AWHP superheat & sub-cooling

[Gary]

The trade is going to be between a pool-boiling design & convective boiling design - I'm going to have to think about this a little more.

Pool boiling would interfere with oil return. We need to maintain velocity.

[desA]

Pool boiling would interfere with oil return. We need to maintain velocity.

That's very true - hadn't thought through those implications carefully.

It will have to be convective at minimum oil carry velocity - 5 m/s vertical seems to be the suggested minimum.

Thinking more on the VIC(E), I'm wondering about the suitability of a tall, slender PHE, for this application. I actually have enough to ask my supplier to give me an idea on this, if it seems sensible.

Now that I think more on this, I seem to also remember that there may be something in the Alfa-Laval technical book on similar applications, where this could suit the VIC(E).

[desA]

Condenser sub-cooling re-visited

Some pages back, an idea was proposed to develop a ratio of condenser sub-cooling to TD, much along the lines of the Magoo evap tuning rule [SH,e=(0.6-0.7)*TD].

Would you perhaps have any thoughts on a suitable ratio for the condenser SC/TD?

From the current lab experimental data, this value (tube-in-tube) seems to shift from start-up (e.g 30%) to hot condition (e.g. 40%), but remain in a reasonable range, if the system charge is correct.

[Gary]

Have we tried different charge levels to see what subcooling works best?

[desA]

^ I can add, or remove charge, as we require - that's the easy part.

The difficult part will be once the machine configuration looks to be reasonably settled, then performing a full charge versus performance study, with varying charges.

What I was looking for is a yardstick, around which to begin.

As an example, on one machine, with a different condenser type, in the range of calculated refrigerant charge, the SC/TD ratio seems to be around 0.3-0.4. This pretty much corresponds to the original theoretical design estimates.

Now, for the current lab machine, with a tube-in-tube condenser, I'm not sure if this type typically responds with a greater allowable SC/TD, by nature of its long condensing length.

I'd always been under the impression that tube-in-tube condensers typically allow more SC than for instance plate condensers.

[Gary]

As an example, on one machine, with a different condenser type, in the range of calculated refrigerant charge, the SC/TD ratio seems to be around 0.3-0.4. This pretty much corresponds to the original theoretical design estimates.

Now, for the current lab machine, with a tube-in-tube condenser, I'm not sure if this type typically responds with a greater allowable SC/TD, by nature of its long condensing length.

I was thinking more along the lines of 0.6-0.7 * TD for everything.

[desA]

^ From what I'm seeing so far, it seems that the SC/TD ratio will probably vary between condenser families, as well as rise some % from start to end of heating cycle.

The correlation to TD looks to be fairly sound, so far.

[Gary]

All that will go out the window with the addition of the water regulating valve, because any measure of subcooling assumes full water flow.

We may end up putting a sightglass on this system.

[Gary]

Here is how I would determine the ideal charge:

Assuming the primary purpose of this system is to fully heat feedwater before sending it to the storage tank:

Further assuming that we want to keep the Te,sat at 13C and the Tc,sat at 75C:

I would empty the tank and hook up the feedwater supply to the condenser water inlet.

Start the system and adjust the water outlet valve to where it maintains Tc,sat=75C.

Adjust the fan speed to where it maintains Te,sat=13C.

Every 15 minutes add refrigerant, adjust Tc,sat and Te,sat if needed, then record all data.

The ideal charge is that which gives us the lowest condenser approach temp (hottest leaving water).

We can then see what condenser outlet subcooling works best for the condenser.

[desA]

All that will go out the window with the addition of the water regulating valve, because any measure of subcooling assumes full water flow.

We may end up putting a sightglass on this system.

Can I ask you to explain further?

[Gary]

Can I ask you to explain further?

If the water flow is controlled to maintain Tc,sat=75C, the TD when there is cold water at the condenser inlet is going to be huge. We may not be able to use a charging formula based upon SC as a percentage of TD... but we need to test this.

I would be very reluctant to put a sightglass on this system because service techs everywhere are going to charge to a clear sightglass plus, and will totally ignore the weigh-in charge.

