Final Project Report CPC

8/11/2019 Final Project Report CPC

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De si gn , anal ysi s and manuf acturing of Cu rrent to Pressure Convert er for

Steam Turbine

A PROJECT REPORT

SUBMITTED BY Darshik Sheth (100410119031)

Meet Patel (100410119010)

Shravan Ranade (100410119018)

Vaishal Shah (100410119035)

In partial fulfillment for the award of the degree

Of

BACHELOR OF ENGINEERING

In

MECHANICAL ENGINEERING DEPARTMENT

Of

SARDAR VALLABHBHAI PATEL INSTITUTE OF TECHNOLOGY,

VASAD

GUJARAT TECHNOLOGICAL UNIVERSITY,

DECEMBER, 2013


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ACKNOWLEDGEMENTS

This project would not have been complete without the support of many people. We

would like to give a special thanks to Mrs. Hetal Chauhan for her assistance and

guidance throughout this project work. We would also like to thank Mr. Hitesh Vakil,

Mr. Pankaj Shah and Mr. Vipul Shah from SIEMENS Ltd. for their expert guidance,

comments and suggestions. Without the help of their knowledge and expertise in every

facet of the study, from helping to find the relevant formulas and to analyzing the results,

this project could not have been completed. We would finally like to give special thanks

to all our family and friends who have given us support throughout the past 6 months,

because without them none of this would have been possible.


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ABSTRACT

As we know, the world today is unimaginable without electricity. Large scale electricalenergy production largely depends on the use of turbines. Nearly all of the world's powerthat is supplied to a major grid is produced by turbines. A turbine is a simple device withfew parts that uses flowing fluids (liquids or gases) to produce electrical energy. Fluid isforced across blades mounted on a shaft, which causes the shaft to turn. The energy

produced from the shaft rotation is collected by a generator which converts the motion toelectrical energy using a magnetic field.

The controlling of the turbines was done initially by mechanical governors. Now, with thetechnological advancement, they are replaced by the electronic governors.

Steam turbine governing is the procedure of controlling the flow rate of steam into asteam turbine so as to maintain its speed of rotation as constant

This project is dedicated to the study and design of a Current to Pressure Converter,which converts the electric signal from the governor to the hydraulic signal, which furthercontrols the servomechanism.

It forms an integral part of the controlling. CPC quickly and accurately converts the inputsignal into a proportional output pressure. This is important as the turbine should functionas expected in load variant conditions and also stop in abnormal conditions.


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SARDAR VALLABHBHAI PATEL INSTITUTE OF TECHNOLOGY, VASAD

MECHANICAL ENGINEERING DEPARTMENT

CERTIFICATE

This is to certify that the dissertation entitled De sign, analysi s an d manuf acturing of

Current to Pressure Converter for Steam Turbine has been carried out by Darshik

Sheth, Meet Patel, Shravan Ranade, Vaishal Shah under my guidance in partial

fulfillment of the degree of Bachelor of Engineering in Mechanical Engineering (VII th

Semester) of Gujarat Technological University during the Academic Year 2013 14.

Internal Guide: External Guide:

Prof. Hetal R. Chauhan, Mr. Pankaj F. ShahAssistant Professor, Designing Head, Core Engineering,Mechanical Engineering Department Siemens Ltd.

External Examiner: Dr. P. V. Ramana

Head, Mechanical Engg. Department


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3.6 Solenoid Valve (US5469886A)..35

3.7Direct Solenoid operated directional control valve (US4338966A)..38

Chapter 4: Design Phase

4.1 Introduction ....44

4.2 Design of Orifice.....45

4.3 Bore Diameter.46

4.4 Oil Seal Selection47

Chapter 5: Conclusion and Future plans

5.1 Conclusion...54

5.2 Future Plans.54

References.56


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LIST OF TABLES

Table No. Table Description Page No.

2.1 Comparison of Throttle & Nozzle control 19

4.1 Temperature Range for common elastomers 49

4.2 Comparison of commonly used elastomers 50


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LIST OF SYMBOLS

SYMBOL ABBREVIATIONS

I Current

SG Specific Gravity

Q Flow rate

Cv Flow coefficient

N Turbine Rpm

P Pressure

oC Degree Centigrade

oF Degree Fahrenheit

F Frequency

p Pole

D Diameter

L Length

mA Milli-Ampere

E,V Voltage

R Resistance

r Radius

A C/S area


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LIST OF FIGURES

Figure No. Figure Description Page No.

