Solenoid Operated Piston Pump Engineering Essay

Solenoid Operated diver Pump Engineering EssayThis project is aimed at analysing and tendencying a solenoid operated plunger nitty-gritty which is cap open of delivering solution (this report assumes water) at a f modest rate of 1 litre/min. However, the customer usage requires the flow rate to retain mingled with 0.9 and 1.1 litre/min at an ambient shove of ab protrude 1 bar.The operation mode of the plunger pump is described down the stairs exploitation the diagram Oscill diverPumpFig 1.1 Solenoid Operated diver Pump1The solenoid coil (4) generates an electromagnetic field by the single wave diode rectified current flowing through the coil. for each(prenominal) 1 current pulse moves the plunger (5) against the pressure imprint (3). This reason reduces the spate in the suction chamber causing an increase in pressure (P a 1/V), which opens the valve (6) in the piston, thereby allowing the liquid to hold up into the pressure side. When the current acting on the soleno id pulse is off, the pressure restrain pushes back the piston toward the pressure side. The increase of pressure caused by the piston movement closes the piston valve (6) and the liquid flows through the valve (7) castigate in the pressure connection (8) and into the pressure pipe. The piston movement alike simultaneously increases the volume in the suction chamber, thereby reducing the pressure below the chamber. The low pressure in the suction chamber opens the valve (2) set in the suction connection (1), and the liquid is sucked into the pump and the cycle starts again. The piston size and the length of its displacement define the flow rate. The pump willing run with step up damage when the liquid flow is stopped momentarily1.This design concentrated on the piston, suction chamber and pressure gives design. Although references were made to the valves and solenoid ride, engineering compendium were non carried come in on them.CHAPTER 2INITIAL ENGINEERING DESIGN ANALYSISThi s indorsementtion considered the engineering digest of the operation of the piston pump to achieve the require specifications. The stipulation specifications beFlow rate Q = 1 Lit/minFrequency F = 60 cycles/secAmbient Pressure = 1 bar utilise the above specifications, the length of cut of the piston, which is also termed as the Swept Volume, can be cypher using the relation belowQ = Volumetime=Volume frequence= d2 L4 fL= 4Qd2fWhere Q = Flow respect =1 lit/min= 1.667 104 mm3/secf = Frequency (cycles/sec)L = Length of stroke/Swept volume (mm)d = Diameter of piston/suction chamber (mm)The diameter was varied from 1 to 20 mm and the be lengths of stroke were obtained at different frequencies of 40, 45, 50, 55 and 60 cycles/sec. The results obtained were plotted (See appendix 1). After cargonful look, the frequence at 40 cycle/sec, so subsequent calculations would be based on this. It was also noticed that reasonable pair of dimensions of the diameter and length occurred mor e or less the diameters 5-10mm, therefore subsequent calculations were based on this range.2.1 LOAD ANALYSISThe load depth psychology was carried out on each component designed as indicated belowA. piston The load analysis on the piston was done by isolating the piston and analysing the draw outs acting it. The different sop ups acting on the piston are as shown below take out on piston causing accelerationMagnetic jam from solenoid coil termination spring powerfulnessKinematic grindingal powerfulness gravitative attract result hydraulic take (including assumed honied effect)This is presumptuous that atomic, sign static frictional legions and temperature effects are negligible.The force analyses were carried out considering three different cases under which the pump operation can undergo. The intake and jut strokes were also analysed separately to reduce complications. The difference between the intake and ejection stroke is that, the magnetic force from the solenoid is zero during ejection, because the solenoid is offCase I This is when the piston pump is used swimmingly, that is, it is used to pump fluid on the same datum. This means that the gravitational effect and the height difference in the hydraulic force will be zero. The relationship between the forces will therefore beIntake strokeForce causing motion = Force from solenoid Resultant spring force Resultant hydraulic force Frictional forceEjection strokeForce causing motion = Resultant spring force Resultant hydraulic force Frictional forceCase II This considered the case when the pump is used to transfer fluid from a higher level to a lower level. This means that the gravitational effect will favour the direction of flow therefore reducing the force needed to drive the piston. The relationship between the forces will therefore beIntake strokeForce causing motion = Force from solenoid Resultant