[desA]

Here is how I would determine the ideal charge:
Assuming the primary purpose of this system is to fully heat feedwater before sending it to the storage tank:
Further assuming that we want to keep the Te,sat at 13C and the Tc,sat at 75C:
I would empty the tank and hook up the feedwater supply to the condenser water inlet.
Start the system and adjust the water outlet valve to where it maintains Tc,sat=75C.
Adjust the fan speed to where it maintains Te,sat=13C.
Every 15 minutes add refrigerant, adjust Tc,sat and Te,sat if needed, then record all data.
The ideal charge is that which gives us the lowest condenser approach temp (hottest leaving water).
We can then see what condenser outlet subcooling works best for the condenser.

Thanks for this. It seems to be a very cunning way of getting the system settled on a once-through (direct) basis.

I'll have to adjust my current water piping, to implement this (currently a pump-around loop), but this is no problem, as it really only needs a tee, a valve & a few connectors to set it up. I'll get the bits in today & modify to suit.

This will get us set for the direct heating option. I predict a very low water flowrate for this option.

Pump-around loop
How would we go about determining correct system charge under a pump-around scenario, with typical dTw in the region of 2-3K? Water lifts from around 30'C to around 65'C exit in final pass. Total circuits around 11-12. This is the common scheme - alternative to the direct method.

[Gary]

Pump-around loop
How would we go about determining correct system charge under a pump-around scenario, with typical dTw in the region of 2-3K? Water lifts from around 30'C to around 65'C exit in final pass. Total circuits around 11-12. This is the common scheme - alternative to the direct method.

Why would anyone let their tank temp drop down to 30C?

I think it might be very helpful to have a piping diagram of the "typical" system you are referring to. My idea of typical may be entirely different from what is typical in your area.

[Gary]

Here is how I would determine the ideal charge:
Assuming the primary purpose of this system is to fully heat feedwater before sending it to the storage tank:
Further assuming that we want to keep the Te,sat at 13C and the Tc,sat at 75C:
I would empty the tank and hook up the feedwater supply to the condenser water inlet.
Start the system and adjust the water outlet valve to where it maintains Tc,sat=75C.
Adjust the fan speed to where it maintains Te,sat=13C.
Every 15 minutes add refrigerant, adjust Tc,sat and Te,sat if needed, then record all data.
The ideal charge is that which gives us the lowest condenser approach temp (hottest leaving water).
We can then see what condenser outlet subcooling works best for the condenser.

Since the secondary purpose of this system is to maintain the tank temperature, once the tank is full we can connect the tank back to the water inlet and see if the charge is right for this purpose as well. Again adjusting the water flow for Tc,sat=75C and the fan for Te,sat=13C, the best charge is that which gives us the highest dTw.

If these two ideal charges are different, then splitting the difference gives us the optimum charge for the pump around loop system.

IOW, if one charge is ideal for say W,in=30C and the other charge is ideal for say W,in=60C, then we would need to charge it in between for going from 30C to 60C.

[desA]

Originally Posted by desA
Pump-around loop
How would we go about determining correct system charge under a pump-around scenario, with typical dTw in the region of 2-3K? Water lifts from around 30'C to around 65'C exit in final pass. Total circuits around 11-12. This is the common scheme - alternative to the direct method.

Gary:
Why would anyone let their tank temp drop down to 30C?

I think it might be very helpful to have a piping diagram of the "typical" system you are referring to. My idea of typical may be entirely different from what is typical in your area.

Good comment.

An example:
Take an establishment which only uses most off its water (21 m3) during a peak time of 19h00 - 21h30. Hot water storage tanks are extremely pricey in this region - more than the heat-pumps, for instance.

Current system as follows:
1. Small inventory of hot water, via small fuel-fired boiler - pump-around. Start of cycle 30'C - end ~65'C - hold when hot. Enough for off peak usage.
2. At peak usage, hot water storage not enough, so heat mains water directly from 30'C to 65'C on the fly - for use during peak time only.
3. Cost of fuel incredibly high, need to go to alternative energy source.

Proposal (by consultant):
4. To install a heat-pump in either of two configurations:
4.1 Direct heating of small volume flow of water, through heat-pump (direct), into hot-water storage tanks - sufficient for peak usage. Heat-pumps operate off-peak, taking advantage of off-peak electrical rates. Fuel-boiler on make-up call if water storage begins to run short.
4.2 Pump-around heating of larger volume flow of water, through heat-pump, into hot-water storage tanks - sufficient for peak usage. Heat-pumps operate off-peak, taking advantage of off-peak electrical rates. Fuel-boiler on make-up call if water storage begins to run short.

