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TURBOEXPANDER HP CALCULATION 1
- Thread starter tomato5
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Hey Tomato5,
For expanders an isentropic efficiency is used. This ratio is the ratio of actual enthalpy change for the expansion to the isentropic enthalpy change, which is equal to actual work over isentropic work.
The starting enthalpy if a function of P1 and T1. The isentropic enthalpy is calculated at P2 via the isentropic path from thermodynamic tables or equation of state. The actual enthalpy is based on the actual outlet temp (T2) and vapor fraction at P2 if there is condensation. The actual outlet conditions are needed to calculate the efficiency (and actual process work).
Once the isentropic efficiency is known, the power can be reasonably calculated at new conditions such as a different outlet pressure. It is the relatively constant efficiency for expansion over a wide range which makes isentropic efficiency useful for expanders and steam turbines, where as polytropic efficiency is used for compression for the same reason.
Best wishes,
Sshep
For expanders an isentropic efficiency is used. This ratio is the ratio of actual enthalpy change for the expansion to the isentropic enthalpy change, which is equal to actual work over isentropic work.
The starting enthalpy if a function of P1 and T1. The isentropic enthalpy is calculated at P2 via the isentropic path from thermodynamic tables or equation of state. The actual enthalpy is based on the actual outlet temp (T2) and vapor fraction at P2 if there is condensation. The actual outlet conditions are needed to calculate the efficiency (and actual process work).
Once the isentropic efficiency is known, the power can be reasonably calculated at new conditions such as a different outlet pressure. It is the relatively constant efficiency for expansion over a wide range which makes isentropic efficiency useful for expanders and steam turbines, where as polytropic efficiency is used for compression for the same reason.
Best wishes,
Sshep
- Thread starter
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salmanazim
Chemical
hey tomato & sshep
our plant capacity is 125mmscfd and we hv turbo-expeder but if v run plant at full capacity (125mmscfd) then v r facing high expender thrust problem.if any one knows its remidal action thn plzzz..................
our plant capacity is 125mmscfd and we hv turbo-expeder but if v run plant at full capacity (125mmscfd) then v r facing high expender thrust problem.if any one knows its remidal action thn plzzz..................
You need to talk to your turbine supplier's technical service rep
Most of the time the thrust problems are related to the monitoring device and not true thrust problems.
What is happening is as the gas flow increases, the pressure also increases causing the increased thrust forces.
The rotating assembly then moves towards the thrust bearings. If this "axial displacement" is too great, the thrust bearings can be quickly worn or damaged.
Do you have a rotating machinery expert you can utilize?
Most of the time the thrust problems are related to the monitoring device and not true thrust problems.
What is happening is as the gas flow increases, the pressure also increases causing the increased thrust forces.
The rotating assembly then moves towards the thrust bearings. If this "axial displacement" is too great, the thrust bearings can be quickly worn or damaged.
Do you have a rotating machinery expert you can utilize?
Hey Tomato5,
Sorry to take so long to get back- i have been having a great holiday with little "work".
I don't have much experience with liquid power recovery, but theoretically the max that could be recovered is calculated as hydraulic horsepower same as a pump. Outlet temperature is not a good alternative to direct power measurement for efficiency calculation (such is done with compressible fluids) because high liquid heat capacity means the dT is small. In all cases (liquid/vapor/mixed) a power balance using data on the other side of the shaft is the ultimate efficiency check.
I have seen that cryogenic turbo-expanders can achieve efficiencies of up to 85%, but the key (in my experience) is to operate at design speed. If the other side is over or under loaded, then the efficiency will be poor. I have seen this a few times in expander-compressor systems. If you try and demand too much compression work then you cut your own throat- The symptoms will be low speed, low efficiency, and consequently low dP on the compressor side. The control system needs to facilitate the optimum power balance and get the expander to design conditions (speed, inlet P, dP) at design gas composition. The result of a proper power balance will be low temp outlet, maximum liquid condensation, and maximum work recovered.
