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H. A. WardaMechanical Engineering Department,

Alexandria University,

Alexandria 21544, Egypt

e-mail: [email protected]

E. M. Wahba1

Mechanical Engineering Department,

American University of Sharjah,

Sharjah 26666, United Arab Emirates;

Mechanical Engineering Department,

Alexandria University,



E. A. Selim

Mechanical Engineering Department,Alexandria University,



Integral Pumping Devicesfor Dual Mechanical Seals:Experiments and Numerical

Simulations An experimental and numerical investigation is carried out to evaluate the performanceof alternative pumping ring designs for dual mechanical seals. Both radial-flow and axial-flow pumping rings are considered in the present study. An experimental setup isconstructed, and appropriate instrumentation are employed to measure the pressure, temperature, and flow rate of the barrier fluid. A parametric study is carried out to investi-gate the effect of pump rotational speed, barrier fluid accumulator pressure, and barrier fluid inlet temperature on the performance of the pumping rings. Experiments are alsoused to evaluate the effect of different geometric parameters such as the radial clearancebetween the pumping ring and the surrounding gland, and the outlet port orientation. Moreover, a numerical study is conducted to simulate the flow field for the radial pump-ing ring designs under different operating parameters. The computational fluid dynamics(CFD) model implements a multiple reference frame (MRF) technique, while turbulenceis modeled using the standard k-epsilon model. Numerical simulations are also used to

visualize the flow of the barrier fluid within the dual seal cavity. Present results indicatethat the pump rotational speed and the orientation of the outlet port have a significant effect on the performance of the pumping ring. On the other hand, the effects of barrier fluid accumulator pressure and inlet temperature are minimal on the performance. Thestudy also shows that reducing the radial clearance between the rotating ring and the sta-tionary outer gland would significantly improve the performance of axial pumping rings. Moreover, comparisons between the computational and experimental results show good agreement for pumping ring configurations with tangential outlet (TO) ports and at moderate rotational speeds. [DOI: 10.1115/1.4028384]

Keywords: dual mechanical seal, integral pumping device, radial pumping ring, axial pumping ring, computational fluid dynamics

1 Introduction

Mechanical seals are commonly used in turbomachinery to pre-vent leakage of fluids [1]. The mechanical seal typically consistsof two parts: a rotating seal and a stationary seal. The rotating sealis fixed to the shaft and rotates with it, whereas the stationary sealis mounted on the housing. The primary sealing occurs at theinterface between both seal faces. O-rings are used as secondaryseals to prevent leakage between the rotating shaft and the rotatingseal, and also between the housing and the stationary seal. For proper functioning of the mechanical seal, a fluid film should bemaintained between the faces.

Present legislation and environmental issues surrounding thecontainment of pumped fluids dictate that the consequences of leakage can no longer be tolerated for many liquids. As a result,dual mechanical seals were introduced where a barrier fluid isused to fill the space in between the two sets of seal faces. The

barrier fluid also acts as a coolant to remove heat from the seal[2]. Providing an adequate barrier flow rate is, therefore, crucialfor removing heat in order to sustain proper functionality and toprevent premature seal failure. The barrier fluid is pumped from aseparate tank through an inlet port into the seal. The fluid then cir-culates through the seal, picking up heat and is removed from theseal via an outlet port. It then flows back into the tank, forming aclosed loop.

The barrier fluid is circulated within the closed loop through ei-

ther a convective thermosyphon system, an integral pumping de-vice or via an external pumping device [3]. In general, an externalpumping device is more expensive to run, and it consumes moreenergy. Moreover, integral pumping devices are preferred over thermosyphon systems. This is due to the fact that seal face tem-peratures should be kept relatively cool and sufficiently below theboiling temperature of the liquid film between the faces, which ismore difficult to achieve using the thermosyphon system. Integralpumping devices ensure cooler operation, which explains whythese devices have generally been well received in industry. Sev-eral types of integral pumping devices are currently available inthe market [4]. Thus, it is important to study and understand thefundamental differences between them. Understanding these dif-ferences will assist in improving the design of such pumping devi-ces and help in selecting the most advantageous device for

