13. LOBE PUMP
1. As the lobes come out of mesh, they create expanding volume on the inlet side of the
pump. Liquid flows into the cavity and is trapped by the lobes as they rotate.
2. Liquid travels around the interior of the casing in the pockets between the lobes and
the casing -- it does not pass between the lobes.
3. Finally, the meshing of the lobes forces liquid through the outlet port under pressure.
24. Total Static Head
The total static head is the total vertical distance the pump must lift the water. When pumping from a
well, it would be the distance from the pumping water level in the well to the ground surface plus the
vertical distance the water is lifted from the ground surface to the discharge point. When pumping
from an open water surface it would be the total vertical distance from the water surface to the
discharge point.
Pressure Head
Sprinkler and drip irrigation systems require pressure to operate. Center pivot systems require a
certain pressure at the pivot point to distribute the water properly. The pressure head at any point
where a pressure gage is located can be converted from pounds per square inch (PSI) to feet of
head by multiplying by 2.31. For example, 20 PSI is equal to 20 times 2.31 or 46.2 feet of head.
Friction Head
Friction head is the energy loss or pressure decrease due to friction when water flows through
pipe networks. The velocity of the water has a significant effect on friction loss. Loss of head due
to friction occurs when water flows through straight pipe sections, fittings, valves, around
corners, and where pipes increase or decrease in size. Values for these losses can be
calculated or obtained from friction loss tables. The friction head for a piping system is the sum
of all the friction losses
Velocity Head
Velocity head is the energy of the water due to its velocity. This is a very small amount of energy
and is usually negligible when computing losses in an irrigation system
25. The Total Dynamic Head (TDH) is the sum of the total static head, the total
friction head and the pressure head. The components of the total static head for
a surface water and well wate pumping system are shown
TOTAL DYNAMIC HEAD
26. PB = Barometric pressure in feet absolute.
VP = Vapor pressure of the liquid at maximum pumping temperature, in feet absolute.
P = Pressure on surface of liquid in closed suction tank, in feet absolute.
Ls = Maximum static suction lift in feet.
LH = Minimum static suction head in feet.
hf = Friction loss in feet in suction pipe at required capacity
NET POSITIVE
SUCTION HEAD
27.
28. NPSHA = Atmospheric pressure(converted to head) + static head + surface pressure head -
vapor pressure of your product - loss in the piping, valves and fittings
29. The NPSH available in a
flooded suction system is:
Atmospheric Pressure (- )
Vapor Pressure (+) Liquid
Height (-) Friction in the
Suction Line
The NPSH available in a suction
lift system is:
Atmospheric Pressure (-) Vapor
Pressure (-) Liquid Ht. (-) Friction
in the Suction Line.
Net Positive Suction Head Available (NPSHA)
The net positive suction head available is a function of the pump suction system.
The Net Positive Suction Head is the absolute total suction head in feet.
30.
31. Where
N = Pump speed RPM
Q = GPM = Pump flow at best efficiency point at impeller inlet
(for double suction impellers divide total pump flow by two).
hsv = NPSHR = Pump NPSH required at best efficiency point.
Suction specific speed (S or NS) is defined as:
32.
33.
34.
35. BHP = Flow(GPM) x TDH(FT) x SG /3960 x EFFICIENCY(%)
Example: BHP = (100 GPM) x (95 Ft) x (1.0) / 3960 x 0.6
BHP = 4.0
Horsepower at the output shaft of
an engine, turbine, or motor is
termed brake horsepower or shaft
horsepower, depending on what
kind of instrument is used to
measure it. Horsepower of
reciprocating engines, particularly
in the larger sizes, is often
expressed as indicated
horsepower, which is determined
from the pressure in the cylinders.
Brake or shaft horsepower is less
than indicated...
36.
37. PUMP CHARACTERISTICS
Kurva sistem
Dari diameter impeller dan kecepatan yang ditentukan, pompa sentrifugal mempunyai
“performance curve” yang tertentu dan dapat diprediksi. Titik dimana pompa
dioperasikan dalam kurvanya tergantung pada sistem dimana pipa tersebut dioperasikan,
biasanya dinamakan “System Head Curve” – atau, hubungan antara aliran dengan
:hydraulic losses” dalam sistem. Representasi hal ini ada dalam bentuk grafik
dan, karena friction losses bervariasi menurut kuadrat laju alir, kurva sistem berbentuk
parabola.