2.1 Steam Turbine Speed Control System 18

2.2 Nozzle governing 19

2.3 Working principle of CPC 20

2.4 Logic of CPC 21

3.1 Pressure difference v/s Voltage analogy 23

3.2 Spool (Source: Parker O-Ring Manual) 25

3.3 Static seals (Source: Parker O-Ring Manual) 26

3.4 Dynamic seals (Source: Parker O-Ring Manual) 27

3.5 Rotary seals (Source: Parker O-Ring Manual) 27

3.6 Seat seals (Source: Parker O-Ring Manual) 28

3.7 O-ring (Source: Parker O-Ring Manual) 29

3.8 O-ring Operation (Source: Parker O-Ring Manual) 31

3.9O-ring under pressure (Source: Parker O-Ring

Manual)31

3.10 O-ring Extruding (Source: Parker O-Ring Manual) 31

3.11 O-ring Failure (Source: Parker O-Ring Manual) 31


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CHAPTER 1

Introduction


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1.1 Company Background

R. S. No. 144-A,

Opp. Makarpura Railway Station,

Maneja, Vadodara

Industry Coordinator : Mr. Pankaj Shah

The Siemens Group in India has emerged as a leading inventor, innovator and

implementer of leading-edge technology enabled solutions operating in the core business

segments of Industry, Energy, Healthcare and Infrastructure and Cities. The Groups

business is represented by various companies that span across these various segments.

Siemens brings to India state-of-the-art technology that adds value to customers through a

combination of multiple high-end technologies for complete solutions. The Group has the

competence and capability to integrate all products, systems and services. It caters to

Industry needs across market segments by undertaking complete projects such as

Hospitals, Airports and Industrial units.

The Siemens Group in India comprises of 17 companies, providing direct employment to

over 18,000 persons. Currently, the group has 21 manufacturing plants, a wide network of

Sales and Service offices across the country as well as over 500 channel partners.

Today, Siemens, with its world-class solutions plays a key role in Indias quest for

developing modern infrastructure.

Siemens unit at Maneja, Vadodara comes under Energy Sector (Oil and gas Division) and

has main products including Steam Turbine systems.


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1.2 Company Products

Industry

Industrial Automation

Drive TechnologyBurning Technology

Osram

Industrial Solution

Mobility

Energy

Oil & Gas

Fossil Power Generation

Renewable Energy DivisionPower Transmission

Power Distribution

Healthcare

Workflow & Solution

Diagnosis

1.3 Objective of Project

To facilitate the localization of the product and make the product in-house

1.4 Topic Overview

This Project Report is based on developing the current to pressure converter, which is

basically a solenoid valve. It is an important component in the controlling of the steam in

case of the turbines.

The aim of this project is to develop this product to facilitate the in-house production of

this product.


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1.5 Project Methodology

Understanding the basic concept

Literature study

Recognition of new system

Design of new mechanism

Preparation of model

Conducting of tests

Comparison with existing systems


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CHAPTER 2

Brief Introduction to need of CPC


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2.1 Why Fuel control is required in turbines?

For getting electricity at constant frequency turbine should rotate at constant speed & the

relation between frequency & RPM in generator is given by:

N= (120f/P)

Where

N = Rpm of turbine

f = Frequency

P = poles

so for F= 50Hz, P=4, N=1500 RPM

So at different Electrical load, if speed will remain same then we will get constant

frequency.

Initially when steam/gas turbines were developed, mechanical governors were used. But

due to less efficiency, wear and friction problems now it is replaced by electronic

governors.


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2.2 Speed Control System in Steam Turbine

Steam turbine governing is the procedure of controlling the flow rate of steam into

a steam turbine so as to maintain its speed of rotation as constant. The variation in load

during the operation of a steam turbine can have a significant impact on its performance.

In a practical situation the load frequently varies from the designed or economic load and

thus there always exists a considerable deviation from the desired performance of the

turbine. The primary objective in the steam turbine operation is to maintain a constant

speed of rotation irrespective of the varying load. This can be achieved by means

of governing in a steam turbine.

Principal types of Steam turbine governing are:

Throttle Governing Nozzle Governing

Bypass Governing

Combination Governing

Emergency Governing

In nozzle governing the flow rate of steam is regulated by opening and shutting of sets of

nozzles rather than regulating its pressure. In this method groups of two, three or more

nozzles form a set and each set is controlled by a separate valve. The actuation of

individual valve closes the corresponding set of nozzle thereby controlling the flow rate.

These individual valves are controlled by servo mechanism, actuated by the CPC, based

on the governor signal.


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Fig 2.1 Steam Turbine Speed Control System

DETAILS ON NOZZLE GOVERNING (Mechanism and working)

The efficiency of steam turbines is considerably reduced if throttle governing is carried

out at low loads. An alternative and more efficient of governing is by means of nozzle

control.

Figure shows a diagrammatic arrangement of nozzle governing system in which 3 to 5 or

more nozzles are worked together and supply steam to turbines. At full load nozzles are

fully opened and at less loads nozzle openings are controlled by valve chest which is

controlled by servo mechanism.


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Fig 2.2 Nozzle Governing

Table 2.1 Comparison of Throttle & Nozzle control

Sr No. Aspects Throttle control Nozzle control

1 Throttling losses Severe No throttling losses

2 Partial admission losses Low High

3 Heat drop available Lesser Larger

4 Suitability Small turbines Medium and larger turbines


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2.3 Working of CPC

A set signal of 4-20 mA is generated by the electronic governor. This signal is based on

the load on the turbine/governor. The signal generates a magnetic force in the solenoid,

the limits of which can be adjusted by means of the potentiometers and which is then

transmitted onto the control piston via tappet.

The hydraulic force is generated due to the pressure of incoming hydraulic oil at 9 bar (g)

pressure. The hydraulic force being proportional to the output signal pressure acts against

this force.