different applications.A very limited number of studies in the literature investigatedthe performance of integral pumping devices and the flowdynamics inside the dual seal cavity. Clark and Azibert [2] usedFLUENT to simulate the turbulent flow field in the barrier fluiddomain of a dual seal in a centrifugal pump. They concluded thatdesign features favoring axial circulation include larger radialgaps between rotating and stationary components, as well as axi-ally tapered surfaces that would propel cooler fluid toward theheat-producing interfaces of the seal. Nevertheless, a larger radialgap can also increase the risk of leakage. Moreover, Carmodyet al. [5] studied the barrier fluid flow characteristics of doublemechanical seals in API 682 sealing arrangements. In this study,Carmody et al. [5] used CFD to simulate and improve the flow

1Corresponding author.Contributed by the Structures and Dynamics Committee of ASME for publication

in the JOURNAL OF E NGINEERING FOR G AS TURBINES AND P OWER. Manuscript receivedJuly 11, 2014; final manuscript received July 18, 2014; published online September 16, 2014. Editor: David Wisler.

Journal of Engineering for Gas Turbines and Power FEBRUARY 2015, Vol. 137 / 022504-1CopyrightVC 2015 by ASME

wnloaded From: http://asmedigitalcollection.asme.org/ on 12/02/2014 Terms of Use: http://asme.org/terms


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characteristics through the internal cavities of typical seals. Car-mody et al. [5] demonstrated the possibility of providing adequatehead and flow rate requirements for double seals at clearancesspecified in the API 682 standard [3].

The present study aims to fill the gap in the literature by provid-ing a detailed experimental and numerical investigation of the per-formance of a group of radial- and axial-flow integral pumpingdevices. The radial-flow devices include two configurations: astandard paddle wheel (SPW) design and a modified paddle wheel(MPW) design. The axial-flow devices also includes two configu-rations; a single pumping scroll (PS) complying with the API 682standard [3]; and a single PS with a tighter-clearance configura-

tion. A parametric experimental study is carried out to investigatethe performance of the pumping rings under different operatingconditions and also to reveal the effect of different geometric pa-rameters such as the radial clearance between the stationary glandand the pumping device, and the orientation of barrier fluid outletport. Numerical simulations are also carried out to provide a better understanding of the flow dynamics within the dual seal cavity.

2 Experimental Setup

The experimental setup consists of two main loops: the processfluid loop and the barrier fluid loop. As shown in Fig. 1, the pro-cess fluid loop consists of a pressurized tank, a centrifugal pump,two pressure transmitters, and two temperature transmitters. The

centrifugal pump has an open type impeller with a diameter of 8.1875 in. and is operated by a 3.5 kW electric motor.A frequency inverter is used for changing the frequency of theelectric motor and thus controlling the rotational speed of the

Fig. 1 Schematic diagram of the process fluid loop and the barrier fluid loop. (a ) Process fluid loop and (b ) barrier fluid loop.

Fig. 2 Radial-flow and axial-flow pumping ring designs. (a ) Standard paddle wheel, (b ) modi-fied paddle wheel, and (c ) single pumping scroll.

Fig. 3 Tangential and RO port orientations for the MPW. (a )Tangential outlet and (b ) radial outlet.

022504-2 / Vol. 137, FEBRUARY 2015 Transactions of the ASME



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pump to be able to study the performance of the pumping rings atdifferent speeds.

The pump receives water from the bottom side of the pressur-ized tank and the discharged water is returned back to the top sideof the tank. The tank is pressurized through a compressed air

source connected to a ball valve installed at the top side of thetank. The pump discharge is controlled through a gate valve in-stalled at the pump discharge port. In order to facilitate pump dis-assembly, a ball valve is installed at the bottom side of the tank tobe able to isolate the pump for maintenance purposes.

