Hydrocarbon Processing Copying and distributing are prohibited without permission of the publisher
Email a friend
  • Please enter a maximum of 5 recipients. Use ; to separate more than one email address.

Control-valve seat leakage

08.01.2011  |  Sanders, D. ,  GE Energy, Atlanta, Georgia

Enhancing initial performance, extending service life of valves offers benefits

Keywords: [Valves] [seat leakage] [psi pressure] [PTFE seals] [piston ring seals] [pilot balanced designs] [standards]

Two of the leading causes of user concern regarding control-valve performance may surprise you. Field experience shows that the biggest contributor to control-valve maladies is often oversizing. It is a classic example of having too many cooks in the kitchen. Each of the engineers contributing to the final process specification adds what he or she believes is a constructive safety factor, and the resulting installed product is substantially oversized for the application. However, the focus of this article is what may be the second-leading cause of concerns: shutoff performance that does not meet the end user’s expectations.

In all fluid-processing industries, tight-sealing control valves are vital to ensuring product quality, efficiency, safety and environmental protection. But how do you determine how tight is tight enough?

Tolerance of leakage can vary widely from application to application; tight enough in one case can be overkill in another and insufficient in a third. And to top it off, the various industry standards that classify seat leakage in industrial valves fail to address some of the practical issues that confront valve manufacturers, specifiers and end users. In fact, it is quite possible to successfully specify, manufacture and test a valve according to a well-established industry standard, yet still experience less-than-satisfactory results in the field.

This article will help address the technical and practical issues related to seat leakage, discussing the fundamentals behind the governing industry standards and offering guidance that users can apply to enhance initial seat leakage performance and help extend the life of their valve assets.

Defining seat leakage.

Seat leakage is defined as leakage that is internal to a valve—between the inlet and outlet sides of the valve—when the valve is in its closed position. It is not limited to leakage across the valve seat, but also encompasses all leakage across the valve trim when the valve is in the closed position. Leakage across internal trim seals, such as piston rings, and across trim-to-body seals, such as gaskets, can be counted as seat leakage.

It is important to note that, while leakage through valve stem packing is of growing concern in the industry, governing industry standards address this type of leakage separately and do not consider it to be a form of seat leakage.

Industry gold standard.

ANSI/FCI 70-2, Control Valve Seat Leakage, published by the American National Standards Institute and the Fluid Controls Institute, is widely recognized as the defining standard for leakage in control valves. It categorizes seat leakage into six groups (Class I to Class VI). Generally, each higher class defines tighter or more stringent leakage criteria. While these classes are not linear in their progression, they follow a logical succession from lowest to highest class with some added footnotes.

Additionally, for each leakage class, ANSI/FCI 70-2 provides detailed test procedures and defines the maximum allowable leakage (MAL). The test procedures include provisions for using water and/or air as the test medium, along with allowable pressure and temperature ranges. IEC Standard 60534-4, Industrial Process Control Valves—Part 4, Inspection and Routine Testing (2006), published by the International Electrotechnical Commission, contains the same control valve seat leakage criteria.

Class I.

Class I leakage is relatively undefined and is simply stated as agreed upon between purchaser and manufacturer. It does not identify a test procedure or specify a standard test pressure and, therefore, does not define the maximum allowable leakage.

So why does it exist? While not stated in the standard, Class I is normally used when the valve specification has no leakage-specific criteria based on its application. Possible examples include a valve that is always in the open position, or a system in which the control valve is always accompanied by an adjacent remote-operated isolation valve. In these cases, the user has no expectation that leakage through the valve will negatively impact the process, so Class I may be appropriate.

Classes II–IV.

The definitions of the other leakage classes are not quite as liberal. The Class II, III and IV definitions share many common attributes and are easily discussed as a set. For each of these classes, there is a common test procedure that permits the use of water or air at a temperature of 50°F to 125°F and a pressure of 45 psig to 60 psig. The maximum allowable leakage for each class is expressed as a function of the rated valve capacity as follows:
• Class II: 0.5% of rated valve capacity
• Class III: 0.1% of rated valve capacity
• Class IV: 0.01% of rated valve capacity.

There are two observations to note at this stage. First, the difference between these classes is not trivial. Class III is five times tighter than Class II and Class IV is 50 times tighter than Class II.