In the case of the two forces being equal, the control piston is positioned in the hydraulic

center. In this situation, the output signal pressure corresponds to the set signal. In the

hydraulic center position the control piston performs minimum oscillating movements

in the area of the guiding edges P PA and P A T, in order to keep the output pressure P A

on the value set by magnetic force.

Fig 2.3


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Fig. 2.4 Logic of CPC

When increasing the set signal and thus magnetic force from this condition, the control

piston position changes and thus connects the output pressure P A to the feed pressure P

and blocks P A towards the tank return line. Now pressure P A will increase until the same

has returned the control piston to the "hydraulic center" and P A corresponds to the new set

signal. The spring force of the control spring generates a force-offset in order to guarantee

the I/H converter function for output pressures of approx. 0 bar, too. The internal leakage

is fed back into tank return line.


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Chapter 3

Literature Review


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4.1 Orifice

Fundamentals

Pressure difference and volta

So,

I (Current) Q (Fl

V (Voltage) P (Pr

In electrical circuit we used t

But in Hydraulic circuit findi

the reciprocal of the resistan

Pressure = Flow*Res

Pressure = Flow*(1/

ge difference analogy

Fig 4.1

w) &

ssure)

o find resistance.

ng out resistance is difficult so we find conduct

e. Denoted by C Or K

istance

onductance

23

ance which is


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And this C value is known as Flow coefficient.

4.1 Flow coefficient

The flow coefficient of a device is a relative measure of its efficiency at

allowing fluid flow.

It describes the relationship between the pressure drop across an orifice, valve or other

assembly and the corresponding flow rate.

Mathematically the flow coefficient can be expressed as:

C V = ??

?

where:

Cv = Flow coefficient or flow capacity rating of valve.

F = Q=Rate of flow (US gallons per minute).

SG = Specific gravity of fluid (Water = 1).

P = Pressure drop across valve (psi).

In more practical terms, the flow coefficient Cv is the volume (in US gallons) of water at

60F that will flow per minute through a valve with a pressure drop of 1 psi across the

valve.


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The use of the flow coefficient offers a standard method of comparing valve capacities

and sizing valves for specific applications that is widely accepted by industry.

In our project after finding out Cv value we are able to find out minimum orifice sizethat is required to meet the specifications.

4.2 Spool

Spool is of two types namely sliding and rotary. Sliding spool is cylindrical in cross

section, and the lands and grooves are also cylindrical. Rotary valves have sphere-like

lands and grooves in the form of holes drilled through the spheres.

The type of spool used in our project is of sliding type.

Fig 3.2 Spool


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4.3 Seals

Types of Seals

1. Static Seals

2. Dynamic Seals

3. Rotary Seals

4. Seat Seals

1.Static Seals

They are used when no relative movement occurs between the mating parts.The seal is

usually compressed between two adjacent parts securing the two stationary parts together by fasteners.

Example : Static seal may be used between the pump housing and the end plate.

Fig 3.3 Static Seals


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2.Dynamic Seals

They are used between the surfaces of hydraulic parts where movement occurs and

controls both leakage and lubrication. In certain dynamic seals wipers or covering boots

are used to keep away dirt and foreign materials.

Example : Use of road wiper.

Fig 3.4 Dynamic seals

3.Rotary Seals

In a rotary seal, either the inner or outer member of the sealing elements turn (around the

shaft axis) in one direction only. This applies when rotation is reversible, but does not

allow for starting and stopping after brief arcs of motion, which is classed as an

oscillating seal. Examples of a rotary seal include sealing a motor or engine shaft, or a

wheel on a fixed axle. See Figure 4.5.

Fig 3.5 Rotary Seals


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4. Seat Seals

In a seat seal, the O-ring serves to close a flow passage as one of the contact members.

The motion of closing the passage distorts the O-ring mechanically to create the seal, in

contrast to conditions of sealing in previously de ned types. A sub-classi cation is

closure with impact as compared with non-impact closure. Examples of a seat-seal

include O-ring as a washer on the face of a spiral threaded valve, a seal on the cone of a

oating check valve, and a seal on the end of a solenoid plunger. See Figure 4.6

Fig 3.6 Seat Seals

We have selected O-ring type of seal according to shape.

4.4 O-ring

What is an O-Ring?

An O-ring is a torus, or doughnut-shaped ring, generally molded from an elastomer,

although O-rings are also made from PTFE and other thermoplastic materials, as well asmetals, both hollow and solid.

O-rings are used primarily for sealing. O-rings are also used as light-duty, mechanical

drive belts.


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What is an O-Ring Seal?

An O-ring seal is used to prevent the loss of a uid or gas. The seal assembly consists of

an elastomer O-ring and a gland. An O-ring is a circular cross-section ring molded fromrubber . The gland usually cut into metal or another rigid material contains and supports

the O-ring . The combination of these two elements; O-ring and gland constitute the

classic O-ring seal assembly.

Fig 3.7 O-Ring

4.4.1 Advantages of O-Rings

They seal over a wide range of pressure, temperature and tolerance.

Ease of service, no smearing or retightening.

