The casing, or volute as it is sometimes called, captures the fluid energy generated by the impeller and directs it out of the pump through the discharge nozzle.

Casing types

As with impeller designs, manufacturers each have their own “special design” for casings, but most can be described by how they fall into a few short categories. The first is how they are “split.” Within this category, you have axially or radially split.


Axial split casings have mating surfaces running on the same axis as the shaft. This design allows the maintenance department to open the pump for inspection without disturbing the bearings, seals, or piping.


The pump on the left is often referred to as a “Horizontal Split Case Pump,” leaving out the term “Axial.”  Although not technically correct, the word horizontal, in this case, denotes both the orientation of the shaft and the direction of the split. (“Axial Split Case Horizontal Pump” might have been a better designation.”)






Due to the sheer numbers, the term “Axial split case” is most often associated with horizontal pumps, but the design is not limited to this single arrangement. The illustration below is an example of a vertical split case design.

Radial split casings have mating surfaces that are at 90 degrees on the axis of the shaft, as the picture below shows.

This concept is used extensively by the manufacturers of both rubber and metal-lined pumps as it greatly simplifies the changing of worn parts.

Tangential discharge casings have a discharge flange centerline that is in a plane offset from the pump’s centerline. The smooth path out of the pump leads to higher pump efficiencies and less wear. The offset design, however, requires cast-in lifting lugs to handle the offset weight. It also calls for special offset piping connections that may need independent support structures.

Top left discharge configuration pump casing

This casing design is only self-venting when in a “top left” discharge configuration.






The centerline discharge casing has the discharge flange center in the same plane as the pump’s centerline.

Centerline discharge pump casing


The advantages this design offers include self-venting for vertical discharge arrangements, the casing hanging straight when lifted by the discharge flange, and the weight of the casing being equally distributed to both feet.





The additional change in flow direction as the fluid exits the pump does result in a very slight loss in efficiency. This may be seen as a problem by some engineers; however, others prefer working with a centred discharge flange, which simplifies piping layouts.

Centerline discharge pump casing

Most small and general industrial pumps are tangential, not centerline.

Recessed impeller casing, as the name denotes, has the impeller recessed back into the casing.  With the impeller in this position, it allows the casing to pass very large particles and or “stringy” slurries.

Recessed impeller pump casing


The max particle passageway is often similar to the discharge flange opening dimension.







Arguably, I have reviewed the most common terms you will hear when discussing casings. There are, however, a host of others. Terms such as double suction,  double discharge, split volute, diffusers, end suction, top discharge, and self-priming just to name a few.  Alas, subjects for another day.



Previously we have discussed net positive suction pressure but avoided the role vapour pressure may have on NPSHA. Today, we will shed some light on this subject.

The study of vapour pressure starts with developing an understanding of evaporation. Evaporation of a liquid is a concept most of us generally understand, but do we really know the intricacies of the process?

Let’s start by examining a beaker of material in a liquid state. If that material is water, the beaker at a molecular level would be full of H2O water molecules. If the water was cold, say 1 degree C, then these molecules would be moving slowly and would not possess much kinetic energy. The molecular attraction between the molecules would keep almost all the surface molecules contained.

Periodically, due to random collisions of molecules, one surface molecule may accumulate enough kinetic energy to break the molecular attraction and leave the water’s surface. This molecule is now in a gaseous state and has “evaporated”.

As the temperature of a liquid increases so does the kinetic energy of the individual molecules.  As the drawing below illustrates, the increase in energy within the molecules at the elevated temperature translates into more molecules obtaining enough energy to transition into a gaseous state.

If the beaker is open to the atmosphere, air currents carry these water vapour molecules away, and the level in the beaker slowly drops. However, if the beaker is sealed as in the illustration above, the water vapour is contained and pressure starts to build.

While evaporation is occurring within the sealed beaker, a few gaseous molecules lose kinetic energy due to collisions with other molecules and return to the liquid state. A process we know as condensation.

If the liquid within the sealed container is held at a stable temperature for a long enough period of time, an equilibrium is reached. That is to say, water vapour is being formed by the process of evaporation and at the same rate vapour is condensing back to a liquid state. The pressure at which a state of equilibrium is reached at any given temperature is known as the liquid’s vapour pressure at that temperature.