Second, a value such as “0.01% of rated valve capacity” is not of much use to someone who wants to know how many cups or buckets or gallons of leakage per minute is acceptable for a particular valve. The fact is, “rated valve capacity” can be a rather elusive value to those who do not crunch valve data every day.

Fortunately, control-valve manufacturers rate valve capacities in terms of Cv, the universal flow coefficient for valves as defined by ANSI/ISA Standard S75.01.01, and a basic conversion to Cv can help put the “rated valve capacity” leakage criteria into a more tangible context.

Per ANSI/ISA Standard S75.01.01, Cv is defined as the number of gallons of water per min., at 60°F, that will pass through a flow restriction with a differential pressure of 1 psi. The equation is:

Cv = Flow coefficient
Q = Flow quantity gallons per min. (gpm)
SG = Specific gravity
dP = Differential pressure psi.

When quantifying seat leakage, the flow restriction is the subject valve. Assuming that valve capacity (Cv) is known, the equation can be solved for flow (Q) as follows:


The SG of water between 50°F and 125°F only varies from 1.0007 to 0.9887 and thus can be rounded to 1.0. Therefore, the MAL for Classes II through IV can be expressed in terms of gpm as a function of test pressure as:

• Class II:

• Class III:
• Class IV:

To illustrate how these equations play out in the real world, consider a common 10-in. nominal, cage-guided, metal-seated control valve that meets standard ANSI/ISA 75.10.02 face-to-face dimensions. A typical Cv for this type of valve is 950. Using 50 psi of water at 60°F, the MAL rates for this valve are:
• Class II: 33 gpm
• Class III: 6.7 gpm
• Class IV: 0.67 gpm.

Class V.

Class V represents what is commonly referred to as an “effectively zero-leakage” control valve. It should be remembered, however, that the Class V category still retains some allowance for leakage. And, as will be discussed later, a seemingly zero-leakage shop test does not always lead to zero leakage under service pressures and temperatures. Class V, however, is very tight and certainly takes a step far beyond Class IV.

A different test procedure is used to calculate maximum allowable leakage for Class V. Rather than a test pressure of 45 psig to 60 psig, Class V requires the use of water at a pressure differential that is +/-5% of the service pressure differential, not to exceed the 100°F-rated pressure of the valve body. The formula for MAL is:

Class V (ml/min.): (5 x 10–4) x (seat diameter in in.) x (test pressure in psi). (6)

For the purposes of comparison, consider the 10-in. valve in the previous example and assume that the service pressure equals the 50-psig test pressure. The Class V MAL would be:

Class V: (5 x 10–4 ml/min.) x (10 in.) x (50 psi) = 0.25 ml/min. = 6.6 x 10–5 gpm.

Roughly speaking, this is 10,000 times less leakage than the same Class IV valve’s MAL of 0.67 gpm.

Class VI.

Like Class V, Class VI bears no relationship to the other classes. The specified test procedure requires the use of air or nitrogen gas at a temperature of 50°F to 125°F and at a pressure differential of 50 psi.

The maximum allowable leakage is not given in equation form, but is instead presented as a table of values per nominal valve size. And, since the test pressure is fixed, the MAL values are discrete. The values range from 0.15 ml/min. for a 1-in. nominal valve to 28.4 ml/min. for a 16-in. nominal valve. Valves that are 8-in. nominal and smaller are also given an equivalent MAL in bubbles per minute. 

  Fig. 1. An unbalanced control
  valve plug. The solid plug
  leaves one potential leakage
  path at the seat. 


Several aspects of the ANSI standard can be confusing and frustrating for users. For starters, the maximum allowable leakage for Classes II, III and IV do not vary with service pressure. The test pressure for these classes is fixed between 45 psig and 60 psig, while most control valves operate at pressures above that range. For example, the 100°F maximum working pressure of ASME B16.34 Class 2500-rated valves is in the neighborhood of 6,250 psig.

It may be tempting to work around this by always opting for a Class V valve, but this is not a practical solution. As illustrated earlier, a Class V valve can be up to 10,000 times tighter than its Class IV cousin. That is obviously a huge leap in performance—one that, understandably, comes with a higher price tag. The frustration comes for a user faced with a situation in which a Class IV valve is not tight enough but Class V performance is far more than is required. The ANSI standard offers no middle ground.