Dengan memplotkan system
head cuve dan pump curve
sekaligus, dapat diketahui :
(1)Dimana pompa akan
dioperasikan dalam kurva
(2)Perubahan apa yang terjadi
jika system head curve
atau pump performance
curve berubah
38. NO STATIC HEAD - ALL FRICTION
As the levels in the suction and discharge are the same (Fig. 1), there is no static
head and, therefore, the system curve starts at zero flow and zero head and its
shape is determined solely from pipeline losses. The point of operation is at the
intersection of the system head curve and the pump curve. The flow rate may be
reduced by throttling valve.
39. POSITIVE STATIC HEAD
The parabolic shape of the
system curve is again
determined by the friction
losses through the system
including all bends and valves.
But in this case there is a
positive static head involved.
This static head does not affect
the shape of the system curve
or its "steepness", but it does
dictate the head of the system
curve at zero flow rate.
The operating point is at the
intersection of the system curve
and pump curve. Again, the
flow rate can be reduced by
throttling the discharge valve.
40. NEGATIVE (GRAVITY) HEAD
In the illustration below, a
certain flow rate will occur by
gravity head alone. But to
obtain higher flows, a pump Is
required to overcome the pipe
friction losses in excess of "H" -
the head of the suction above
the level of the discharge. In
other words, the system curve
is plotted exactly as for any
other case involving a static
head and friction head, except
the static head is now negative.
The system curve begins at a
negative value and shows the
limited flow rate obtained by
gravity alone. More capacity
requires extra work.
41. MOSTLY LIFT- LITTLE FRICTION HEAD
The system head curve in the illustration below starts at the static head "H" and
zero flow. Since the friction losses are relatively small (possibly due to the large
diameter pipe), the system curve is "flat". In this case. the pump is required to
overcome the comparatively large static head before it will deliver any flow at all.
*Hydraulic losses in
piping systems are
composed of pipe
friction losses, valves,
elbows and other
fittings, entrance and
exit losse (these to the
entrance and exit to
and from the pipeline
normally at the
beginning and end not
the pump) and losses
from changes in pipe
size by enlargement
or reduction in
diameter.
46. CHARACTERISTIC
CURVE
OF PUMP
NPSH, Efficiency, HP, Impeller Diameter
Sebuah sistem perpipaan
mempunyai kapasitas 800
gpm dan head 26 ft.
Rancangan pompa yang
dipakai
•Diameter impeller 7.13”
•BHP=6.5
•Efisiensi = 80%
•NPSHr = 8 ft
47.
48. How to Read Pump Curves
STEP 1: The basic pump curves are no different than reading any other
head - flow curve. For a known head value, follow the head over to the
pump curve then drop down to the capacity axis and this will be the flow
rate. What you are trying to figure out here is what diameter impeller is
needed to get the required head and capacity.
STEP 2: The next thing to figure out is what motor is needed to drive
this impeller without overloading. To do this use the dashed horsepower
lines. To the right of the horsepower line is overloading and to the left is
non-overloading.
STEP 3: The last thing to determine is at what pump efficiency the pump
will operate. Look at the U-shaped lines and interpolate to get the
efficiency.
Now let's try an example using ZM1570, Performance Data for Models 6650-
6671 (5-15 BHp 4" discharge units). For the example we will size a pump for
400 GPM at 54 feet of total dynamic head.
49. STEP 1: Locate the point of 400 GPM at 54 feet on the pump curve. This
point is slightly above the 8.31" impeller but well below the 8.63"
impeller so I would go with an 8.38" impeller to hit the duty point.
50. STEP 2:
Next, draw a new pump curve that passes through the duty point and is
parallel to the existing pump curves. This will give you a close
representation of the actual performance the pump will deliver. Look to
see where this curve crosses the horsepower line to the right of the
design point.
In this example the pump curve crosses the 10 BHp curve at about 48
feet and crosses the 12.5 BHp curve at about 21 feet. We will not oversize
an impeller on a pump if the overload point on the pump curve is greater
than the static head for the system.
So for this example, if the static head is greater than 48 feet then we can
use the 10 BHp unit. If the static head is between 21 feet and 48 feet, use
the 12.5 BHp motor. If the static head is less than 21 feet then use the 15
BHp motor.
51.
52. STEP 3:
Now let's figure the pump efficiency we can expect. The design point
is about half way in between the efficiency lines of 60% and 63%. So,
for the design point of 400 GPM at 54 feet, we would expect about
61.5% pump efficiency.