spring force Resultant hydraulic force Frictional force Gravitational forceEjection strokeForce causing motion = Resultant spring force Resultant hydraulic force Frictional force + Gravitational forceCase III This considered the case when the piston pump is used to deliver fluid from a lower level to a higher level. The difference between this case and case II is in the gravitational effect and the datum difference in the hydraulic effect. The design load analysis was done under this circumstance because pumps are usually used for this particular purpose. Even with this design concept, the pump can still be used for otherwise cases, but it might deliver fluid at a higher flow rate, which could still be in the boundaries of the given margin of the flow rate. The relationship between the forces will therefore beIntake strokeForce causing motion = Force from solenoid Resultant spring force Resultant hydraulic force Frictional force + Gravitational forceEjection strokeForce causing motion = Resultant spring force Resultant hydraulic force Frictional force Gr avitational force.The different forces were calculated as follows using the free body diagram of the piston shown belowFigure 2.1 Boundary conditions of intake and ejection strokesForce from solenoid coil= FsForce on piston causing motion = MpaWhere Mp = mass of piston kg and a = acceleration of piston (mm/s2)Mp= V = Density of material (Stainless steel) =810-6 (kg/mm3)V=Volume of fluid displced in one stroke mm3= Q t= Qfwhere f=45 cycles/sec=90 strokes/sec (2 strokes=1 cycle)Mp= Qf=810-6 1.667 10490=1.48210-3From law of motion v2= u2+ 2aSu = 0 a=v22S Also v= St= S fv=Velocity (mm/s) and S= L=Length of stroke (mm)a=L f22L= L f22= L 9022The length was varied from 5 to 10 mm, and different accelerations were obtained (See appendix 2).Resultant spring force = K2x- K1x= xK2- K1= xK Where K1 and K2=Stiffness of springs 1 and 2 respectively (N/mm)x=L=Stoke length (mm)Kinematic frictional force = kN= kMpgWhere k=Coefficient of kinematic friction N=Normal force= Mpgg=acceleration pay able to gravity=9810 mm/s2Gravitational force = MpgHydraulic force = Total neuter in Pressure P (N/mm2)Surface Area of Piston A (mm2)From Bernoulllis comparability P1g+ V122g+ Z1= P2g+ V222g+ Z2P= P1-P2=V22-V122+ ZgQ= A1V1= A2V2 , then(prenominal) V2= QA2= A1V1A2 and V1= QA1P= A1V1A22-V122+ Zg= V122 A1A22- 1+ ZgP= Q22A12A1A22- 1+ ZgWhere Q= Flow rate (mm3/s) , =density of water =110-6 (kg/mm3)A1and A2=Area mm2 and V1 and V2=Velocity (m/s)Z=L=Length of Stroke mmIncluding the campaign coefficient C = 0.98 to account for viscous effect, P therefore becomesP= Q22C2A12A1A22- 1+ Lg Hydraulic force = Q22C2A12A1A22- 1+ LgSurface Area of Piston A mm2= Q22C2A12A1A22- 1+ LgA2- A1The forces were algebraically added according the ejection stroke par developed above (case III) to obtain ?K at different diameter of pistons, fixing inner diameter of Piston D2 (corresponding to A2) = 0.5, 1, 1.5, 2 and 2.5mm (See appendix 3).Force causing motion = Resultant spring force Resultant hydrauli c force Frictional force Gravitational force.Mpa= L K- Q22C2A12A1A22- 1+ LgA2- A1- kMpg- MpgK= 1LMpa+ kg+g+ Q22C2A12A1A22- 1+ LgA2- A1The hydraulic effect is due to the fluid forced out from the suction chamber into the outlet. Therefore the A1 and A2 will be the area of the piston and the outlet, corresponding to diameters D1 and D2 respectively. Also the outlet diameter was assumed to be equal to the inner diameter of the piston.The results obtained for difference in stiffness ?K above, were used to obtain the force from solenoid coil Fs using the injection stroke par above. Also different diameter of piston were used objet dart varying the inner diameter of piston D2 (corresponding to A2) = 0.5, 1, 1.5, 2 and 2.5mm (See appendix 4).Considering the intake stroke equation for case IIIForce causing motion = Force from solenoid Resultant spring force Resultant hydraulic force Frictional force + Gravitational forceMpa= Fs-LK- Q22C2A12A1A22- 1+ LgA1- kMpg+ MpgFs= Mpa+ kg-g+LK + Q22C2A12A1A22- 1+ Lg A1The hydraulic effect is due to the change in pressure as the fluid passes through the piston, because of the reduction in area. Therefore the A1 and A2 will be the area of the piston outer and inner diameter, corresponding to diameters D1 and D2 respectively.B. Pressure Springs The load analysis of the spring was also done by isolating the spring and analysing the forces acting it. Considering the ejection stroke of upper spring (spring 1), the different forces acting on the spring are as shown belowForce on piston causing accelerationSpring forceResultant hydraulic force (including assumed viscous effect)This is assuming that the frictional force on spring is negligible because the surface area contacting the wall is small.Force causing motion = Spring force + Resultant hydraulic forceMpa= LK1+ Q22C2A12A1A22- 1+ Lg