No critical torque on tightening, therefore unlikely to cause structural damage.

O-rings normally require very little room and are light in weight.

In many cases an O-ring can be reused, an advantage over non-elastic at seals and

crush-type gaskets.

The duration of life in the correct application corresponds to the normal aging period of

the O-ring material.


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O-ring failure is normally gradual and easily identi ed.

Where differing amounts of compression effect the seal function (as with at gaskets), an

O-ring is not effected because metal to metal contact is generally allowed for.

They are cost-effective.

4.4.2 O-ring Operation

Operation All robust seals are characterized by the absence of any pathway by which

uid or gas might escape. Detail differences exist in the manner by which zero clearance is

obtained welding, brazing, soldering, ground ts or lapped nishes or the yielding of a

softer material wholly or partially con ned between two harder and stiffer members ofthe assembly. The O-ring seal falls in the latter class.

The rubber seal should be considered as essentially an incompressible, viscous uid

having a very high surface tension. Whether by mechanical pressure from the surrounding

structure or by pressure transmitted through hydraulic uid, this extremely viscous uid is

forced to ow within the gland to produce zero clearance or block to the ow of the

less viscous uid being sealed. The rubber absorbs the stack-up of tolerances of the unitand its internal memory maintains the sealed condition. Figure 4.8 illustrates the O-ring

as installed, before the application of pressure. Note that the O-ring is mechanically

squeezed out of round between the outer and inner members to close the uid passage.

The seal material under mechanical pressure extrudes into the micro ne grooves of the

gland. Figure 4.9 illustrates the application of uid pressure on the O-ring. Note that the

O-ring has been forced to ow up to, but not into, the narrow gap between the mating

surfaces and in so doing, has gained greater area and force of sealing contact. Figure 4.10

shows the O-ring at its pressure limit with a small portion of the seal material entering the

narrow gap between inner and outer members of the gland. Figure 4.11 illustrates the

result of further increasing pressure and the resulting extrusion failure. The surface

tension of the elastomer is no longer suf cient to resist ow and the material extrudes

( ows) into the open passage or clearance gap.


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Fig 3.8 O-ring installed

Fig 3.9 O-ring under pressure

Fig 3.10 O-ring Extruding

Fig 3.11 O-ring failure


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4.5 Solenoid controlled

A pilot operated flow regulaenergization to a solenoid.

pressure-differential-actuate

valve employing a solenoidcontrol the flow of fluid thmain valve element.

Figures

Description

The present invention proviart, provides a flow valve wsupplied to a solenoid coil.

flow valve (US RE32644 E)

ting valve which can be remotely controlled bThe valve consists of a normally-closed,

, main valve element and a pressure-compensa

acting like a remotely controllable electromagnough the pilot valve and the pressure differen

.12 Solenoid controlled flow valve

es a flow valve which overcomes the problemich may be easily regulated by the amount of e

32

y varying the pring-biased,ed, pilot flow

etic spring toce across the

s of the priorectric current


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In accordance with the invention, a remotely controllable fluid flow valve is provided,comprised of a three-ported, main valve section including a mechanical spring biased,

pressure-differential-actuated, main valve element and a two ported, pressure-compensated, pilot-type flow valve including a solenoid acting as a variableelectromagnetic spring to variably control the volume of fluid flowing through the pilotvalve and the pressure difference across the main valve element.

More specifically in accordance with the invention, a flow valve is provided comprised ofa housing having an elongated cylindrical cavity with a pressure inlet and a first outlet

port, a first valve element operable to control the flow of fluid between the ports andhaving a first surface exposed to inlet pressure and a second oppositely facing surfaceexposed to pressures in a variable pressure chamber, mechanical spring means lightly

biasing the element to an initial position relative to the ports, fixed orifice meanscommunicating the inlet port with the chamber, a second outlet port communicating thechamber to low pressure, a second valve element operable to control the flow of fluid

through the second port and an electro-magnetic spring for controlling the position of thissecond valve element relative to the second port comprised of a solenoid including asolenoid coil, a sleeve-like, armature-attracting pole piece and an axially spaced, sleeve-like, armature-supporting pole piece with the armature so positioned relative to the pole

pieces that as the armature moves into the attracting pole piece, the magnetic forcedecreases. To provide this, the armature must substantially overlap the supporting pole

piece and slightly overlap the attracting pole piece.

Further in accordance with the invention, the valve is comprised of: a housing having an

elongated cylindrical cavity with an inlet port and a main outlet port; a valve element inthe form of a piston slidable in the cavity to restrict the outlet port and having one endsurface exposed to the inlet pressure and an opposite surface defining with the cavity avariable pressure chamber; mechanical spring means biasing the element to the valveclosed position; a fixed orifice through the piston communicating inlet pressure with thevariable pressure chamber; an outlet port communicating the variable pressure chamber tolow pressure and a pilot valve element movable to restrict this latter port; a magnetically

permeable armature operatively associated with the pilot valve element; a solenoid coiland magnetic field poles arranged to exert an axial force on the armature; the armatureand the field pole being so arranged that as the armature moves further into one of thefield poles, the magnetic force decreases; and, an axially facing orifice operativelyassociated with the pilot valve element for exerting a flow force thereon and provide

pressure compensated flow through the pilot valve.