As you can tell from the above example, we would consider oversizing
an impeller on a unit and not overload the unit due to engineering the
right pump for the system. If this were the case we would also able to
provide a more competitively priced unit since pricing is based on
motor size (i.e. smaller motors cost less).
The only exception to this rule is a single-phase unit. ZOELLER
COMPANY DOES NOT SELL SINGLE-PHASE UNITS WITH OVERSIZED
IMPELLERS because we feel that this will compromise the life of a
single-phase unit.
53.
54. Determining Flow and Head
The pump is installed and running, but how do you know if it is
operating at its design point? There is a simple way to check. Knowing
that a pump will provide a certain flow at a given head, we can
determine the point at which the pump is operating. To determine the
head, a few gage readings will be necessary. Take one reading from
the suction of the pump and one from the discharge after the system is
balanced and with all the control valves wide open. The difference
between the two gage readings will give you the head that the pump is
providing. Remember to convert your gage readings to feet of head.
Knowing the head and the impeller size, you can determine the flow of
the pump.
Now that we have the flow and head of the pump, let’s see how close
we are to the design point. Most often, the head will be less than what
we expected, and the flow will be more. Why does this happen? There
are many reasons, but it does no good to blame anyone. Let’s just fix
the problem.
55.
56. Solutions
Trimming the impeller is one of best solutions. Before we can trim the impeller,
we need to determine where the pump is operating. In the pump curve above,
let’s call point “D” the design point, and draw the system curve that corresponds
with that design point. Point “A” is where we actually are, which we determined
from our gage readings. Along with that is our actual system curve. Remember
that we are concerned with the actual system curve. This shows us how our
system operates, not how it was designed. Operational and design points are
often completely different.
We would like to be on the unmodified actual system curve, but where on that
curve? If our load has not changed and our heat transfer is the same, we want to
be at our design flow. That is“I,” the ideal point.
Trimming the Impeller
But how do we get there? Although it’s off our impeller curve, we can trim our
impeller down to the right size. In this particular case, our ideal impeller size falls
between 10-1/2” and 11-1/2 (actually about 11”). Fortunately, trimming an
impeller is not too difficult or expensive, and in fact it pays for itself very quickly.
Notice from the figure that when we trim our impeller we lose some pump
efficiency, but we’re more concerned about the cost of operating our pump and
that cost has dropped tremendously. In this case we have dropped from 85Hp to
40 Hp-that’s a lot. Even if your electric rates are low and you don’t operate all
year long, there is still the potential for great energy savings.
57.
58. Given:
Atmospheric pressure = 14.7 psi
Gage pressure =The tank is at sea level and open to atmospheric pressure.
Liquid level above pump centerline = 5 feet
Piping = a total of 10 feet of 2 inch pipe plus one 90° long radius screwed elbow.
Pumping =100 gpm. 68°F. fresh water with a specific gravity of one (1).
Vapor pressure of 68°F. Water = 0.27 psia from the vapor chart.
Specific gravity = 1
NPSHR (net positive suction head required, from the pump curve) = 9 feet
NPSHA = Atmospheric pressure(converted to head) + static head + surface pressure head - vapor pressure of your
product - loss in the piping, valves and fittings
Static head = 5 feet
Atmospheric pressure = pressure x 2.31/sg. = 14.7 x 2.31/1 = 34 feet absolute
Gage pressure = 0
Vapor pressure of 68°F. water converted to head = pressure x 2.31/sg = 0.27 x 2.31/1 = 0.62 feet
Looking at the friction charts:
100 gpm flowing through 2 inch pipe shows a loss of 17.4 feet for each 100 feet of pipe or 17.4/10 = 1.74 feet
of head loss in the piping
The K factor for one 2 inch elbow is 0.4 x 1.42 = 0.6 feet
Adding these numbers together, 1.74 + 0.6 = a total of 2.34 feet friction loss in the pipe and fitting.
NPSHA (net positive suction head available) = 34 + 5 + 0 - 0.62 - 2.34 = 36.04 feet
The pump required 9 feet of head at 100 gpm. And we have 36.04 feet so we have plenty to spare.
59. Given:
Gage pressure = - 20 inches of vacuum
Atmospheic pressure = 14.7 psi
Liquid level above pump centerline = 5 feet
Piping = a total of 10 feet of 2 inch pipe plus one 90° long radius screwed elbow.
Pumping = 100 gpm. 68°F fresh water with a specific gravity of one (1).