Figure 1 also shows the barrier fluid loop, which consists of anaccumulator-based seal support system, the mechanical seal, twopressure transmitters, two temperature transmitters, and a flow me-ter. The test mechanical seal is John Crane type 48. It is a 1.875in. diameter, multiple-spring dual pusher seal with carbon/SiC/car-

bon/SiC faces and fluorocarbon rubber elastomers throughout. Thepressurized barrier fluid circulating within this loop is responsiblefor cooling the mechanical seal. Pressurized water enters the sealfrom the inlet port at the bottom, where it picks up heat within theseal and comes out through the top outlet port. Water is thendirected to the accumulator-based seal support system, where it iscooled through a water-cooled heat exchanger. Cooling water issupplied to the heat exchanger through an external tank which isfitted with a 0.37 kW centrifugal pump for water circulation. Thepressures and temperatures of the barrier fluid are measured usingpressure and temperature transmitters installed at the seal inlet andoutlet ports. This loop also includes a Coriolis mass flow meter for the measurement of barrier fluid flow rate. The experimentalsetup is used to investigate the performance of a group of radial-and axial-flow pumping rings. In what follows, the investigatedpumping ring configurations are presented in detail.

2.1 Radial-Flow Integral Pumping Devices. Two radial-flow pumping ring designs are investigated in the present study; aSPW and a MPW. Figure 2 provides a three-dimensional model of the SPW. The slots on the circumference of the wheel carry thefluid as the shaft rotates. When each slot passes by the dischargetap located in the seal housing, the fluid is pushed out into the sealpiping. A three-dimensional model of the MPW is also given in

Fig. 4 Overall and close-up views of the medium grid for theMPW with a TO port. (a ) Overall mesh, (b ) modified paddlewheel mesh, (c ) seal inlet port mesh, and (d ) seal outlet portmesh.

Fig. 5 Performance curves for the different pumping ring designs

Journal of Engineering for Gas Turbines and Power FEBRUARY 2015, Vol. 137 / 022504-3



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Fig. 2. The MPW design is introduced in an effort to reducethe effect of eddies and internal recirculation that occur in theclearance between the pumping ring and the surrounding outer gland.

2.2 Axial-Flow Integral Pumping Devices. The single PS,shown in Fig. 2, belongs to the family of axial-flow pumpingrings. It consists of a rotating unit with an outer helical thread.Unlike the screw thread of a fastener, these screw threads have arectangular cross section and multiple leads. The fluid entersthrough the inlet port, then tangentially through two ports into theseal chamber. Then, the fluid is transferred axially through thehelical grooves of the ring, which lead to the outlet port.

Two versions of the single PS are investigated in the presentstudy. The API-complying single PS is a version of the singlescroll in compliance with the API 682 standard [3]. The main dif-ference between this API-complying version and the second APInoncomplying version is the radial clearance between the pump-ing ring and the outer gland. API 682 standard defines a radialclearance of 1.5mm between the pumping ring and the outer gland. Nevertheless, some dual seals applied in industry aredesigned with radial clearances in the order of 0.25 mm and0.5 mm. The API noncomplying version investigated in the pres-ent study uses a radial clearance of 0.5mm. The importance of specifying the clearance stems from two sets of conflicting

requirements. One is to make the radial clearance safe and contactfree as per the API 682 standard [3], and the other is to provideadequate barrier fluid flow rate to remove heat from the seal. Theconflict arises when many integral pumping devices become

inefficient as the radial clearance is increased to comply with theAPI 682 standard [3].

2.3 Orientation of Outlet Port. To investigate the effect of the outlet port orientation on the performance of the radial andaxial pumping rings, experimental runs are carried out using twooutlet port orientations: a radial outlet (RO) port and a TO port.Figure 3 shows the two outlet port orientations for the case of theMPW.

3 CFD Model

In what follows, details of the CFD model are presented interms of the governing equations, numerical methods, boundaryconditions, and grid generation.