Now, which option is better to use - bearing in mind that the COP,hp for the direct system at Tc,sat~70'C of 3.08, whereas the average COP,hp over the range Tc,sat=40'C to 70'C is close to 4.96 (theoretical)? The COP's improve further if the hot water is held at say 55'C, or 60'C instead of 65'C.

The average power output of the heat-pump over the range Tc,sat=40'C to 70'C is around 9.1% higher than that at Tc,sat~70'C.

--------

I'd be very interested in how things are done in your part of the world.

[Gary]

An example:
Take an establishment which only uses most off its water (21 m3) during a peak time of 19h00 - 21h30. Hot water storage tanks are extremely pricey in this region - more than the heat-pumps, for instance.

Are the storage tanks pressurized?

Are there pumped loops to keep the water hot at point of use?

[desA]

Are the storage tanks pressurized?

As I understand things, many operations have the water storage tanks on the roof & then gravity feed down to the lower floors, except for the booster pump down to the upper few floors.

Some of the other places may be different.

I'd like to know more about how the pressurised systems work, as I've heard a fair bit about them in Europe, Australia etc.

Are there pumped loops to keep the water hot at point of use?

Can you explain this a little more? I'm not a building/piping engineer, to be honest.

Remembering that in Asia, ambient temp is often around 25'C at night.

[Gary]

http://shop.solardirect.com/pdf/wate...dfos-guide.pdf

[Gary]

Now, which option is better to use - bearing in mind that the COP,hp for the direct system at Tc,sat~70'C of 3.08, whereas the average COP,hp over the range Tc,sat=40'C to 70'C is close to 4.96 (theoretical)? The COP's improve further if the hot water is held at say 55'C, or 60'C instead of 65'C.

The average power output of the heat-pump over the range Tc,sat=40'C to 70'C is around 9.1% higher than that at Tc,sat~70'C.

... at what W,in temps?

I'm thinking the COP for a direct system will be much higher with cold feedwater entering the condenser.

[desA]

http://shop.solardirect.com/pdf/wate...dfos-guide.pdf

Thanks - nice article.

Now, the tank-type heater takes care of its own internal circulation - whether just natural convection - some are also pump-around, to get better heat-transfer.

The loop they seem to show is more on the water usage end, & would be independent of the storage pump-around loop, which is there to continuously pass water through the heat-source (heat-pump) & back to the tank, slowly raising its temp.

Heat-pumps are 'slow heaters' & either one has ridiculously low water flow-rates on a one pass system - then store at high temp, or use a pump-around system which gradually brings up the whole mass water temp, over time.

The water makeup - even as for the Grunfos example - just mixes inside the hot storage tank.

The direct systems seem to like to use a series of storage tanks, linked top-to-bottom, making use of the natural buoyancy of the water, to allow the hot water to rise to the top, ready for instant hot water draw-off, on demand. I've seen some this in Japanese designs & some European concepts. Apparently it is difficult to NOT mix the water contents & so these systems have limitations, it seems.

[desA]

Originally Posted by desA
Now, which option is better to use - bearing in mind that the COP,hp for the direct system at Tc,sat~70'C of 3.08, whereas the average COP,hp over the range Tc,sat=40'C to 70'C is close to 4.96 (theoretical)? The COP's improve further if the hot water is held at say 55'C, or 60'C instead of 65'C.

The average power output of the heat-pump over the range Tc,sat=40'C to 70'C is around 9.1% higher than that at Tc,sat~70'C.

Gary:
... at what W,in temps?

I'm thinking the COP for a direct system will be much higher with cold feedwater entering the condenser.

The thermodynamics for the direct system, have the Tc,sat~70'C - this is at lowest COP,hp for the system ~ 3.08.

For the indirect system, the maximum COP,hp is at start of cycle, something around 6.92. This reduces as the Tc,sat rises with the incoming feedwater temp, until eventually it hits the 3.08 mark at Tc,sat=70'C.