In essence, when the solenoid is energized to a given level, the valve element is moved toeffect communication of the chamber with the pilot valve outlet port and permit flow offluid through the main valve orifice and the pilot valve orifice in a volume which is

pressure compensated, that is, it is constant above a certain minimum pressure regardless


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of inlet system pressure above that minimum pressure. This pilot flow creates a pressuredifferential across the main valve element which then is moved to the open position bythe pressure differential across the valve element creating a pressure force in oppositionto the mechanical spring bias. This permits flow of fluid from the inlet port through themain outlet port in an amount proportional to the pressure differential which will remainconstant for all values of inlet pressure above a minimum.

Summary

This invention pertains to the art of fluid valves, and more particularly to an electricallycontrolled flow valve.

The invention is particularly applicable to hydraulic flow valves operating at pressures upto 6,000 pounds per square inch and will be described with particular reference thereto,although it will be appreciated that the invention has broader applications and may beused in many types of valves either for controlling the flow of liquids or gases.

Flow valves are used extensively in industry to control the volume of fluid flowing from afixed volume hydraulic pump to a motor or other apparatus, the speed of which must becontrolled.

Such valves in the past have usually included an adjustable variable orifice, the openingof which is controlled by a threaded element which may be manually adjusted or driven

by a remotely controlled electric motor.


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4.6 Solenoid valve (US

In a solenoid valve comprisi

multifunctional solenoid val

of the first solenoid valve is

immediately opened, thereby

solenoid valve comprises a fi

which is movable by the sole

disposed to be brought into c

provided for urging the diap

moved away from the diaphr

said solenoid valve for contr

valve.

Figures 3.13 Solenoid

469886 A)

g a first solenoid valve for regulating pressure

e integrated with the first solenoid valve, even

nergized, a diaphragm valve of the solenoid va

providing a safe & compact solenoid valve. Th

rst spring provided between the upper end of a

noid and stator, while the lower end of the plun

ontact with the diaphragm valve and a second s

ragm valve in the direction to close it even if th

agm valve and the second solenoid valve, integ

lling to selectively open/close a closing valve a

alve

35

nd a second

f a solenoid

lve is not

e first

lunger

ger is

ring is

e plunger is

ated with

nd a one way


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Problems with Prior art

(a) Since the plunger is directly connected to the diaphragm valve, the diaphragm valve isimmediately opened when the solenoid is energized. Accordingly, when an atmosphereopening passage of the canister which is open to the atmosphere is clogged or resistancethere through is large, the negative pressure in the intake manifold is introduced into thesecond connecting passage, then passes through the diaphragm valve and the firstconnecting passage, whereby the fuel tank is negatively pressurized to apply stress in thefuel tank. As a result, there is a likelihood of breakage of the fuel tank.

(b) Since the stopper formed of the rubber or the nonmagnetic material is attached to thestator by way of sticking of the stopper to the stator, if the stopper is detached from thestator, the plunger is prevented from being slid. As a result, there is likelihood that thesolenoid valve is prevented from working normally.

(c) Since the pressure regulating solenoid valve and the atmosphere introduction valve areseparately provided, both valves need individual spaces for attachment thereof.

(d) Since the evaporated fuel generated in the fuel tank is condensed in the first andsecond connecting passages and liquefied therein, there is a likelihood that the valve

performance is changed or the evaporated fuel leaks outside, which causes a fire.

Summary

The present invention has been made in view of the problems of the prior art solenoid

valve. It is an object of the present invention to provide the solenoid valve which is safe

and compact since a diaphragm valve is not immediately opened even if the pressure

regulating solenoid valve is energized.

It is another object of the invention to provide a solenoid valve comprising two solenoidvalves which are integrated with each other to thereby make the solenoid valve compact.

To achieve the above objects, the solenoid valve according to a first aspect of theinvention comprises a first connecting passage connected to a fuel tank side, a second


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connecting passage connected to an intake manifold side, a diaphragm valve for permitting both the first and second connecting passages to communicate with each other,a solenoid body having a solenoid, a plunger and a stator wherein the plunger verticallymovable by the solenoid and the stator is disposed above the plunger, a first springdisposed between the upper end of the plunger and the stator wherein the lower end of the

plunger is disposed to be able to contact with the diaphragm valve, and a second springdisposed between the diaphragm valve and the solenoid body for urging the diaphragmvalve in the direction to close it in the state where the plunger is moved away from thediaphragm valve.

A solenoid valve according to a second aspect of the invention is characterized in thatthere are provided in the first aspect of the invention a hole or groove defined vertically inthe stator and a stopper formed by mold of the nonmagnetic material provided in the holeor groove for stopping the plunger.

A solenoid valve according to a third aspect of the invention is characterized in that thereare provided in the first aspect of the invention a solenoid valve for adjusting pressure

between the first and second connecting passages, a bypass passage between the first andsecond connecting passages, a one way directional valve (hereinafter referred to as oneway valve) between the bypass passage and a third connecting passage, a closing valve

between the bypass passage and the second connecting passage and a second solenoidvalve which is integrated with the first solenoid valve for controlling to selectively openor close the closing valve and the one way valve.