3.1 Governing Equations and Numerical Methods. TheCFD model is based on the Reynolds–averaged Navier–Stokesequations:

@ u j

@ x j ¼ 0 (1)

@ ðuiu j Þ

@ x j ¼

1

q

@ p

@ xiþ

1

q

@

@ xil

@ ui

@ x j þ

@ u j

@ xi

þ

1

q

@ sturb

@ xi(2)

Turbulence is modeled using a two-equation eddy viscositymodel, namely, the standard k-e model [6]. One drawback of eddy-viscosity models is that they are insensitive to streamline

Fig. 6 Effect of rotational speed on the performance of the integral pumping device. (a ) SPW, RI TO, (b ) MPW, RI TO, (c ) PS, TITO, and (d ) PS, API, TI TO.




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curvature and system rotation. A modification of the productionterm has been derived by Smirnov and Menter [7] to sensitize thetwo-equation turbulence models to these effects and is appliedherein.

A major difficulty in modeling the fluid flow inside a mechani-cal seal is the presence of rotating and stationary boundaries

within the same fluid domain. Such a configuration requires theuse of a fully transient CFD code with a deformable mesh, whichis computationally very expensive. Another approach is to use theMRF technique proposed by Luo et al. [8]. This approach was

used by the authors for the prediction of the flow field induced byimpellers in mixing vessels using steady-state CFD models. Themodel uses two frames of reference. The first frame of referencerotates at the impeller speed and is used to compute the flowwithin the impeller in steady-state mode; the second is stationaryand is used to compute the flow away from the impeller. At theinterface between the two computational regions, the two flowfields are matched in an implicit manner to conserve mass, mo-mentum and energy across the interface. Luo et al. [8] concludedthat this approach is able to predict the details of the flow field

both within the impeller region and that outside it, with reasonableaccuracy. Compared to the full unsteady calculation with adynamic mesh, the computing time is nearly an order of magni-tude smaller. In a more recent study, Deglon and Meyer [ 9] dem-onstrated that the MRF technique and the standard k-e turbulencemodel could be efficiently used in CFD simulations of stirredtanks. The MRF technique was also successfully applied in CFDsimulations of regenerative and centrifugal pumps [10,11]. Thelocation of the interface between the two frames of reference isimportant in order to minimize numerical errors associated withapproximations at the interface. This topic was addressed byZadravec et al. [12], who found that as the size of the rotatingfluid domain increases, the influence of numerical errors due tointerpolation inaccuracies is reduced, hence resulting in moreaccurate results. A similar approach is adopted in the presentstudy, where the size of the stationary domain is limited to only a

small region near the stationary outer gland, while the rest of theflow field is within the rotating domain.

Numerical simulations are carried out in the present study usingthe commercial software ANSYS FLUENT [13], which employs a

Fig. 7 Comparison between the present experimental resultsand the affinity laws for SPW with a TO port

Fig. 8 Effect of outlet port orientation on the performance of the integral pumping device. (a )SPW, (b ) MPW, (c ) PS, and (d ) PS, API.




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finite volume discretization of the governing equations. Convec-tive terms are discretized using a second-order upwind scheme[13]. The SIMPLE algorithm [14] is used for pressure–velocitycoupling. The near-wall region is modeled using wall functions.

In the present study, the scalable wall function [15] is used toallow for systematic grid refinement without solution deteriora-tion. The numerical solution is deemed fully converged when theresidual for the discretized governing equations is reduced bythree orders of magnitude.

3.2 Boundary Conditions and Grid Generation. Theboundary condition at the seal inlet port is defined as a constantmass flow rate, while a specified gage pressure is defined at theseal outlet port. The values for the barrier fluid mass flow rate and

the outlet gage pressure are obtained from the correspondingexperimental measurements.Unstructured grid generation using tetrahedral elements is

employed in the present study. Three grids, coarse, medium, andfine, are generated for each pumping ring. The coarse grid consists

Fig. 9 Effect of barrier fluid accumulator pressure on the performance of the integral pump-ing device. (a ) MPW, RI TO and (b ) PS, TI TO.