Now, the efficiency of heat-transfer in the condenser is a different story, for the direct, versus pump-around systems.

[desA]

The VIC(E) re-visited: (I'm presently attempting to model this)

Can we review the sequence of connections for the VIC(E) system?

T,liq,hot -> VIC -> T,liq,cool -> TXV,in -> TXV,out -> evap inlet -> Tevap,out -> VIC -> Tvap,sh -> capillary measure

In other words, evap output vapour (partially evaporated) moves to VIC, where it is completely vapourised, then superheated.

Is this the correct sequence?
How will this look on the pressure-enthalpy diagram?

[desA]

CPR re-visited

Are there any electronically-driven pressure regulating valve systems that can be used in place of a CPR?

These CPR's are substantially-priced items, but, of course are mechanical & if treated properly, will probably be expected to last way beyond an electronic lifetime.

[Gary]

The thermodynamics for the direct system, have the Tc,sat~70'C - this is at lowest COP,hp for the system ~ 3.08.

For the indirect system, the maximum COP,hp is at start of cycle, something around 6.92. This reduces as the Tc,sat rises with the incoming feedwater temp, until eventually it hits the 3.08 mark at Tc,sat=70'C.

Now, the efficiency of heat-transfer in the condenser is a different story, for the direct, versus pump-around systems.

It seems very strange that the COP would be identical for both strategies at Tc,sat=70C despite the difference in Tw,in.

And what is the cause of the increased COP at start of the cycle for the indirect system?

I'm thinking the answer must be the liquid temp at the TXV inlet... in which case the VIC will change everything.

[Gary]

The VIC(E) re-visited: (I'm presently attempting to model this)

Can we review the sequence of connections for the VIC(E) system?

T,liq,hot -> VIC -> T,liq,cool -> TXV,in -> TXV,out -> evap inlet -> Tevap,out -> VIC -> Tvap,sh -> capillary measure

In other words, evap output vapour (partially evaporated) moves to VIC, where it is completely vapourised, then superheated.

Is this the correct sequence?
How will this look on the pressure-enthalpy diagram?

I'm not sure what you mean by "capillary measure" but otherwise seems right.

[desA]

Originally Posted by desA
The VIC(E) re-visited: (I'm presently attempting to model this)

Can we review the sequence of connections for the VIC(E) system?

T,liq,hot -> VIC -> T,liq,cool -> TXV,in -> TXV,out -> evap inlet -> Tevap,out -> VIC -> Tvap,sh -> capillary measure

In other words, evap output vapour (partially evaporated) moves to VIC, where it is completely vapourised, then superheated.

Is this the correct sequence?
How will this look on the pressure-enthalpy diagram?

Gary:
I'm not sure what you mean by "capillary measure" but otherwise seems right.

Ok, thanks.

I meant measurement by TXV capillary bulb - tried to shorten the wording. Apologies.

Now, are there examples of a VIC-type system in applications?

Why I ask, is that, I'm currently running the VIC setup on a simulator I use to model all kinds of stuff, including the refrigerant & heat-pump circuits. No matter how hard I try, the thing will not converge cleanly. The last time I had this was in simulating CO2 circuits. When I eventually researched a little deeper, I found out that CO2 circuits exhibit 'bi-stability' or even 'multi-stability'. In other words, there are different stable solutions & the solver jumps between these - sometimes settling on one, or the other.

Practically, this would seem to be taking place due to the change in solution modes as follows:
1. Evap fully evaporates refrigerant, adds small amount of superheat. VIC then adds further superheat.
2. Evap fully evaporates refrigerant, adds no superheat. VIC then adds superheat.
3. Evap partially evaporates refrigerant - exits wet. VIC then continues evaporation & adds superheat (ie. is an evaporator).

Using the sequence as outlined above, one single stable solution does not seem to exist, on a simulator, at least.

This could actually mean, in practice, that an evap-VIC combination could possibly dance around between stable solutions, if not carefully managed - in theory.

This is why I would really like to know how these may have shaped up in the field.

[Gary]

CPR re-visited

Are there any electronically-driven pressure regulating valve systems that can be used in place of a CPR?

These CPR's are substantially-priced items, but, of course are mechanical & if treated properly, will probably be expected to last way beyond an electronic lifetime.