Note that the 2006 revision of ANSI/FCI 70-2 permits the use of air at 50 psig for the Class V test. While this adds convenience, it negates one of the attractive attributes of the Class V criteria, which is the use of a test pressure that matches the service pressure.

Because the difference between Classes IV and V is so large, it may be prudent for the end user to consider the value of the process fluid when making a decision on leakage class. Over the life of a control-valve trim, when leakage means lost revenue or lost energy, the additional cost of an upgraded trim can pay for itself in a short period of time.

The parallel standard, IEC Standard 60534-4, does offer an intermediate step between Class IV and Class V, known as Class IV-S1. It uses the Class IV test pressure, with MAL defined as:
Class IV-S1 MAL = (5 x 10–6) x rated valve capacity.

Solving for gpm, the equation becomes:
Class IV-S1 MAL (gpm) =


or 100 times less leakage than Class IV.

Returning to the 10-in. example valve at 50 psig test pressure:
Class IV-S1 MAL = 0.0067 gpm.

For Classes II through IV, neither ANSI/FCI nor IEC supports the use of test pressures that approach service pressures. Class V MAL does at least use a representative (dP) value, but even it can leave the informed user wanting.

Recall Eq. 2, the driving equation behind the Cv definition:

This is the fundamental equation that the valve industry uses to engineer, sell and deliver flow capacity. It states that flow (Q)—or, for the purpose of this discussion, leakage—is a function of the inverse of the square root of differential pressure (dP).

Then consider the equation for MAL for Class V (Eq. 3):
Leakage = (5 x 10–4 ml/min.) x (seat diameter) x (dP).

Here, the flow is directly proportional to the pressure differential. Needless to say, Eqs. 2–3 produce diverging results.

If it can be assumed that the Eq. 2 curve reflects the fundamental relationship between flow and differential pressure, then even the more sophisticated Class V definition has to be taken in narrow context. It is no wonder that ANSI/FCI 70-2 states that “the standard cannot be used as a basis for predicting leakage at conditions other than those specified.” In other words, the standard should not be used to predict leakage at conditions other than the test conditions. This leaves the user with little or no prediction of field performance.

Also note that the 10,000:1 example shown previously only applies to the test pressure of 50 psi. Due to the disconnect between the driving equations for MAL and classical physics of flow through a restriction, comparison of the two equations at alternate pressures yields alternate ratios.

For instance, in the previous 10-in. example valve, the ratio at a 1,000 psi test pressure is:

(Class IV) / (Class V) = (3 gpm) / (0.0013 gpm) = 2,308.

This discussion of ANSI/FCI 70-2 provides some insight into the attributes to consider when looking at another commonly referenced leakage standard—MSS SP-61, Pressure Testing of Steel Valves.

MSS SP-61. The Manufacturers Standardization Society of the Valve and Fittings Industry publishes MSS SP-61. Among other topics, it addresses valve leakage, although it uses the term Seat Closure Test.

It is important to begin by acknowledging that this standard, by its own definition, does not apply to control valves, but instead to valves used in “full open” and “full closed” service. More specifically, it is intended for use with isolation, stop, and check valves. Nonetheless, its use has been creeping into control-valve specifications. Thus, it is important to address its capabilities and appropriate use.

Section 5 of MSS SP-61 defines test procedures as well as acceptance criteria of seat closure tests. The test pressure is specified as 1.1 times the 100°F rating, which, if applied to an ASME B16.34 control valve, means 1.1 times the cold working pressure.

The standard defines only one class of leakage, although either liquid or gas can be used as the test fluid. The maximum allowable leakage is specified as:

For a liquid test: MAL = (10 ml/hr) x (valve NPS) where NPS is nominal seat size in inches.

There are a number of modifying circumstances named in the standard that can impact test pressure and maximum allowable leakage criteria. For instance, there are provisions that, for certain sizes and classes of valves, permit the use of 80 psi air rather than 1.1 times the cold working pressure. Additionally, there are provisions that, for certain types of valves, permit the allowable leakage to be increased by a factor of four.