Some people crudely describe the vapour pressure as the push the liquid has to jump from a liquid to a gaseous state.  When the liquid is in a container that is open to the atmosphere, the force exerted by the vapour pressure is pushing against atmospheric pressure.

When water is heated to 100 degrees C (212 F) it has a vapour pressure of approx 14.7 psi.  The pressure exerted by the water, in an effort to jump to a gaseous state, is now equal to the downward pressure on the water exerted by atmospheric pressure at sea level.  Atmospheric pressure at this point is incapable of stopping the rapid formation of water vapour throughout the entire volume of the liquid, and the water boils.

As discussed in earlier blogs, pumps often rely on atmospheric pressure to push liquid into the eye of the impeller. If the vapour pressure is pushing up against the downward force of atmospheric pressure, then it is detracting from the pressure available to push liquid into the pump.  In terms of earlier blogs, it is reducing the net positive suction head available. (NPSHA)

Looking at the formula used to calculate the net positive suction head available,

NPSHA = Ha +/- Hs – Hf + Hv – Hvp

we see vapour pressure is shown as negative or subtractive in nature. This coincides with today’s discussions of how vapour pressure acts against atmospheric pressure and detracts from the pressure available to feed liquid to a pump.

Fortunately, the majority of liquids pumped are water-based and not particularly hot but if you have a more volatile liquid or is at an elevated temperature best consider its vapour pressure if you have low NPSHA or you may be in for pumping problems.

If you need help or advice, the Hevvy/Toyo’s team of application engineers are well versed on this subject and are always willing to help.



In the realm of heavy centrifugal pumps and industrial rotary equipment, ensuring the longevity and optimal performance of your assets is paramount. When it comes to repairs, there’s an often-underestimated gem in the world of maintenance: true OEM repair.

  1. Authenticity translates to reliability

When your industrial equipment undergoes an OEM repair, you are essentially breathing new life into it with components designed and crafted by the very same experts who created the equipment originally.

This ensures a seamless fit, impeccable compatibility, and adherence to the exacting standards that the manufacturer upholds.

  1. Cost Savings

By leveraging the expertise only a manufacturer can have of its products, OEM repair services conduct a thorough diagnosis to identify and replace only the damaged or worn-out parts, sparing you from the expense of an entirely new pump.

This financial advantage not only safeguards your bottom line but also liberates valuable capital that can be allocated to other critical areas of your business operations.

  1. Reduced Downtime

Downtime can result in significant losses. An OEM dedicated repair team, well-versed in the intricacies of your equipment, can swiftly pinpoint the issue; sourcing of the necessary components isn’t an issue as they often have the most-used parts in stock.

This efficiency can be particularly crucial if you operate around the clock or have tight production deadlines.


Industrial pump manufacturers always have available pump curves for fixed-speed pumps with full-size impellers, but what if you want to run the pump at a different RPM or use a trimmed impeller?  Well, the manufacturer should be able to provide you with a modified curve but if that is not easily obtainable just use affinity laws to modify the manufacturer’s standard curve.

Affinity laws are not courtroom jargon, just some simple formulas that can be applied to operating points on a pump curve to predict a new point when pump RPM or impeller diameter is modified.

Changing the Pump RPM: When the impeller diameter of a centrifugal pump is held constant the effect of changing the speed (RPM) is in accordance with the following formulas, where  N= RPM   Q = Capacity   H = Head, and  BHP = brake horsepower.

Q1/Q2 = N1/N2

H1/H2 = (N1/N2)2

 BHP1/BHP2 = (N1/N2)3

Changing the Impeller Diameter: The effect of trimming the impeller without changing the pump speed is virtually identical to what happens when you alter only the pump speed. Therefore the formulas are also very similar, as illustrated below, where  D= impeller diameter   Q = Capacity   H = Head and  BHP = brake horsepower.

Capacity: Q1/Q2 = D1/D2

Head: H1/H2 = (D1/D2)2

BHP: BHP1/BHP2 = (D1/D2)3

Note; Before  I show you how to use these formulas to generate a special curve there are a couple of cautionary notes.  Firstly, the accuracy of affinity laws or formulas diminishes as the percent change increases.  Generally changes of approximately 15 % or less still yield acceptable accuracy.  Secondly, when trimming an impeller you are modifying the effective length of the impeller vanes and that is what results in a modified performance.  Since the impeller vane on most pumps does not start at the center of the impeller the percent change in impeller diameter does not accurately reflect the percent change in vane length. Since impeller diameter is easier to work with than vane length, affinity laws substitute it. On high flow- low head impellers this substitution can introduce significant error and actual test tank data should be used to create modified curves instead of affinity laws.