CPR's last pretty much forever and almost never lose their adjustment. Their downside is that they have an adjustment screw for people to play with.

[desA]

It seems very strange that the COP would be identical for both strategies at Tc,sat=70C despite the difference in Tw,in.

And what is the cause of the increased COP at start of the cycle for the indirect system?

I'm thinking the answer must be the liquid temp at the TXV inlet... in which case the VIC will change everything.

COP is determined thermodynamically from Te,sat & Tc,sat temps.

In our case, Te,sat is relatively fixed - will be more so when using the CPR.

When Tc,sat is low (at start of cycle), then COP is high, since motor power is low & Q'cond high, at start. When Tc,sat is high (at end of cycle), then COP is low, since motor power is high, & Q'cond is lower at end.

For the direct flow through the condenser, we push Te,sat to the highest Tc,sat~70'C, with Te,sat~12.5'C. Thermodynamically, this is the same as that for the hot (end) pump-around case. For thermodynamcs, the condenser (waterside) doesn't feature at all in the COP calcs.

[Gary]

Ok, thanks.

I meant measurement by TXV capillary bulb - tried to shorten the wording. Apologies.

Now, are there examples of a VIC-type system in applications?

Why I ask, is that, I'm currently running the VIC setup on a simulator I use to model all kinds of stuff, including the refrigerant & heat-pump circuits. No matter how hard I try, the thing will not converge cleanly. The last time I had this was in simulating CO2 circuits. When I eventually researched a little deeper, I found out that CO2 circuits exhibit 'bi-stability' or even 'multi-stability'. In other words, there are different stable solutions & the solver jumps between these - sometimes settling on one, or the other.

Practically, this would seem to be taking place due to the change in solution modes as follows:
1. Evap fully evaporates refrigerant, adds small amount of superheat. VIC then adds further superheat.
2. Evap fully evaporates refrigerant, adds no superheat. VIC then adds superheat.
3. Evap partially evaporates refrigerant - exits wet. VIC then continues evaporation & adds superheat (ie. is an evaporator).

Using the sequence as outlined above, one single stable solution does not seem to exist, on a simulator, at least.

This could actually mean, in practice, that an evap-VIC combination could possibly dance around between stable solutions, if not carefully managed - in theory.

This is why I would really like to know how these may have shaped up in the field.

#3 is correct. #1 and #2 are incorrect.

As far as I know, there are none in the field... although this seems too obvious not to have been tried in the past.

[desA]

CPR's last pretty much forever and almost never lose their adjustment.

I think that they are a really nice concept. Very robust.

Their downside is that they have an adjustment screw for people to play with.

How do system manufacturers guard against this? Does it help to dab a spot of paint on the final as-built screw setting?

[desA]

#3 is correct. #1 and #2 are incorrect.

As far as I know, there are none in the field... although this seems too obvious not to have been tried in the past.

Never fear, I'll rig up something on the lab machine, once we have got all the charge settings sorted out. We can then see what on earth happens in reality, not in simulator.

I'll carry on with a hand calculation based on #3. The evaporator can be 'downsized' by cutting back the fan, or blocking off part of the face with insulation.

[Gary]

Never fear, I'll rig up something on the lab machine, once we have got all the charge settings sorted out. We can then see what on earth happens in reality, not in simulator.

I'll carry on with a hand calculation based on #3. The evaporator can be 'downsized' by cutting back the fan, or blocking off part of the face with insulation.

As to simulation, the VIC does two things: It adds a suction/liquid heat exchanger and it (ideally) adds evaporator capacity sufficient to cool the liquid from Tc,out to Te,sat.

Another way to look at this: A conventional suction/liquid heat exchanger calculation balances the gain from cooling the liquid against the loss from increased superheat. The VIC eliminates that loss from the equation as the superheat remains the same.

[Gary]

The evaporator can be 'downsized' by cutting back the fan, or blocking off part of the face with insulation.

I would very much prefer the cutting back the fan option... even if it means intermittantly cutting off the fan.

[Gary]

How do system manufacturers guard against this? Does it help to dab a spot of paint on the final as-built screw setting?

I would think that a dab of paint coupled with a "do not adjust or warranty is void" label should deter most, but nothing can stop a determined idiot.

