Sizing hydraulic lines, SUS and velocity

[grandnat_6]

Hi,

I'm trying to size up a suction line, for a positive displacement gear pump. I'm flowing 3.4 GPM. The oil I'm looking at using is Case IH Hy-tran. It has a SUS # of 208 at 100 degrees F and 47 at 210 degrees F. cSt of 40 at 40 degrees C and 6.3 at 100 degrees C.

I have sized up a 7/8" I.D. suction line that gives me a velocity of about 1.8 ft/sec

Reynolds number shows I have laminar flow.

If the temperature drops to 0 degrees F or lower. I know it will take more horse power to move the oil. I also know about the tube I.D. ruffness. I'm worried about the pump cavitating on cold days until the fluid warms up.

Will the lower temp cause my Reynolds's number to change?
Will it change the velocity?
If so, I would think a larger I.D. suction line is in order...

Is there a way to calculate the known SUS # or cSt # for the temp change? Or do I have to contact the oil manufacture and find this out? I have a feeling if the SUS and cSt is graphed by temp; it would be some kind of curve and not a straight line?

I'm hoping someone can straighten me out on this and get me in the right direction.

Thanks!

 

[SteamKing]

Hi,

I'm trying to size up a suction line, for a positive displacement gear pump. I'm flowing 3.4 GPM. The oil I'm looking at using is Case IH Hy-tran. It has a SUS # of 208 at 100 degrees F and 47 at 210 degrees F. cSt of 40 at 40 degrees C and 6.3 at 100 degrees C.

I have sized up a 7/8" I.D. suction line that gives me a velocity of about 1.8 ft/sec

Reynolds number shows I have laminar flow.

If the temperature drops to 0 degrees F or lower. I know it will take more horse power to move the oil. I also know about the tube I.D. ruffness. I'm worried about the pump cavitating on cold days until the fluid warms up.

You need to determine the Net Positive Suction Head Required for your pump (NPSHR) to make sure the suction lines are properly designed and your pump won't cavitate.

http://www.pumpschool.com/applications/NPSH.pdf

Will the lower temp cause my Reynolds's number to change?
Will it change the velocity?
If so, I would think a larger I.D. suction line is in order...

For circular piping, the Reynolds No. is calculated as Re = v * D / nu, where v is the velocity of the fluid in the pipe, D is the internal diameter, and nu is the kinematic viscosity of the fluid. Generally, the stoke or the centistoke is used as the unit of kinematic viscosity for calculating Reynolds No. instead of SUS. SUS can be converted to stokes. You must be careful to use consistent units when calculating Re, since that quantity is a dimensionless number.

For a given flow, increasing the size of the line also results in a lower velocity, because the flow Q = A * v, where A is the internal cross section area of the flow and v is the flow velocity. For the same Q, if A gets bigger, v gets slower.

For a given line size and flow velocity, the Re will increase as this oil is heated, because the viscosity (nu) gets lower with increased temperature. If you can heat the oil to reduce the viscosity without degrading its other physical characteristics, then it might be possible for the oil to flow in a turbulent rather than laminar flow regime in the pipe, which means that pipe friction assumes a certain fixed value. (See the Moody Chart)

https://en.wikipedia.org/wiki/Moody_chart

Is there a way to calculate the known SUS # or cSt # for the temp change? Or do I have to contact the oil manufacture and find this out? I have a feeling if the SUS and cSt is graphed by temp; it would be some kind of curve and not a straight line?

It would be better to get this information directly from the oil's manufacturer. It is standard engineering design information. It also appears that the name Case IH Hy-Tran is used for a number of different types of fluid, so be sure to specify what you are using the fluid for.

I'm hoping someone can straighten me out on this and get me in the right direction.

Thanks!

 

[grandnat_6]

SteamKing,

Thank you for the information and links, my fluid power book doesn't talk about NPSHa. I'll search and study more on it.

 

[SteamKing]

SteamKing,

Thank you for the information and links, my fluid power book doesn't talk about NPSHa. I'll search and study more on it.

The NPSHa figure is different for each piping system. For a range of fluid flows, you calculate the pressure drop which is produced by all the fittings, valves, and piping in the system. As long as NPSHa > NPSHr, your pump should not cavitate.