Because this standard specifically applies to on/off valves, rather than to control valves, it is difficult to interpret the exceptions in terms that clearly apply to control valves. To help prevent misunderstanding when these exceptions are applied, it is recommended that they be specifically identified and integrated into the product specification.

MSS SP-61 can be readily compared to ANSI/FCI Class V because both standards consider the nominal seat size and, in a roundabout way, both standards define MAL as variable in direct proportion to test pressure. Interestingly, the two standards converge when the test pressure is 320 psi. Any pressure above that value and MSS SP-61 is more stringent, while test pressures below that value show Class V to be more stringent.

API Standard 598. A full discussion of seat leakage would not be complete without recognizing API Standard 598, Valve Inspection and Testing, published by the American Petroleum Institute. Like MSS SP-61, API 598 is not intended for control valves but has begun creeping into the control-valve specification process, so the test parameters are presented here for comparison.

API 598 uses test pressures that are 1.1 times the maximum allowable working pressure at 100°F. The MAL is generally more stringent than ANSI/FCI 70-2, to the extreme that it is expressed in drops and bubbles per minute. Like MSS SP-61, this standard was not written for control valves and further discussion of the unique requirements of control-valve performance will reveal why.

So now armed with the factual data regarding the various standards, it is appropriate to discuss how tighter seat leakage is obtained with various trim designs. While this article will not provide an exhaustive exploration of control-valve trim designs, the primary methods used to address seat leakage will be treated.

Plug-to-seat interfaces.

Whether for rotary or reciprocating valves, the most basic trim design is a simple metal-to-metal interface between the valve seat and valve plug. This interface is the primary path for seat leakage, and its design and control is the single largest contributor to leakage differentiation.

From a broad viewpoint, it would appear that controlling this leakage is a simple matter of the appropriate mating of parts. However, even when mating high-quality machined surfaces together, microscopic imperfections can allow leakage. Additionally, in spite of superior plug-to-seat geometry, if the parts are not self-aligning within the trim assembly, the surfaces will not contact appropriately and repeatedly across multiple open and close cycles, and the desired seat tightness will not be achieved.

For example, it is clear that a simple drilled hole that is 0.12-in. in diameter would be considered a leakage path of unwanted magnitude. Although its equivalent flow area is only 0.011 square in., when tested to Class V standards at 1,000 psi, this hole would lead to flow of roughly 10 gpm. However, if the same flow area were evenly distributed around the circumference of a 10-in. plug-to-seat interface, it would amount to a gap of only 0.0004-in., or approximately 1/10 the thickness of a human hair. Class V leakage for this 10-in. valve at 1,000 psi would only allow 0.001 gpm—1/10,000 the rate of the single-orifice flow. Obviously, there is a need for more than simple dimensional control of the mating parts.

For this primary seat, the plug-to-seat contact geometry and alignment are critical. The plug should have some means of self alignment when approaching the seat in the closed configuration. This can be accomplished with appropriate lead-in angles and/or continuous guiding along the stem or cage. The plug-to-seat contact geometry should result in a single line of contact around the circumference of the parts. This creates large unit loading, which is critical to closing the microscopic irregularities of the mating surfaces.

Compliant material interfaces. One method used to address the inherent imperfections in the plug-to-seat interface is to utilize a compliant member at the interface. This can be accomplished with soft materials, such as PTFE, that are embedded in either part. The compliant material conforms to the mating imperfections and provides the desired closure of micro-gaps. This trim configuration would normally be used in achieving ANSI/FCI 70-2 Class VI leakage.

These compliant materials can be used in applications with temperatures up to 600°F.

A variation of the truly compliant material interface is to use a metallic member on one component that is softer than its mating part. This can be accomplished by mating the softer grades of 300 series stainless steel to harder stainless or alloy steels. The theory is that, when mated together under loading, the harder material actually causes the softer metal to deform. For this reason, these parts are often overloaded at assembly to create a coining effect. It is important to note that this coining results in a unique matched pair assembly, and subsequent substitution of parts, and even disassembly and reassembly, negates the matched affect.

Key specification considerations.

A fundamental element of all plug-to-seat interfaces is that part loading between the two components can greatly affect leakage performance. This loading effect is applied by the force of an actuator, which is always present in control-valve applications. During the specification process, the valve manufacturer should provide guidance (based on an understanding of the behavior of the assembled parts and their tendency to align, mate and conform) to ensure appropriate plug-to-seat loading.