[micknick]

We are testing our chillers at 55C (131F) ambient and experienced compressor overload at various points.
Our solution was a suction line heat exchanger with a txv feed the opposite direction of the suction flow. Insulate the heck of the suction line. My suction temperature 6-9C with a below normal amp draw.
5 ton unit 1 ton TXV.

[Gary]

We are testing our chillers at 55C (131F) ambient and experienced compressor overload at various points.
Our solution was a suction line heat exchanger with a txv feed the opposite direction of the suction flow. Insulate the heck of the suction line. My suction temperature 6-9C with a below normal amp draw.
5 ton unit 1 ton TXV.

I'm thinking the same effect could be had without the heat exchanger by simply injecting the output of the TXV directly into the suction line, with the bulb mounted downstream, i.e. liquid injection.

But once again we see that its about keeping the compressor cool.

[mad fridgie]

Hi desA, if your heat pump is to be manufactured, if so ensure that you keep within the compressors manufactures working envelope. protection against warranty issues.
AWHPs will always be in state flux, to cure your varying refrigeration load, temperatures and pressures you are best to use a double ported TXV, giving your a more constant control over your evap coil.
As far as your heat exchanger, you need to focus on thermal length, not just surface area.

[mad fridgie]

and I should say the TXV should be rated as MOP at your maximum allowable SST

[desA]

As to simulation, the VIC does two things: It adds a suction/liquid heat exchanger and it (ideally) adds evaporator capacity sufficient to cool the liquid from Tc,out to Te,sat.

Another way to look at this: A conventional suction/liquid heat exchanger calculation balances the gain from cooling the liquid against the loss from increased superheat. The VIC eliminates that loss from the equation as the superheat remains the same.

Now, we can look at the following concept, which his stunningly stable under simulation. I think I did see something similar in the Alfa-Laval book:

Place VIC before evaporator...

T,liq,hot -> VIC -> T,liq,cool -> TXV,in -> TXV,out -> VIC -> evap,in -> evap,out -> capillary bulb

So, basically the liquid line wraps in/out of the VIC, & makes use of the return 2-phase fluid from the TXV to interchange heat.

This configuration suffers from no dynamic (simulation) instability at all, & settles in one iteration, whereas the previous configuration cannot settle in 200 iterations.

[desA]

I would very much prefer the cutting back the fan option... even if it means intermittantly cutting off the fan.

No problem on the lab machine, I have a very nice dimmer switch installed. It's been smooth even down to pretty low fan speeds, with only the occasional hum.

[desA]

I would think that a dab of paint coupled with a "do not adjust or warranty is void" label should deter most, but nothing can stop a determined idiot.

Good point.

In Asia the electronics & homeware goods suppliers place a small, very sticky, thin label on equipment - it is dated. If you break this seal, you get no deal...

I'll use some of these in the systems.

[desA]

We are testing our chillers at 55C (131F) ambient and experienced compressor overload at various points.

Thanks micknick, for your very useful comment. 55'C is quite some temperature - I'd imagine this is the local ambient surroundings where the compressor is operating. What were the symptoms of the compressor overload?

Our solution was a suction line heat exchanger with a txv feed the opposite direction of the suction flow. Insulate the heck of the suction line. My suction temperature 6-9C with a below normal amp draw.
5 ton unit 1 ton TXV.

So, basically a suction line cooler, using 2-phase feed out of the TXV.

How did your evaporator respond to the rise in vapour fraction?

[desA]

Originally Posted by micknick
We are testing our chillers at 55C (131F) ambient and experienced compressor overload at various points.
Our solution was a suction line heat exchanger with a txv feed the opposite direction of the suction flow. Insulate the heck of the suction line. My suction temperature 6-9C with a below normal amp draw.
5 ton unit 1 ton TXV.

Gary:
I'm thinking the same effect could be had without the heat exchanger by simply injecting the output of the TXV directly into the suction line, with the bulb mounted downstream, i.e. liquid injection.

But once again we see that its about keeping the compressor cool.

Can you elaborate a little more on the liquid injection approach a little more?

I see that some compressors, for instance, have a liquid injection option.

[desA]

Hi desA, if your heat pump is to be manufactured, if so ensure that you keep within the compressors manufactures working envelope. protection against warranty issues.