Most of the methods for calculating pressure drops are valid only for fluid flow which is fully turbulent. If you have a system where laminar flow may occur, then you must use different methods to calculate the pressure drop in that situation, in order to obtain accurate results.

The Crane Technical Paper 410 or the Hydraulic Institute Handbook contain charts and examples of how to calculate pressure drop for a piping system with fully turbulent flow.

 

[grandnat_6]

I did a lot of studying this weekend on NPSHa and NPSHr. I have a few more questions.

A publication I found talked about gauge pressure for surface pressure head. Is this the pressure or vacuum on the surface of the fluid on the tank? Or is it an actual suction pressure reading from the running pump?

Vapor pressure and specific gravity change with temperature? If so, then I have to check NPSHa for each temp extreme?

I think this will clear up NPSHa for me. I didn't know it was only for turbulent flow. I'll start to study the laminar flow end.

Thank you again; SteamKing.

 

[SteamKing]

I did a lot of studying this weekend on NPSHa and NPSHr. I have a few more questions.

A publication I found talked about gauge pressure for surface pressure head. Is this the pressure or vacuum on the surface of the fluid on the tank? Or is it an actual suction pressure reading from the running pump?

If you could provide a copy of the text from this publication, it might make it clear what is being discussed. The actual pressure or vacuum inside the oil storage tank is part of the determination of NPSHa.

Vapor pressure and specific gravity change with temperature? If so, then I have to check NPSHa for each temp extreme?

Yes. All of these properties are sensitive to the temperature of the oil.

Since the properties of your oil vary with temperature, then you'll have to investigate what happens when the oil has its maximum viscosity, which should give the maximum pressure drop. If there is a way to pre-heat the oil before pumping, this may eliminate temperature from the list of the variables in your study of NPSHa.

I think this will clear up NPSHa for me. I didn't know it was only for turbulent flow. I'll start to study the laminar flow end.

The type of flow, laminar, turbulent, or something in between, determines the amount of pressure drop in the suction lines for different flow rates.

 

[grandnat_6]

If you could provide a copy of the text from this publication, it might make it clear what is being discussed. The actual pressure or vacuum inside the oil storage tank is part of the determination of NPSHa.

http://www.mcnallyinstitute.com/11-html/11-12.html

• Static head. Measure it from the centerline of the pump suction to the top of the liquid level. If the level is below the centerline of the pump it will be a negative or minus number.
• Surface pressure head. Convert the gage absolute pressure to feet of liquid using the formula:
  • Pressure = head x specific gravity / 2.31
• Vapor pressure of your product . Look at the vapor pressure http://www.mcnallyinstitute.com/Charts/Vapor_Pressure_Chart.html in the "charts you can use" section in the home page of this web site. You will have to convert the pressure to head. If you use the absolute pressure shown on the left side of the chart, you can use the above formula
• Specific gravity of your product. You can measure it with a hydrometer if no one in your facility has the correct chart or knows the number.
• Loss of pressure in the piping, fittings and valves. Use the three charts in the "charts you can use" section in the home page of this web site
  • Find the chart for the proper pipe size, go down to the gpm and read across to the loss through one hundred feet of pipe directly from the last column in the chart. As an example: http://www.mcnallyinstitute.com/Charts/friction_2_2.5.html, 65 gpm = 7.69 feet of loss for each 100 feet of pipe.
  • For valves and fittings look up the resistance coefficient numbers (K numbers) for all the valves and fittings, add them together and multiply the total by the V2/2g number shown in the fourth column of the friction loss piping chart. Example: A http://www.mcnallyinstitute.com/Charts/Friction_fittings.html has a K number of 0.4 and a 2 inch http://www.mcnallyinstitute.com/Charts/Friction_valves.html has a K number of 8. Adding them together (8 + 0.4) = 8.4 x 0.6 (for 65 gpm) = 5 feet of loss.

