...

Hi People,

After a long wait the necessary spares have arrived and we have now changed the expansion valve to PHT 125. Yesterday, I received another clarification from the manufacturer of the heat exchanger - this system is supposed to be a gravity fed system and not a DX chiller. The way it is installed now is that the expansion valve feeds the heat exchanger directly. According to the contractor, using some glycol will permit the system to operate at the required capacity but use about 2-4% more refrigerant throughput.

Does any one have any experience on using a gravity fed PHE in a DX type installation?
Theoratically - does this sound correct? I think the residence time in the PHE is too small to allow for necessary heat transfer.
The options before me are:
1. Allow the contractor to prove his point ( a variance of 5% on the heat capacity may be acceptable for non technical reasons - shut down etc)
2. Insist that the PHE is installed in the righ way - hence a different setup, controls etc
3. Insist that he changes the PHE to match the set up

Any cues for me?

regards,

Some of the plate heat exchanger manufacturers use a refrigerant distributor for DX feed. If yours was designed for gravity flooded feed, perhaps you can ask the plate manufacturer if they have a retrofit device to convert this.

You should also ask the manufacturer what impact this can have on the capacity of the plate heat exchanger.

You asked about several options:

Quote:

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1. Allow the contractor to prove his point ( a variance of 5% on the heat capacity may be acceptable for non technical reasons - shut down etc)
2. Insist that the PHE is installed in the right way - hence a different setup, controls etc.
3. Insist that he changes the PHE to match the set up

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I would vote for #2. If you or your firm paid for the original specification using gravity flooded evaporators, you should insist on getting what you are paying for.

Personally, I like gravity flooded for these applications much better than DX.