A solenoid valve according to a fourth aspect of the invention in characterized in that thethird connecting passage serves as an atmosphere passage, the one way valve serves as anatmosphere introduction valve, and the closing valve serves as a negative pressureintroduction valve.

A solenoid valve according to a fifth aspect of the invention is characterized in that thereis provided an evaporation device for evaporating liquefied material due to capillarity inthe second solenoid valve, the first connecting passage or the second connecting passagein the third aspect of the invention


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4.7 Direct solenoid ope

A spherical valve memberdirect application of supplymechanical force imposedeliminating the use of movsticking in the presence of so

Figure 3.14 Direc

ated directional control valve (US 433

s moved within a housing between operative pressure and operation of a solenoid arranged

n the valve member while minimizing sole ble close fitting cylindrical parts which canlid contaminants.

solenoid operated directional control

38

8966 A)

positions byto maximizeoid size and

be subject to

valve


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Description

Referring first to FIG. 1, the invention directional control valve is illustrated in one preferred embodiment configured for installation in a fluid submerged environment suchas the valve housing of an automotive automatic transmission as including generally a

valve housing, a valve assembly carried in the valve housing, and an actuating solenoidassembly carried with the housing and operatively engageable with the valve assembly.

The valve housing is illustrated as including a ported body to which is secured by screwsor the like a cover plate. An elongated valve chamber is formed in the body by a stepped

bore having a reduced diameter inner portion and an enlarged diameter outer portion. Afirst passage communicates with the valve chamber portion and is connectable by meanswell known in the hydraulic arts with a source of pressurized fluid such as a transmission

pump (not shown). A second passage communicates with the valve chamber portion andis connectable by similar means with a fluid responsive device such as a fluid actuatingcylinder of an automotive automatic transmission (not shown).

Turning next to the valve assembly, it is illustrated as comprising a seat member, aretainer member, a guide member, a ball, and an actuating plunger. The seat member is

preferably formed as a disc having a central bore and received in closely fitting diametralrelationship in the enlarged diameter chamber portion, abutting the terminal shoulder 31thereof.

The retainer member includes a large diameter portion which is likewise received inclosely fitting diametral relationship in the enlarged diameter chamber portion 30 andabuts the seat member to effect axial retention. The retainer member also includes areduced diameter neck portion extending outward (rightward as viewed in FIG. 1) fromthe large diameter portion, and has formed through it a central stepped bore having anenlarged diameter valve portion traversed perpendicularly by a cross port and a reduceddiameter connecting portion. The inner terminus of portion and the outer terminus of seatthrough bore together form a pair of spaced, axially aligned valve seats for a purpose to

be hereafter described.

The guide member is formed as a substantially cylindrical member having an elongatedsolenoid mounting portion and an inner terminal radially extending flange portion. Itfurther includes a central through bore having an inner portion sized to receive the neck


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portion of retainer member in close fitting diametral relationship, a smaller plunger access portion, and a retaining shoulder there between. A cross port extends through the bore at a position adjacent the outer terminus of the neck portion of retainer member.

The ball member is loosely carried within the enlarged diameter valve portion of theretainer member and is movable to selectively engage the valve seats in a manner to behereafter described.

The plunger member is carried for axial movement within the guide member and isformed as an elongated rod having an actuating stem portion, a guide portion, anelongated rod portion, and a solenoid connecting portion.

The stem portion is preferably cylindrical and passes along the axis of the connecting portion of retainer stepped bore to abuttingly engage the ball. Its outer terminus preferably includes a blend radius to the larger diameter of the guide portion.

Guide portion is preferably, but not necessarily formed as a generally cylindrical flutedstructure (as may best be seen in FIG. 4) defining an outer diameter slidingly received inloose fitting relationship with the bore of guide member for movement between theshoulder and the outer terminus of the retainer member. The mentioned flutedconstruction is preferred, since it minimizes the surface areas in sliding contact, thusreducing the tendency to stick; but in the FIG. 1 embodiment the critical functionallimitation on the structure of the guide portion is that it permit communication betweenthe connecting bore portion of retainer member 38 and at least the cross port of guidemember when the guide portion abuts the outer terminus of the retainer member.

Rod portion is preferably cylindrical and extends outwardly from the guide portion 76 tothe solenoid connecting portion.

Solenoid connecting portion includes inner enlarged diameter portion and outer retainerreceiving groove. Inner portion includes an outward facing chamfer connecting its outerdiameter with connecting portion.


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Turning next to the solenoid assembly, it is illustrated as a clapper type solenoidincluding essentially a mounting bracket, a solenoid core assembly, and a clapper typearmature.

The mounting bracket is preferably formed as a generally L-shaped member having avertical (as shown in FIG. 1) leg configured to be conventionally secured as by threadedfasteners, as seen in FIG. 3, to the valve body. The inner surface of the leg may include arecess for receiving the flange portion of guide member, thereby effecting axial retentionof the valve assembly. It may also include an aperture for conventionally receiving andretaining the core assembly. Extending outward from the vertical leg is a generallyhorizontal leg, preferably including a portion turned vertically to support the armature

providing a large flux area therewith. Notches are formed in the sides of the leg, and theformation of the portion leaves support posts as may best be seen in FIG. 3.