Creating a curve for a special pump speed or impeller trim is basically the same procedure. I will therefore only demonstrate one today and that is the creation of a  pump “Head–Flow” curve.

Below is a pump curve for a 1750 rpm pump.  If we needed to create a curve for 1650 rpm we would start by calculating the shut-off point. The current shut-off point is zero flow at 125 meters as indicated by the green star. The new shut-off point at the reduced rpm would be calculated using the following two calculations

The new flow point at shut-off is calculated using the flow affinity formula  N1/N2=Q1/Q2


1750/1650 = 0/Q2     therefore  Q2 is zero M3/hr

The new head point is calculated using the head affinity formula  (N1/N2)2=H1/H2


 (1750/1650)2 = 125/H2    therefore H2 is 111 meters

Plotting zero M3/hr at 111 meters on the curve below, we establish the calculated shut-off point for the pump at an rpm of 1650.  (indicated by the blue star below)

pump curve

We would next calculate the new run-out point. The current run-out point is 17.2 M3/hr at a head of 85 meters, as indicated by the red star. The run-out point at the reduced rpm would be calculated using the same two formulas as used with the shut-off point



1750/1650 = 17.2/Q2     therefore  Q2 is 16.2 M3/Hr

The new head point is calculated using the flow affinity formula  (N1/N2)2=H1/H2


 (1750/1650)2 = 85/H2    therefore H2 is 75.6 meters

Plotting 16.2 M3/hr at 75.6 meters on the curve above, we establish the calculated run-out point for the pump at an rpm of 1650.  (indicated by the orange star above)

Additional points can be calculated to fill in the points between shut off and run out using the same procedure as used above, thereby filling in some points on the 1650 rpm curve as shown in the illustration below. (black stars)

pump curve

Finally, as shown below, connect the points, and you have a 1650 pump curve.

pump curve

In closing, this is how you can create a curve by yourself, however, if it is a Toyo pump curve, you may forget this blog altogether and just call 604-298-1213 and have our application team e-mail you what you need!



In a previous blog, I quoted the number 2.31 when discussing the relationship between water pressure  & water head.  Since then I have had a request to elaborate on the number 2.31 and where it comes from.  Frankly, with many gauges still calibrated in PSI  I still relate to the US system of measure. The quoted factor therefore only applies to pressure in units of PSI and water head expressed in feet. The basis of the factor 2.31 is as follows:

  • One cubic ft of water weighs 62.4 lbs and contains 1728 cubic inches. Therefore one cubic inch weighs 62.4/1728 or 0.0361 lbs.
  • A column of water with a cross-section of one square inch, that is 12 inches tall (one foot), would weigh 0.0361 x 12 or 0.433 lbs. Exerting a pressure of .433 lbs per square inch.
  • Based on this, a pressure of 1 lb per square inch would require a one square inch column of water equal to 1/.433 or 2.31 ft tall

PS —  In the metric system the conversion factor is 1 Kilopascal [kPa] = 0.101 974  Meters of water column [mH2O]  My apologies to my fellow Canadians for not providing the metric Stay tuned, and bye for now.


A couple of months ago I used the term “Bulk Density Ratio” and said that it was a subject for another day. So today is the day.

Pea gravel

The term bulk density ratio describes the relationship of a material or substance mass as found in a particular sample vs its theoretical mass, assuming “complete compactness” and no voids of any type.

For example, Pea gravel is made up of small particles of stone, and a solid block of stone usually weighs approximately 168 lbs per cubic foot  ( 2.7 times the weight of water, an SG of 2.7).

When we measured a particular sample of pea gravel we found that it weighed  108 lbs per cubic foot (1.76 times the weight of water). But stone is stone!! Why the difference?

Clearly, the difference is the pea gravel sample is full of voids between the stones. The ratio of solid stone to the sample of pea gravel is 168 lbs / 108 lbs or 1.55

Said differently,  pea gravel occupies 1.55 times more space than solid rock for the same mass of product. This ratio is referred to as the “Bulk Density Ratio”.