Because the various components must work together to deliver the required loading, particular care must be taken when pairing a control valve and actuator from different vendors. The separate parties, or a third-party consolidator, may not coordinate the offerings. The end user must, therefore, understand the required load and ensure that it is provided by the actuation platform and supply pressure.

And because actuation load is a function of not only actuator type and size, but also of supply pressure, it is important that the party responsible for packaging the valve and actuator factor in the entire range of possible supply loads—normal, minimum and maximum—as they can positively or negatively impact seating performance. The obvious misstep would be a supply pressure that is too low to provide the necessary part loading under service conditions. It is equally important, however, to ensure that the actuator does not overload the parts and deform the trim or body components or cause the plug and stem to engage too tightly and become wedged together, negatively impacting the control valve’s ability to respond to an opening signal.

An additional feature of these simple plug-to-seat interfaces is that the direction of flow, and thus fluid pressure, can either assist or detract from part loading. The common terms used to determine pressure and flow tendencies relative to trim design are flow-to-open (FTO) and flow-to-close (FTC).

FTO trim designs. In FTO trim designs, the flowing fluid and its associated static and dynamic pressures tend to force the valve plug off of the seat, thus compromising seat load and increasing leakage tendency. For this type of trim, it is important that the worst-case conditions are factored into the design so that fluid pressures do not relieve the necessary minimum seating load that closes those micro-gaps. This is especially important in applications involving metal-to-metal seat interfaces.

FTC trim designs. Alternately, FTC trim designs are assisted by the fluid pressures, potentially reducing leakage tendency. Again, there is opportunity for misapplication when mating valves and actuators. On one hand, the actuator must be adequately sized to overcome the worst-case active fluid pressure loads. However, if oversized, the combination of actuator loading and fluid pressure can overload a plug-to-seat interface and cause damage, manifested in excessive part friction and sticking.

It is reasonable to question the use of FTO trim designs versus their FTC cousins, when FTC would appear to provide the most generous seat loading and, thus, more effectively reduce leakage. While this argument is valid, FTO trim designs provide other beneficial performance traits. They are inherently more stable, as the flow under the plug does not create the recirculating flow eddies that can cause the valve to be sucked into the closed position, a concept known as negative gradient. Additionally, an FTO design may be preferable based on the desired failure mode when actuation energy is lost.

In many trim designs, the seating surfaces of both the valve seat and the valve plug are in the area of flow modulation. In fact, when a valve plug is lifted just off of its seat, the inherently small available flow area leads to high fluid velocity, creating prime conditions for deterioration of component geometry. Many designs attempt to address this issue by providing geometry that shields the critical surfaces from the erosive effects of high-velocity flow. Proper materials selection is also critical to ensuring that these surfaces survive the rigors of modulating flow. One of the most straightforward means to enhance service life in these areas is to provide high-hardness materials via base materials, process hardening or weld overlays using hardened alloys.

Collectively, the combination of plug-to-seat interface geometry, materials selection, actuation loads and pressure assist forces all come together to create an effective primary seating surface. The requirements for control valves and on/off valves are quite different, however. Control valves must be able to move off of their seated position with a subtle change in instrument signal. If the above combination creates friction that is difficult for an actuator to overcome, or if the friction changes over time, then the control valve is rendered ineffective in a manner that would be less problematic for an on/off valve.

Balanced trim designs. As noted previously, these brute force, or unbalanced, trim designs often require large actuation forces, and thus costly actuation packages. Early solutions included the advent of double seated valves (Fig. 2). In these designs, the valve seat geometry is arranged such that high-pressure fluid is present on both sides of the valve plug, providing the necessary force balancing to minimize actuation loads. These designs still exist but their use is largely limited to smaller nominal sizes. 


  Fig. 2. In double seated
  valves, the valve seat
  geometry is arranged such
  that high-pressure fluid is
  present on both sides of the
  valve plug. 

The challenge with double seated designs is that it is difficult to exactly match the dimensional values between the two valve body seat regions and the two plug seating regions. Even when good dimensional matching is provided, thermal excursions during service can cause differential thermal growth and loss of seat contact.