Vapor pressure of 68°F water = 0.27 psia from the vapor chart.
NPSHR (net positive suction head required) = 9 feet
Now for the calculations:
NPSHA = Atmospheric pressure(converted to head) + static head + surface pressure head
- vapor pressure of your product - loss in the piping, valves and fittings
Atmospheric pressure = 14.7 psi x 2.31/sg. =34 feet
Static head = 5 feet
Gage pessure pressure = 20 inches of vacuum converted to head
inches of mercury x 1.133 / specific gravity = feet of liquid
-20 x 1.133 /1 = -22.7 feet of pressure head absolute
Vapor pressure of 68°F water = pressure x 2.31/sg. = 0.27 x 2.31/1 = 0.62 feet
Looking at the friction charts:
100 gpm flowing through 2.5 inch pipe shows a loss of 17.4 feet or each 100 feet of
pipe or 17.4/10 = 1.74 feet loss in the piping
The K factor for one 2 inch elbow is 0.4 x 1.42 = 0.6 feet
Adding these two numbers together: (1.74 + 0.6) = a total of 2.34 feet friction loss in the
pipe and fitting.
NPSHA (net positive suction head available) = 34 + 5 - 22.7 - 0.62 - 2.34 = 13.34 feet.
This is enough to stop cavitation also.
68. Discharge Cavitation
Discharge Cavitation occurs when the pump discharge is extremely high. It normally occurs
in a pump that is running at less than 10% of its best efficiency point. The high discharge
pressure causes the majority of the fluid to circulate inside the pump instead of being allowed
to flow out the discharge. As the liquid flows around the impeller it must pass through the
small clearance between the impeller and the pump cutwater at extremely high velocity. This
velocity causes a vacuum to develop at the cutwater similar to what occurs in a venturi and
turns the liquid into a vapor. A pump that has been operating under these conditions shows
premature wear of the impeller vane tips and the pump cutwater. In addition due to the high
pressure condition premature failure of the pump mechanical seal and bearings can be
expected and under extreme conditions will break the impeller shaft.
Suction Cavitation
Suction Cavitation occurs when the pump suction is under a low pressure/high vacuum
condition where the liquid turns into a vapor at the eye of the pump impeller. This vapor is
carried over to the discharge side of the pump where it no longer sees vacuum and is
compressed back into a liquid by the discharge pressure. This imploding action occurs
violently and attacks the face of the impeller. An impeller that has been operating under a
suction cavitation condition has large chunks of material removed from its face causing
premature failure of the pump.
69.
70. The affinity laws express the mathematical relationship between the several
variables involved in pump performance. They apply to all types of centrifugal
and axial flow pumps.
With impeller diameter D held constant:
With speed N held constant:
Where:
Q = Capacity, GPM
H = Total Head, Feet
BHP = Brake Horsepower
N = Pump Speed, RPM
THE AFFINITY LAWS
When the performance (Q1, H1, &
BHP1) is known at some particular
speed (N1) or diameter (D1), the
formulas can be used to estimate the
performance (Q2, H2, & BHP2) at some
other speed (N2) or diameter (D2). The
efficiency remains nearly constant for
speed changes and for small changes
in impeller diameter
71. Example:
To illustrate the use of these laws, refer to Fig. 8 below. It shows the performance of a particular
pump at 1750 RPM with various impeller diameters. This performance data has been determined by
actual tests by the manufacturer. Now assume that you have a 13" maximum diameter impeller, but
you want to belt drive the pump at 2000 RPM.
72. The affinity laws listed under 1 above will be used to determine the
new performance, with N1 1750 RPM and N2 = 2000 RPM. The first
step is to read the capacity, head, and horsepower at several points
on the 13" dia. curve in Fig. 9 below. For example, one point may be
near the best efficiency point where the capacity is 300 GPM, the
head is 160 ft, and the BHP is approx. 20 hp.
This will then be the best efficiency point on the new 2000 RPM
curve. By performing the same calculations for several other points
on the 1750 RPM curve, a new curve can be drawn which will
approximate the pump's at 2000 RPM, Fig. 9. performance
Trial and error would be required to solve this problem in reverse. In
other words, assume you want to determine the speed required to
make a rating of 343 GPM at a head of 209 ft. You would begin by
selecting a trial speed and applying the affinity laws to convert the
desired rating to the corresponding rating at 1750 RPM. When you
arrive at the correct speed, 2000 RPM in this case, the corresponding
1750 RPM rating will fall on the 13" diameter curve.