If we take the weight of the pea gravel sample and put it over the weight of a sample of the same volume of solid stone and then convert it to a percentage (108/168  X 100 =  64%) we find that stone occupies 64 % of the space. Assuming it is a dry sample, air must occupy the balance of 36% of the sample’s volume.

Great information but how is it useful in the pump world? Well, the answer to that question is that Bulk Density Ratio is the key to calculating production rates when moving products like sand or gravel that are often measured by volume.  The easiest way to explain this is by using an example.

You are a contractor and your job is to remove a sand bar made up of coarse sand that is 100 meters wide by 100 meters long by 5 meters thick, or 50,000 m3 of material.  If you can pump at a rate of 800 m3/hr of slurry with a density of 1.23,  how many hours must you operate?

You can assume the coarse sand has a dry Sg of 2.7. But the pumpage is not all sand, it is a slurry with water occupying the space between the solids.  The formula below can help you determine the percent solids, by volume.

Cv =( Sm-Sl)/(Ss- Sl) = (1.23 – 1.0)/(2.7 – 1.0) = .135 or 13.5 %

  • Cw = Percent solids by weight
  • Cv = Percent solids by Volume
  • Ss = Sg of the dry solid
  • Sm = Sg of the slurry
  • Sl = Sg of the liquid

At a slurry flow rate of 800 m3/hr this equates to .135 x 800 m3 or 108 m3 of “solid” rock per hour, but the end product is not solid!!  This is where you need to know a Bulk Density Ratio to properly estimate production by volume.

Applying the bulk density ratio of 1.55 for coarse sand would mean the contactor is removing the sand bar at a rate of 108 x 1.55 or, 168 m3/hr. Based on this rate of production, the project of moving 50,000 cubic meters being pumped at a rate of  168 cubic meters per hour will take 297hrs to complete.

Hopefully, this short blog helps to clear up some of the confusion around this subject, but if you still have some questions or need help on a specific project please feel free to contact our very competent applications team.

Bye for now


It is very important when working with centrifugal pumps that there is a clear understanding of how specific gravity, commonly abbreviated to just  SG, affects or doesn’t affect pump output in terms of head and or pressure.

Let’s first define SG. The dictionary definition is simply “ the ratio of the density of a substance to that of water”.   As shown below one cubic meter of water weighs 1000 kg, while an equal volume of granite weighs 2700 Kg. ,  2.7 times as much. Hence granite is said to have an SG of 2.7.

A word of caution! Whatever the sample size, there must be no voids. Materials such as sand or gravel must be adjusted by a factor to allow for the air space between the particles. This factor is referred to as the “bulk ratio”.  A subject for another day!

Turning attention to the head. It is sometimes defined as “the vertical distance (in feet or meters) from the elevation of the energy grade line on the suction side of the pump to the elevation of the energy grade line on the discharge side of the pump.”

When applied to a centrifugal pump’s output, it is more simply described as “the height a pump can lift a liquid to”.  For example, if a centrifugal pump tries to pump straight up a 300 ft pipe, but flow stops as the water in the pipe reaches 231 ft, the pump is said to have a shut-off head of 231 ft.

You may ask, why not express the output of a centrifugal pump in PSI? Well, the answer to that question is based on the operating physics of a centrifugal pump.

Centrifugal pumps use centrifugal force generated as a liquid is spun down the vanes of the impeller to create pressure within the pump’s casing/volute. The heavier the liquid, the more force or pressure is created.  Some pump manufacturers rate performance at a stated output pressure based on water, usually in terms of PSI, but the majority of pump manufacturers use the term head, as it does not vary with liquid SG.

The basis of the statement in bold above is illustrated to the right.  The pumps are identical, but the blue column represents water being lifted while the orange column represents a liquid with an SG of 1.5 being lifted.  Notice the output pressures vary in proportion to the SG while the rating for lift remains constant.

The centrifugal pump handling a liquid that has an SG of 1.5 is accelerating a liquid weighing 1.5 times that of water down its impeller vanes, creating a force (pressure) that is 1.5 times that of an identical pump pumping water.  The column of liquid, however, also weighs 1.5 times more than water. Hence it needs a pressure equal to the pressure on water times 1.5 to lift the heavier liquid to the same height.