Control-valve manufacturers have addressed this dimensional challenge with the use of balanced trim designs (Fig. 3). Conventional balanced trims are available in cage-guided packages, in which the plug travels inside a ported cylinder. These designs still depend on an effective plug-to-seat interface for their primary sealing interface, but they also have passages that provide flow, and thus pressure, communication between the top and bottom of the plug. This permits pressure to equalize on both sides of the plug, and actuation forces are substantially reduced.

  Fig. 3. Conventional balanced
  plugs have passages that
  provide pressure
  communication between the
  top and bottom of the plug to
  minimize actuation forces. 

A complicating factor in balanced designs is that a circumferential seal is required between the plug and cage to prevent flow from passing through the balancing holes, passing between the plug-to-cage clearance and then exiting the trim. These seals are known as the secondary leakage path, and it is here that the various designs show differentiation. The seal can be either plug-mounted and work against the stationary cage, or cage-mounted and work against the modulating plug. They can be compliant materials, such as PTFE or flexible graphite, or metal seals, such as classic piston rings.

PTFE seals. The design range of PTFE has recently been extended into the 600°F realm and these seals show good resistance to the expected frictional wear of a modulating seal. PTFE can also be formed to create pressure-assisting lips that improve the contact between the stationary and modulating parts. Above the 600°F temperature threshold, however, PTFE is not viable and flexible graphite is the only widely available compliant option.

Flexible graphite seals. Like PTFE, flexible graphite provides the necessary compliance to dynamically deform to fit its mating parts. The primary issue with graphite is that the normal frictional wear of a modulating control valve quickly leads to loss of mass, and thus volume, in the graphite seal, causing it to lose its radial contacting load.

Unlike graphite stem packing, internal trim seals cannot be adjusted by tightening a packing bolt or using a live loading mechanism. Instead, the frictional wear leads to increased leakage at the secondary plug seal, and even though the primary plug-to-seat geometry remains effective, the user experiences unwanted leakage.

Metallic piston ring seals. For high-temperature applications, these secondary seals are more typically metallic. Conventional piston ring designs are widely used as trim seals for balanced designs. A key differentiator among various piston rings is the design of the gaps where the open ends meet. More advanced piston ring designs often use an inner expander ring to create continuous radial loading of the outer ring. Good piston ring design leans on installed circularity, face sealing with the mating groove, and an ability to install the ring without deforming it.

Note also that for cage-guided valves, these secondary piston ring seals also provide the dynamic stability that is necessary for valve plugs with balancing holes. These rings prevent continuous flow across the balancing holes, which can cause pressure pulsations between the top and bottom of the plug and compromise valve stability. This is one of the many areas where control-valve trim has additional dynamic requirements that are not required in on/off valves.

Advanced solutions.

Some recent innovations in secondary sealing technology attempt to combine many of the previously discussed attributes into a single advanced seal. These include metallic seals that are also compliant to provide extended wear. Some compliant metals seals have been utilized to remedy the problems associated with double seated valves by providing dimensional forgiveness between the two seating regions. Some designs also have used separate dynamic seals along with static seating seals to keep the wear of modulation isolated from the static seal.

Each of these secondary sealing techniques has merit, but each also has its Achilles Heel in such areas as durability, scalable size, or temperature gradients. The user should seek evidence of long-term application experience in installations that are similar to the proposed application.

Shop-tested vs. installed leakage. A final topic that must be addressed in this discussion of secondary seals is the issue of shop-tested leakage versus installed leakage. A graphite seal in a new assembly, for example, will typically pass a shop test with ease, even when designed to meet ANSI/FCI 70-2 Class V or MSS SP-61 criteria. However, a short time in modulating service can lead to loss of graphite mass and compromised leakage performance. The fact remains that none of the seat leakage standards makes any guarantees regarding in-service leakage. They simply represent stand-alone shop tests of new components, and thus can be misleading if the user attempts to equate them to field service performance.

Pilot-balanced plugs.

A popular option for high-temperature applications that cannot use nonmetallic compliant seals is the pilot-balanced plug (Fig. 4). This design is a hybrid of unbalanced and balanced designs, incorporating the best attributes of both.

  Fig. 4. Pilot-balanced
  designs have a small
  unbalanced plug that is
  integrated into the larger
  primary plug to create a
  hybrid balanced design. 