Fig. 10 Effect of barrier fluid inlet temperature on the performance of the integral pumping device. (a ) SPW, RI RO, (b ) MPW,RI RO, (c ) PS, TI RO, and (d ) PS, API, TI RO.




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of 4 106 elements, the medium grid consists of 6 106 elementsand the fine grid consists of 8 106 elements. Figure 4 showsdifferent views of the medium grid for the MPW design with aTO port.

4 Results and Discussion

In the present section, results of the experimental study are pre-sented and discussed to evaluate the performance of the radial andaxial pumping rings under different operating conditions. Numeri-cal simulations are presented next and validated against the exper-imental results. Moreover, numerical simulations are used tovisualize the flow in order to provide a better understanding of theflow dynamics within the dual cavity seal.

4.1 Experimental Results. Figure 5 provides a comparisonbetween the different pumping ring designs. The experimentaldata is fitted to the solid lines using a second-order polynomialleast-squares regression procedure [16]. In Fig. 5, the pressure-flow rate characteristics for each pumping ring are plotted at thesame barrier fluid accumulator pressure and inlet temperature.Four comparisons are provided at high and low rotational speedsand for radial and tangential outlet port configurations. As can beseen from Fig. 5, the single PS with tight clearance (noncomply-ing with the API 682 standard) and the MPW design provide thebest performance both at high and low speeds, and for radial andtangential outlet port configurations. On the other hand, a signifi-cant deterioration in performance is evident when the API 682standard [3] is employed with the single PS. This deterioration inperformance is mainly due to the increased radial clearancebetween the rotating ring and the stationary gland which wouldlead to higher internal leakage. Figure 5 also shows that the MPW

Fig. 11 Grid independence study for MPW simulations

Fig. 12 Numerical simulations versus experimental results for the MPW. (a ) MPW, RI RO, (b ) MPW, RI TO, (c ) MPW, RI RO, and(d ) MPW, RI TO.




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provides improved performance compared to the SPW. This ismainly due to the modified slot design, which reduces internalleakage as compared to the standard slot design.

Figure 6 shows the performance of the integral pumping devi-ces with a TO port at different rotational speeds. The increase inrotational speed has a significant impact on the pumping ring per-formance, which results in higher barrier fluid flow rates and anincrease in the pressure rise. Similar behavior is also observedwith the RO port configuration. Moreover, Fig. 7 demonstratesthat the effect of rotational speed could be predicted to a reasona-

ble degree of accuracy by the turbomachinery affinity laws

Q2

Q1¼

N 2

N 1(3)

D P2

D P1¼

N 2

N 1

2(4)

The effect of the orientation of the outlet port is illustrated inFig. 8. The TO port results in an improved performance for allpumping ring designs considered in the present study, the onlyexception being the SPW in which the RO port performsrelatively better.

The effect of the barrier fluid accumulator pressure is studiednext. As can be seen from Fig. 9, the effect of the barrier fluidpressure is insignificant on the performance of both the MPW andthe PS. Similar results are also obtained for the SPW and for thePS complying with the API 682 standard [3]. Moreover, Fig. 10illustrates the effect of the barrier fluid inlet temperature on theperformance of the different pump ring designs. As can be seenfrom Fig. 10, the effect of the barrier fluid temperature is rather in-significant. Moreover, the slight reduction in the flow rate as thebarrier fluid temperature increases could be mainly attributed tothe increase in internal leakage within the seal chamber as the vis-cosity of the barrier fluid decreases. This effect is relatively mag-nified in the case of the PS complying with the API 682 standard,due to the increased clearance between the PS and the outer gland,which leads to higher internal leakage within the seal.