The obvious question then becomes, if pressure gauges measure in PSI, how do we relate that to a pump’s output in ft or meters of the head?  The full explanation of that is again a subject for another day, but the equation below and the diagram to the right will get you by for now.




I hope today’s blog helps explain the relationship between SG, Head, and Pressure. As mentioned in today’s blog, I will in future blogs address the “Bulk Ratio” and fully explain the 2.31 ratio of head to PSI.

Stay tuned, and bye for now.


I  was recently in a discussion with an engineer regarding the use of a cutter fan to supply agitation. I was very surprised to find that the engineer, although having some pump experience did not understand the difference between the title objects. Not only could this lead to backward rotation and subsequent damage to a pump but the lack of understanding would also lead to the implementation of ineffective dredging techniques. Therefore today’s blog is on Cutter Fan agitators and shaft-connected Inducers.

The Cutter Fan also known as an Agitator is connected to the main shaft of the pump. The first manufacturer to employ this type of agitation was Hevvy/Toyo Pumps on their heavy-duty submersible dredge pumps. Their patented design utilized a curved three-blade stirring attachment that was threaded onto the pump shaft just below the suction inlet.

The cutter fan or “agitator” as it is sometimes called is typically protected by a stand attached to the bottom of the pump. For added protection on the larger pumps, Toyo places a stub shaft in between the pump main shaft and the cutter fan. Operationally, all cutter fans redirect a portion of the fluid heading toward the pump suction and push or “fan” fluid away from the pump. The agitation provided by the cutter fan dislodges solids and re-suspends them into a slurry. As these solids are drawn toward the pump inlet some of the slurries is redirected by the cutter fan back down into the solids deposit providing a more effective form of agitation than a jet of purified water. This redirecting of slurry/solids continues in a cyclic fashion forming a “pocket” of high solids content slurry directly in front of the pump suction inlet. This of course maximizes the number of solids being pumped, an important feature for any dredge pump.

The Inducer, like the cutter fan, is attached to the main shaft. It can be located anywhere in front of the impeller. The optimum location is entirely application-dependent. Technically said, its normal function is to raise the inlet head by an amount sufficient to provide the NPSHR, thereby preventing significant cavitation in the pump. In short, it can help in applications where initial priming is difficult or the fluid just refuses to flow well into the pump.  Below is a picture of a style of inducer.

In Summary, cutter fans push product away from the pump to aid in agitation while inducers help draw product into the pump. If your new pump arrives with an item that looks like an inducer or maybe a cutter fan, do not guess as to which it is and use it to confirm correct rotation. Since they are both normally attached to the shaft by some form of thread, reverse rotation may result in components unscrewing during operation and some very expensive repairs. Pumps always have a rotational arrow. The cutter fan,  and sometimes the inducer, can be excellent items to watch when bumping for rotation as they are easily visible, but when wiring the pump observe the marked arrow to obtain correct rotation!!  Once you have set rotation you can now look at the blade/vane angles and it will be easy to determine whether the mystery item is a cutter fan or an inducer.

Bye for now!


No manufacturer that has been around for any length of time makes a bad pump. The bad press that follows an early pump failure is generally the result of a misapplied pump. Unfortunately, it is often the pump’s reputation that takes the rap and not the circumstances that lead to the installation of the “wrong pump”. Any reputable supplier will want the client to be pleased with his pump purchase long after the initial installation. Having said that how does he get it right?

In 1941 the actor and comedian  George Jessel, when referring to “the people”, made the famous quote “Give ’em what they want.” The truth of the matter is in the pump world you need to “give them what they need. To do this the supplier needs to know as much as possible about the customer’s application.

The need to know

I once had a PA call me up and request a quote on a 40 HP submersible pump. When I asked for more info such as head, flow, liquid, and solids content he cut me off and said I don’t have time for this just give me a price for a 40 HP pump. As I was in no position to argue with him I sent him a quote for 6 different pumps. Needless to say, I got a call from the mine foreman the next day and we discussed the details of the application.  As it turned out the head/flow he needed required a 60 hp pump so I sent him a quote, copied in the PA, and soon after received an order.  Everyone was happy.

If the mine had taken delivery of a 40 hp pump, at best output/flow would have been below the required amount or possibly even non-existent. Would it have been a bad pump or a misapplied pump?