The armature is secured to the mounting bracket for pivotal movement thereabout bymeans of a special retaining member formed as by stamping from a resilient material suchas spring steel and including a strap portion and a leaf spring portion. Strap portionincludes apertures. The former engage notches in snap fit relationship, and the latterloosely receive the posts to prevent rotation about the longitudinal axis of the horizontalleg.

The end of the armature remote from its point of pivotal support on the vertical portion isreduced in section as shown in FIG. 3 and includes a through aperture sized to pass theconnecting portion of the plunger. A preferably chamfered surface is formed at the innerterminus of the aperture to cooperatively abuttingly engage the chamfered surface of the

plunger to permit self-centering of the plunger thereby resisting binding. It will beappreciated by those skilled in the art that other cooperating surface configurations suchas matching spherical surfaces might be chosen for this purpose.

As may best be seen in FIG. 3, leaf spring portion is slotted at its free end as indicated atto receive connecting portion of plunger. A conventional retaining ring received in thegroove of the plunger secures the assembly axially. The leaf spring portion is formed atassembly to exert a small preload on the plunger in the rightward direction as viewed inFIG. 1. Armature is also formed by bending at assembly to provide a predeterminedspacing from the core assembly.


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Summary

Responsive to the deficiencies in the prior art, it is an object of the present invention to provide a directly solenoid operated directional control valve that is resistant to sticking.

It is another object to provide such a valve which maximizes the actuation force andtravel produced by the solenoid.

It is yet another object to provide such a valve which minimizes contact stresses inactuation.

It is still another object to provide such a valve which is simple and economical to produce.

According to one feature of the present invention, a solenoid operated directional controlvalve is provided which employs a minimal number of close fitting moving cylindrical

parts.

According to another feature, the solenoid of the valve of the present invention is coupledto an actuating plunger for the valve portion with mechanical advantage and is arrangedsuch that operation against large preload forces is not required.

According to yet another feature, contact between the valve operative structure and theactuating plunger therefore is maintained in the unactuated condition of operation tosubstantially prevent impact loading.


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Research Paper

THE CAVITATION IN ZONE BODY-SPOOL VALVEFOR HYDRAULIC DISTRIBUTOR

Abstract.

The cavitation phenomena which occurring in hydraulic machinery and hydraulicequipment, is for many years a research topic, which lead to great running improvementsof the hydraulic systems. Cavitation occurs in some elements of hydraulic drive devicesas a result of great increases of the flow velocity. In distributors it represents the majorlimits the possibility to reduce the flow capacity beyond some values. The initiation andthe development of the cavitation in hydraulic drive systems present some particularitiesoriginated both in the liquid employed (especially mineral oil) and the severe running

conditions. The cavitation erosion is generally not a very stressing factor. On the otherhand, as a result of important modifications in the characteristics of the working fluid,when the pressure decreases (increasing of the gas and steam content) the running of thesystem is heavily disturbed.

Having as a start point in general the definition of cavitation coefficients for thedirectional valves with cylindrical spools. For this definition, there were taken intoaccount all the important aspects of the flow geometry and the elements depending on thefluid nature. By attentive examinations of the flow through the specific hydraulicresistance it was established a relation to obtain the cavitation reserve of the system.

1. The cavitation phenomenon in distribution section distributor body - spool valve

The hydraulic resistance describes hydraulic system elements with diversefunctional role, and the diversity of type construction assures the purpose for their role.The majority of hydraulic resistance works by strangle the flow vein . From this reasonthe cavitation phenomena can occurs during with increasing speed and the different

pressure values in different points of the hydraulic resistance. In hydraulic distributionapparatus, the cavitation phenomenon occurs in case of command resistance with smallaperture, where the speed will be increased, and from Bernoulli equation results that the

pressure drops in the case when exists a big drop of pressure on the command resistance.The distributor hydraulic track is characterized by the distributor body-spool valvegeometry, and different type of spool valves geometry. For studying the cavitation regimein hydraulic resistance distributor body-spool valve , the research had focus to determineenergetic spectrum of analyse cavitation effect among the hydraulic distribution apparatus


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CHAPTER 4

Design Phase


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3.1 Introduction

The preliminary design for t

entities:

Orifice size

Cylinder diameter

Spool diameter

Seal selection

Gland diameter

Gland thickness

e current to pressure converter includes desig

46

of following


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Fig 4.1 Introduction to CPC

3.2 Design of orifice

The flow coefficient of a device is a relative measure of its efficiency at

allowing fluid flow. It describes the relationship between the pressure drop acrossan orifice, valve or other assembly and the corresponding flow rate.

Mathematically the flow coefficient can be expressed as:

CV = ??

?

WhereCv = Flow coefficient or flow capacity rating of valve.F = Rate of flow (US gallons per minute).SG = Specific gravity of fluid (Water = 1). P = Pressure drop across valve (psi).