Is this a design query or operation of existing equipment? Is the goal power recovery, to maximize cryogenic refrigeration, or just general theory, etc?
Best wishes,
Sshep
Sorry to take so long to get back- i have been having a great holiday with little "work".
I don't have much experience with liquid power recovery, but theoretically the max that could be recovered is calculated as hydraulic horsepower same as a pump. Outlet temperature is not a good alternative to direct power measurement for efficiency calculation (such is done with compressible fluids) because high liquid heat capacity means the dT is small. In all cases (liquid/vapor/mixed) a power balance using data on the other side of the shaft is the ultimate efficiency check.
I have seen that cryogenic turbo-expanders can achieve efficiencies of up to 85%, but the key (in my experience) is to operate at design speed. If the other side is over or under loaded, then the efficiency will be poor. I have seen this a few times in expander-compressor systems. If you try and demand too much compression work then you cut your own throat- The symptoms will be low speed, low efficiency, and consequently low dP on the compressor side. The control system needs to facilitate the optimum power balance and get the expander to design conditions (speed, inlet P, dP) at design gas composition. The result of a proper power balance will be low temp outlet, maximum liquid condensation, and maximum work recovered.
Is this a design query or operation of existing equipment? Is the goal power recovery, to maximize cryogenic refrigeration, or just general theory, etc?
Best wishes,
Sshep
Depending on the application, using the discharge temperature to calculate the expander efficiency can be a poor method. For example, if the process stream at the discharge is two phase (which it usually is) then sometimes depending on the composition/temperature/pressure there can be a small change in temperature for a large change in efficiency. Also, if the process flow is relatively small and very low temperature, then you can have significant seal gas heating effects which will raise the discharge temperature.
The best way to calculate the expander efficiency is to first calculate the compressor horsepower. You can do this using the flow, inlet/discharge pressures, and temperatures. If the surge valve is open then you will need to account for the additional flow and the true inlet (mixed) temperature. I usually calculate the head first, then check versus the original compressor curve for the operating inlet volumetric flow rate. If it aligns with the curve for the operating speed, then I use the curve efficiency (checking versus the calculated operating efficiency to see if it agrees) to calculate power.
Then add mechanical losses (which can be estimated based on bearing size and speed) and you have the expander power. You can then use the expander flow and calculated isentropic head (from the inlet temp, composition, inlet and discharge pressures) to calculate the expander efficiency. I then compare the expected expander discharge temperature to the actual operating for agreement.
The other big unknown in a gas plant (I'm assuming that is what you are looking at, since you list flows in MMSCFD) is the expander flow rate. The flow meters in the cold separator line normally are not meant to be very accurate. If you have a reflux line which takes a portion of the cold sep as feed, then determining an accurate expander flow rate can be a problem. I find that in most cases you can determine the expander efficiency within 3-5 points, but usually not more accurate than that.
The best way to calculate the expander efficiency is to first calculate the compressor horsepower. You can do this using the flow, inlet/discharge pressures, and temperatures. If the surge valve is open then you will need to account for the additional flow and the true inlet (mixed) temperature. I usually calculate the head first, then check versus the original compressor curve for the operating inlet volumetric flow rate. If it aligns with the curve for the operating speed, then I use the curve efficiency (checking versus the calculated operating efficiency to see if it agrees) to calculate power.
Then add mechanical losses (which can be estimated based on bearing size and speed) and you have the expander power. You can then use the expander flow and calculated isentropic head (from the inlet temp, composition, inlet and discharge pressures) to calculate the expander efficiency. I then compare the expected expander discharge temperature to the actual operating for agreement.
The other big unknown in a gas plant (I'm assuming that is what you are looking at, since you list flows in MMSCFD) is the expander flow rate. The flow meters in the cold separator line normally are not meant to be very accurate. If you have a reflux line which takes a portion of the cold sep as feed, then determining an accurate expander flow rate can be a problem. I find that in most cases you can determine the expander efficiency within 3-5 points, but usually not more accurate than that.
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