What to Know?

The simple answer here to is give as much information as possible. The more the supplier knows about the application the better he can respond with appropriate information. It is equally as important for the customer to offer up as much information as he can as it is for the supplier to seek all the application details.

Major pump suppliers like Toyo all have questionnaires that can be sent out to prospective customers to help them convey application details to the manufacturer. These are great tools that help both parties home in on the correct pump for the application. Even if the pump user is well versed in pumps and the application thereof, I would still recommend the completion of one of these questionnaires.

Providing specs like; head, flow, liquid type, etc are covered by all questionnaires but more general questions are sometimes overlooked or just not responded to. These can be critical to the selection process. Below I have listed a few:

  • Why is the customer looking to pump the liquids or materials? It is vitally important to obtain a project overview. This gives the supplier a feel for the customer’s end goal and keeps everyone on track while not getting locked into a fixed project plan. Thinking outside the box can sometimes lead to a better plan.
  • What is the anticipated life of the project? This will help the supplier determine if the customer needs premium products that will last a long time, or if he needs a more budget friendly product that will still last for the life of a shorter project.
  • Is the purpose of the application to relocate a liquid, a liquid that contains solids, or is it to use a liquid to relocate a solid. This along with the project life data, will help a supplier focus on the correct product line to offer.
  • Is this a new application or are the existing pumps giving problems? If it is replacing an existing pump, what type of pump is it and how is it failing?  No one wants to offer similar equipment to what is already there and suffer the same fate.

Who to know

I can not over-emphasize the need to obtain and share information. It is the best defense against failed pumps.  I know that all customers vary with regard to pump knowledge, some will need to seek help from within their organization to provide the details of their application.

However, you can rest assured that the sales and application team at  Hevvy/Toyo are here to support you and are just a phone call away.

Bye for now.


Pumps fail for a host of reasons, but as the chart below depicts, more than two-thirds of the failures relate to sealing device issues. With the movement to conserve water and minimize the quantity of waste liquids requiring treatment, the industry is relying more and more on mechanical seals. As the installed population grows so does the percentage of pump breakdowns resulting directly from mechanical seals.

I, like many pump guys, try and separate pump problems from mechanical seal problems. However it’s a package, and the pump supplier has to support the customer in any way that he can. With that in mind, I thought I should delve into some of the general pump requirements for a successful mechanical seal installation.

There are six distinct parameters that must be correct for a seal to provide maximum performance. Some seal designs are more able to accept dimensional inconsistencies than other designs, so specific tolerances must be obtained from the seal supplier prior to installation.

Just to be clear, the tolerances referenced in the diagrams below are for general reference only. You must confirm the tolerances required for specific seals with the manufacturer of that seal.

1)  Shaft run-out

Shaft run-out should typically not exceed 0,05 mm (.002”) TIR (Total Indicator Reading) at any point along the shaft for ball or roller type bearings. (1000 to 3600 RPM)


2)  Radial shaft movement

Radial shaft movement is generally limited to 0,05 – 0,10 mm (.002” -.004”)  for ball or roller type bearings. For sleeve or journal type bearings, values will generally be in the order of 0,10 – 0,15 mm (.004” – .006”)


3)  Concentricity, shaft to bore

Concentricity of the shaft to the seal chamber bore should normally be within 0,025 mm per 25 mm shaft diameter (.001″ per 1″ shaft diameter) to a maximum of 0,125 mm (.005″) TIR.

4)  Seal chamber squareness

Seal chamber squareness to the shaft centreline should be approximately 0,015 mm per 25 mm seal chamber bore (.0005″ per 1″ seal chamber bore). Note: make sure that shaft endplay does not affect the reading.

5)  Shaft endplay

Shaft endplay should generally not exceed 0,25 mm (.010”) TIR, regardless of thrust bearing type.

6)  Seal  preload

All seals have some type of spring mechanism to provide an initial compression of the seal faces, commonly referred to as a preload. Cartridge seals have a built-in preload that only requires the installer to remove set tabs prior to start-up.

The specification for the preload varies greatly dependent on the specific seal design. You must see the manufactures installation instructions for specifics.

I hope the information in this blog helps some readers confirm that their pumps are ready for a successful mechanical seal install.  Next blog I will try and touch on some operational tips on avoiding premature seal failure.