 

[SteamKing]

http://www.mcnallyinstitute.com/11-html/11-12.html

• Static head. Measure it from the centerline of the pump suction to the top of the liquid level. If the level is below the centerline of the pump it will be a negative or minus number.
• Surface pressure head. Convert the gage absolute pressure to feet of liquid using the formula:
  • Pressure = head x specific gravity / 2.31
• Vapor pressure of your product . Look at the vapor pressure http://www.mcnallyinstitute.com/Charts/Vapor_Pressure_Chart.html in the "charts you can use" section in the home page of this web site. You will have to convert the pressure to head. If you use the absolute pressure shown on the left side of the chart, you can use the above formula
• Specific gravity of your product. You can measure it with a hydrometer if no one in your facility has the correct chart or knows the number.
• Loss of pressure in the piping, fittings and valves. Use the three charts in the "charts you can use" section in the home page of this web site
  • Find the chart for the proper pipe size, go down to the gpm and read across to the loss through one hundred feet of pipe directly from the last column in the chart. As an example: http://www.mcnallyinstitute.com/Charts/friction_2_2.5.html, 65 gpm = 7.69 feet of loss for each 100 feet of pipe.
  • For valves and fittings look up the resistance coefficient numbers (K numbers) for all the valves and fittings, add them together and multiply the total by the V2/2g number shown in the fourth column of the friction loss piping chart. Example: A http://www.mcnallyinstitute.com/Charts/Friction_fittings.html has a K number of 0.4 and a 2 inch http://www.mcnallyinstitute.com/Charts/Friction_valves.html has a K number of 8. Adding them together (8 + 0.4) = 8.4 x 0.6 (for 65 gpm) = 5 feet of loss.

This article is just showing you how to measure the static pressures in the system and convert them to units of feet of head for the liquid being pumped.

The K-factors which are being discussed here are for calculating the pressure drop of the piping system. A word of caution: these factors should be used only for fully turbulent flow in the piping system. When the flow is laminar or in transition from laminar to turbulent, these K-factors will change in value, and the turbulent values may not give accurate pressure drop results for anything but turbulent flow.

 

[grandnat_6]

It is for measuring the static pressure and turning it into feet. It's something I wanted to know how to do to make sure I have a large enough suction line and to make sure a suction filter will work. I did not know the number published was for turbulent flow.

I found out by using this formula that a suction filter will not work based on NPSHr of the pump and I need to put it on the return line. I don't think there is a calculation to figure the PSI in the return line to the tank knowing the GPM, velocity and size of pipe/tube. The series of filters I'm looking at has a 200 psi max working pressure and I can get up to a 25 psi max by pass valve in the filter head.

I want to know what the system pressure is on the return line to ensure not to hit the 200 psi max and also that the by-pass valve isn't open all the time. I know that my return line pressure will vary depending on if the system is running thru the open center or if the directional control valve has been switched as the fluid being dumped from the opposite side of the cylinder should cause higher pressures. I'm sure shifting the control valve the other way with the force on the cylinder will be higher yet as the load on the cylinder will cause the fluid to exit at a higher velocity which will increase pressure thru the return line possibly opening the by pass valve.

Any websites or anything I should look at to find my answer? I couldn't find any by searching.

 

[SteamKing]

It is for measuring the static pressure and turning it into feet. It's something I wanted to know how to do to make sure I have a large enough suction line and to make sure a suction filter will work. I did not know the number published was for turbulent flow.

I found out by using this formula that a suction filter will not work based on NPSHr of the pump and I need to put it on the return line. I don't think there is a calculation to figure the PSI in the return line to the tank knowing the GPM, velocity and size of pipe/tube. The series of filters I'm looking at has a 200 psi max working pressure and I can get up to a 25 psi max by pass valve in the filter head.

IDK about putting a suction filter in the return line. The filter is placed in the suction line to keep trash/water out of the pump and the rest of the hydraulic system. Switching it to the return line is a waste of time, IMO. If the pump can't handle the PD thru the filter, get a different filter, redesign the suction piping, or select a different pump.

I want to know what the system pressure is on the return line to ensure not to hit the 200 psi max and also that the by-pass valve isn't open all the time. I know that my return line pressure will vary depending on if the system is running thru the open center or if the directional control valve has been switched as the fluid being dumped from the opposite side of the cylinder should cause higher pressures. I'm sure shifting the control valve the other way with the force on the cylinder will be higher yet as the load on the cylinder will cause the fluid to exit at a higher velocity which will increase pressure thru the return line possibly opening the by pass valve.

Any websites or anything I should look at to find my answer? I couldn't find any by searching.

You're getting into specific areas which I'm not qualified to handle. You should find a knowledgeable person who can look first hand at what you want to do with your system and advise you.