4.2 Numerical Simulations. Numerical simulations are car-ried out in the present section for the MPW design at differentrotational speeds and with radial and tangential outlet port config-urations. Figure 11 presents a comparison of the simulation resultsusing the coarse, medium, and fine grids. The comparison demon-strates that the fine grid provides satisfactory resolution for thepresent numerical simulations. Therefore, and for the remainder of the study, all reported results are based on the fine gridsimulations.

Figure 12 provides a comparison between the experimentalresults and the corresponding numerical simulations. Figure 12shows that CFD simulations with the standard k-e model and theMRF technique are capable of qualitatively capturing the perform-ance of the MPW at different rotational speeds and with differentoutlet port orientations. Quantitatively, Table 1 provides the maxi-mum relative error in the numerically computed pressure differ-ence for the four simulated configurations. The results reported in

Table 1 show that the present CFD model provides best results for configurations with moderate rotational speeds and a streamlined(i.e., tangential) outlet port. This is mainly attributed to the

limitations of the k-e turbulence model and the MRF technique.Higher rotational speeds would result in larger interpolation errorsat the interface between the rotating and stationary frames, while

a RO port would result in significant flow separation, which wouldnot be adequately captured by the standard k-e model.

To provide a better understanding of the improved performanceof the MPW with the TO port configuration, numerical flow visu-alization in the outlet pipe is carried out next. Figure 13 shows theswirling strength isosurface through the outlet pipe for the MPWdesign. As reported by several researchers [17,18], the swirlingstrength provides a reliable means of identifying and visualizingvortices and swirling regions in a flow field. Figure 13 demon-strates that the RO port forces the flow exiting the pumping ringto significantly change direction, resulting in the development of alarge swirling region that would decay over a long distance of theoutlet pipe. On the other hand, the TO port is more streamlinedwith the flow exiting the pumping ring, resulting in a significantlysmaller swirling region and hence explaining the improved

performance of the MPW with the TO port configuration.Moreover, to explain the improved performance of the MPWover the standard design, Fig. 14 shows the swirling strength iso-surface on the downstream side of the paddle wheel. Figure 14shows that the SPW is characterized by larger swirling regions onthe downstream side of the wheel, which indicate higher internalleakage as compared to the MPW design. The higher internal

Table 1 Maximum relative error in the numerically computedpressure difference for the MPW design

Revolutions per minute (rpm) Outlet port |Maximum relative error| (%)

1500 TO 91500 RO 153600 TO 143600 RO 20

Fig. 13 Swirling strength isosurface (value550 s21) throughthe outlet pipe for the MPW. (a ) MPW, RO and (b ) MPW, TO.

Fig. 14 Swirling strength isosurface (value5250s21) on thedownstream side of the standard and MPWs. (a ) SPW, RI TO(3600 rpm) and (b ) MPW, RI TO (3600 rpm).




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leakage and larger swirling regions would now clearly explain theimproved performance of the MPW compared to the standarddesign.

5 Conclusion

The present study numerically investigated the performance of four pumping ring designs for dual mechanical seals. Experimen-tal results reveal that the performance of the pumping rings is sig-nificantly affected by the pump rotational speed and theorientation of the outlet port. The study also shows that the per-

formance of axial pumping rings is significantly improved byreducing the radial clearance between the rotating ring and the sta-tionary outer gland. Moreover, numerical simulations show goodagreement with experimental results for pumping ring configura-tions with streamlined TO ports and at moderate rotational speeds.

Acknowledgment

The authors would like to acknowledge the financial supportfrom John Crane UK in providing the necessary equipment andinstrumentation for the present study. In particular, the authorswould like to thank Dr. Amrat Parmar, Dipl.-Ing. Klaus Meck,and Dr. Greg Zhu from John Crane UK for many fruitfuldiscussions regarding the present work.

NomenclatureMPW ¼ modified paddle wheel

PS ¼ pumping scrollRI ¼ radial inlet port

RO ¼ radial outlet portSPW ¼ standard paddle wheel

TI ¼ tangential inlet portTO ¼ tangential outlet port

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