Agreed. So far, the agreed upper limit for this window, on the compressor in the lab machine is, Te,sat=13'C, Tc,sat=65.6'C.

I say 'agreed' because the US & European technology centres for this compressor have different published opinions. The Europeans are happy at Te,sat=15'C, Tc,sat=75'C, according to their published compressor envelope. Odd.

AWHPs will always be in state flux, to cure your varying refrigeration load, temperatures and pressures you are best to use a double ported TXV, giving your a more constant control over your evap coil.

Can you explain this in a little more detail?

As far as your heat exchanger, you need to focus on thermal length, not just surface area.

Interesting point. Can you explain the terms 'thermal length' versus 'surface area' a little more?

[desA]

and I should say the TXV should be rated as MOP at your maximum allowable SST

Thanks for that.

I am not overly happy with using the MOP option, for a number of reasons. It is trying to serve two functions - controlling evap pressure & temperature. I think that it could really only do one of these effectively across the wide range of operational window called on for an AWHP.

Also, the MOP option will give maintenance folks a real headache. I feel that this is sub-optimal design.

They also add cost, for little function, in my view.

[Gary]

Now, we can look at the following concept, which his stunningly stable under simulation. I think I did see something similar in the Alfa-Laval book:

Place VIC before evaporator...

T,liq,hot -> VIC -> T,liq,cool -> TXV,in -> TXV,out -> VIC -> evap,in -> evap,out -> capillary bulb

So, basically the liquid line wraps in/out of the VIC, & makes use of the return 2-phase fluid from the TXV to interchange heat.

This configuration suffers from no dynamic (simulation) instability at all, & settles in one iteration, whereas the previous configuration cannot settle in 200 iterations.

Hmmm... the transfer would cool the liquid line, but that heat is transferred right back into the refrigerant on the other side of the TXV, to be carried through the evap. I don't see where there would be any gain there. Perhaps some by eliminating flashing.

Placing the VIC at the evap outlet transfers the heat from liquid to suction without carrying that heat through the evaporator.

[Gary]

Agreed. So far, the agreed upper limit for this window, on the compressor in the lab machine is, Te,sat=13'C, Tc,sat=65.6'C.

I say 'agreed' because the US & European technology centres for this compressor have different published opinions. The Europeans are happy at Te,sat=15'C, Tc,sat=75'C, according to their published compressor envelope. Odd.

I wonder if the difference might be due to 50hz electricity in Europe versus 60hz electricity in US.

[Gary]

Can you elaborate a little more on the liquid injection approach a little more?

I see that some compressors, for instance, have a liquid injection option.

On some systems, the superheat at the compressor inlet can be considerably higher than at the TXV bulb, causing the compressor to run hot. Liquid is injected into the suction line, or sometimes directly into the compressor, to reduce the superheat and cool the compressor.

[desA]

Hmmm... the transfer would cool the liquid line, but that heat is transferred right back into the refrigerant on the other side of the TXV, to be carried through the evap. I don't see where there would be any gain there. Perhaps some by eliminating flashing.

Placing the VIC at the evap outlet transfers the heat from liquid to suction without carrying that heat through the evaporator.

I'm currently doing a study of this arrangement at the moment. The VIC will actually take some of the heating load off the evaporator, if sized correctly - there is a direct benefit in evaporator size reduction, it seems.

Basically, lose evap size, but add VIC - the economics will depend on just how hard you want VC to work.

HE sizing, under this arrangement, will be critical to achieve success.

[desA]

Originally Posted by desA
Agreed. So far, the agreed upper limit for this window, on the compressor in the lab machine is, Te,sat=13'C, Tc,sat=65.6'C.

I say 'agreed' because the US & European technology centres for this compressor have different published opinions. The Europeans are happy at Te,sat=15'C, Tc,sat=75'C, according to their published compressor envelope. Odd.

Gary:
I wonder if the difference might be due to 50hz electricity in Europe versus 60hz electricity in US.

Mmhh... an interesting point. That's got me thinking.

How would the frequency affect operation of the compressor? Would this be in terms of motor heat? At 60Hz, does the compressor rpm change, compared to 50Hz?

Very interesting...