Following the above equation,

for port A, C V=1.5793

for port T, C V=1.9724

Input Parameters:

Specific Gravity of Hydraulic Oil used = 0.9

For P A

Q = 24 Lpm at 1 bar pressure difference

For A T

Q = 30 Lpm at 1 bar pressure difference

*Lpm= Liter per minute of oil

1 LPM= 0.264172052 GPM (US)

1 Bar= 14.503 Psi


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For finding orifice diameter, we have used the Spirax Sarco Calculator. It gives the

experimentally determined value for orifice diameter corresponding for that value of the

flow coefficient. This is taken as a reference.

Fig 4.2 Spirax Sarco Calculator

For ports A and B, the values of diameters are,

dA = 6.37 mm

dB = 7.12 mm

3.3 Bore Diameter

Bore is the inner cylindrical portion of the valve which encloses the spool. The annular

area between the bore and spool should be equal to the orifice area.

We select the largest of the area of A and T.


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Area of A orifice = r 2 = (7.12/2) 2 = 39.79511

This area should be maintained inside cylinder including 5mm of spool.

So,

( R 2-r 2 )= 39.79 mm

Here r = 0.25mm spool radii

D= 2R= 8.7 mm

3.4 Oil Seal Selection

There are lots of physical conditions which need consideration before an engineer opts to

use seal or seal material for the specific application

The governing factors are:

1. Working Pressure and pressure range

2. Environmental condition

3. Fluid medium4. Static or dynamic application

5. Temperature of the system

6. Functional reliability and expected life

But in order to generate a failsafe sealing system, the above factors may not be enough.

One has also to look for the physical and chemical properties of the seal materials used. A

through look at the following material properties, especially in case of elastomeric seals is

therefore very essential for the designers as well as for the maintenance engineers in order

to make an optimum choice.

Some of these properties are:

a. Hardness

b. Friction

c. Volume change


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d. Compression set

e. Tensile strength

f. Elongation probability of change

g. Tear strength and abrasion resistance

h. Thermal effects and heat resistancei. Squeeze and anti extrusion property

j. Stretch or tensile strength

k. Coefficient of thermal expansion

l. Permeability

m. Oil compatibility

n. Ageing

o. Corrosion resistance

p. Ozone and weather resistance

q. Electrical properties

The following calculations are made from the Parker manual for O-Rings, being widely

used for seal selection.

O-rings can be molded in a wide range of compounds in hardness from 40 to 95 Shore A.

These materials include:

Acrylonitrile-Butadiene (NBR)

Butyl (IIR)

Chloroprene (CR)

Ethylene Acrylic (AEM)

Ethylene Propylene (EPDM)

Fluorocarbon (FKM)

Fluorosilicone (FVMQ)

Hydrogenated Nitrile (HNBR)

Perfluoroelastomer (FFKM)

Polyacrylate (ACM)

Silicone (VMQ)


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Table 4.1 Temperature Range for common elastomers

(P.T.O)


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Fig 4.3 Compressibility Chart

Table 4.2 Comparison of commonly used elastomers


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We select Polyurethane due to following reasons

(a)It has high abrasion resistance, low friction

(b)Suitable for high pressure, shock load

(c)Temperature range of -50 to 150 degree Celsius

(d)Shore hardness of 95-105A

According to the sizing charts, following range of products are available:

2-0XX (C/S=1.78 0.08)2-1XX (C/S=2.62 0.08)

2-2XX (C/S=3.53 0.10)

2-3XX (C/S=5.33 0.13)

2-4XX (C/S=6.99 0.15)

Since the bore diameter is required to be 8.7, we may select the O-ring from 2-0XX

series.

(P.T.O)


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Fig 4.4 2-0XX selection chart


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Considering the bore diameter to be 8.70.05, and cross section diameter to be 1.78, we

get the inner diameter to be 5.35 or 4.93. Now, if the limits are such that bore is of 8.65

and the c/s is 1.86, then ID of 5.35 would further compress the ring, providing better

leakage resistance. While the other case would result in loose ring, which would cause

leakage. Thus, we select a ring which has ID closer to 5.35.

So, we select the ring 2-009.

According to it, the inside diameter of the ring is 5.280.13

But in the case where,

Bore=8.75,

ID=5.15,C/S=1.70,

ID+2(C/S)=8.55


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CHAPTER 5

Conclusion and Future Scopes


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Conclusion:

During our work course and study, we have gained the understanding of the governing principles for the CPC, thereby analyzing different possible mechanisms for its optimumfunctioning.

Following the design procedure, facilitating failure safe functioning of the product, wehave reached till the selection of O-rings for static and dynamic operations of the CPC.

Also, in the coming months, we plan to complete the designing procedure thoroughly andcontinue to the next step, i.e. analysis of the proposed design.

Future plans

Gland Design

Valve orifice configuration

Spool design (detail engineering)

Force Calculations

Electronic controller of CPC


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References:

1. Spirax Sarco website ( www.spiraxsarco.com )2. APPLIED INSTRUMENTATION IN PROCESS INDUSTRIES Vol1a by

W.G. Andrew,3. Siemens website ( www.siemens.com )4. Parker O-Ring handbook 5. Hydraulics by S.R. Majumdar

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Final Project Report CPC

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