In the pilot-balanced concept, a small unbalanced plug is integrated into the larger primary plug. When the pilot plug is opened, it provides a balancing path for pressure equalization across the primary plug. When both the pilot plug and primary plug are closed, the assembly is effectively unbalanced.

Pilot-balanced plugs are time-tested and in widespread use. The greatest challenge in working with them is ensuring that the proper stiffness is maintained during mid-travel modulation, as inadequate stiffness can lead to instability. Most designs use internal springs, either a coil or Bellville type, to maintain a target stiffness between the pilot plug and the primary plug. Others incorporate pressure ports to utilize the adjacent high-pressure fluid in a chamber with a pressure area biased toward stiffness.

The positioner’s role.

In addition to the mechanical variations that can impact seat leakage performance, the “brain” on top of the control valve—the positioner—can be an important tool in protecting critical seating surfaces. As mentioned earlier, valve trim that operates just off of its seat creates small flow passages and the resulting high velocities within the trim are highly erosive, even in “clean” fluid applications. Many of today’s digital positioners have options for an override that prevents the plug from operating in this unfavorable travel position. These overrides are typically user-configurable, allowing the user to determine the desired minimum travel position relative to both trim protection and required low-end capacity.

Additionally, the most advanced digital positioners can remember their shop-tested attribute of stem position when the trim is seated. This is quite useful if foreign material or damaged trim components prevent the plug from properly seating on the valve seat. Even though the actuation force may suggest that the plug has reached its proper travel stop, these smart positioners know that this position does not signify true seating and can provide the necessary alarm to process control personnel.

Clearly, there are a multitude of factors that influence seat leakage measurement and performance in control valves. While there are various industry standards that can be utilized in the specification process, ANSI/FCI 70-2 is the most widely used despite its shortcomings. Given that a shop-based seat leakage test provides limited prediction of field performance, the end user is wise to understand the more durable attributes that lead to satisfactory seat leakage performance over the installed life of a control valve. HP

The author 

Don Sanders is product manager for engineered products and the severe service segment for GE Energy, www.ge-energy.com. A 29-year veteran of the control-valve industry, he has held positions in engineering, manufacturing, marketing and management. He has a BS degree in mechanical engineering from the Georgia Institute of Technology

Have your say
  • All comments are subject to editorial review.
    All fields are compulsory.


a vent valve has 11bar differ but in close 3o percent open what is problem

YTC Positioner

Thanks for sharing this kind of information.This post will also help to those people who have less technical knowledge and they can easily operate their system.
YTC Positioner


It was a useful cumulative lesson for the valve passing, many thanks,

Ed Knuckles

Useful and practical insight into various aspects of control valve seat leakage; many thanks


can oversized control valve could contribute to leakage?


Really interesting information; I'm still having problems to calculate the allowed leakage for Class IV valve with a CV of 41 using 50 psi of air.

can anybody help?

ahmed messallam

valuable deep information of leakage inside the body of the valve.if possiple i want more information about this aricle.
thank you in advance.


it's creat articale and very important.
(i need to know, how to select leak classe for shut off valves).


Yamini Panchal

Don, this is a great article. Helped me a lot. :)
Please can you also mention about Leakage in Spool Type Control Valves.

Gerardo Ramirez

Excelent publication Don. I was not aware about it till today when a customer (Mr. Santiago Arizpe from Cerrey) told about it. I am in Dresser Mexico.

Shobhit Desai

Is seat leakage applicable for solid service???


I want leakage rate chat as per Cv in all Leakage class?

Adeel Qaiser

Mr. Don you have cleared so many of my doubts regarding seat leakages Thank you very much,this is the first time i am commenting on a article.

Sonal Singh

Excellent article..Precisely explained

Sonal Singh

Excellent article..Good insight about different standards...precisely explained..

John Simmons

Very good article, Don.

Related articles


Sign-up for the Free Daily HP Enewsletter!

Boxscore Database

A searchable database of project activity in the global hydrocarbon processing industry


Is 2016 the peak for US gasoline demand?




View previous results

Popular Searches

Please read our Term and Conditions and Privacy Policy before using the site. All material subject to strictly enforced copyright laws.
© 2016 Hydrocarbon Processing. © 2016 Gulf Publishing Company.