A1K1=1LMpa- Q22C2A12A1A22- 1+ Lg A1K2=K1+KWhere Force on springs Fsk=KLength of strokeThe values of stiffness of springs 1 and 2 were calcula ted using the relationships above at different outer and inner diameters of the piston. The graphs were plotted to see the variations (See appendix 5 and 6).C. Inlet Valve and Spring Considering also the inlet valves and analysing the forces acting it, the injection stroke is caused by an increase in volume of the suction chamber, causing a corresponding decrease in pressure. Therefore the different forces acting on the inlet valve are given belowInlet spring force at compressionResultant hydraulic force (including assumed viscous effect)This is assuming that the frictional force and gravitational force on the valve is negligible because the valve is light.Resultant Pressure Change= ?PFrom Gas Law P1V1= P2V2P1 and P2 are the initial and final pressures of both the inlet and suction chamber respectively (N/mm2). The initial pressure P1 is assumed to be equal to the external pressure which is given to be equal to the atmospheric pressure Pa = 1 bar = 0.1 N/mm2. That is why fluid is n ot flowing because there is no pressure difference, or P1 was higher than PaP2= P1V1V2= PaV1V2 where V2=V1+Vs and Vs=Swept Volume per stokeVs=Flow rateFrequency in stroke/sec=1.66710490 =185.22 mm2/strokeP2= P1V1V1+VsP1=Change in pressure due to swept volume= Pa-P2P1=Pa-PaV1V1+Vs=Pa V1+Vs-PaV1 V1+Vs=PaV1-PaV1+PaVsV1+Vs=PaVsV1+VsWhere V1 = VT and it is the total volume of the inlet spring area, suction chamber and the inner stead of the piston.P2=Pressure Change due to area changesP2= Q22C2A12A1A22- 1+ Lg The above pressure change is the sum of the pressure changes from the inlet through suction chamber and into pistons inner diameter. This is negligible because the pressure drops as it enters the suction chamber and increases as it enters the inner diameter of piston, thereby almost cancelling out.P=P1=PaVsVT+VsHydraulic force=spring force at compressionP1A3=PaVsA3VT+Vs= K3x3PaVs=K3x3A3VT+ K3x3A3VsVT=PaVs- K3x3A3VsK3x3A3 Where A3=Inlet area mm2, K3=Inlet Spring Stiffness (N/mm)and x3=Spring movement=Valve lifting mmThe values the total internecine volume VT was obtained at different values of the diameter of the inlet D3 (corresponding to A3). The value of the spring force K3x3 was varied from 0.01 to 0.05 N and the variations were plotted to see an appropriate one (See appendix 7).2.2 Component Design and pickax The component design has been carried out along with the load analysis shown above. The desired dimensions for different components were then selected after a careful study and analysis of the graphs plotted. The dimensions were selected based on those that satisfy the required specifications, reasonably able to be manufactured and can be selected from the manufacturers catalogue as in the case of the springs2. Below are the component dimensionsSolenoidSolenoid Frequency 45cycles/sec = 90 strokes/secForce from solenoid coil 108.8NLength of stroke 7.367 mmPistonPiston outer diameter 8 mmPiston inner diameter 2 mmSpringsPressure spring 1 rate = 5.771 N/mm Force on spring 1 = Rate * length of stroke = 5.771 * 7.367 = 42.515 NPressure spring 2 rate = 14.683 N/mm Force on spring 1 = Rate * length of stroke = 14.683 * 7.367 = 108.17 NFrom the above calculations and estimated values of the spring rates, the most accurate spring elect from the compression spring catalogue are (see appendix 8 and 9)Spring 1 C6609150Wire diameter 1.02 mmOuter Diameter 7.62 mm ingenuous length 15.88 mmRate 5.81 N/mmSpring 2 D22110Wire diameter 1.25 mmOuter Diameter 7.55mmFree length 17mmRate 15.03 N/mmInletInlet spring stiffness = 0.02 N/mmInlet spring length = 9.804 mmInlet diameter = 1.78 mm2.3 Stress Analysis The deform analysis was carried out on just two components as shown below. This was because these are the two components whose failure affects the pump operation most.A. Piston The two vehemencees acting on the piston are normal and gazump seekes which is given as.Stress (N/mm2) sij= Force (N)Area (mm2)The notation is to differentiate betwee n the direction and tack of action, where the first digit represents the plane of action and the second digit represents the direction of force. When the notations are different, it signifies shear stress and when the notations are the same it means normal stress.The force on the piston varies as the piston goes through the cycle, therefore the different forces and principal stresses were calculated as the spring compresses and stretches. This was shown in appendix 10 and 11, but the calculations of the maximum and minimum principal stresses at the springs peak are shown below. The principal stresses were calculated because they are the cause of fracture in a component3.Considering the piston and spring 1Fig 2.2 Stresses acting on piston from spring 1 and wall3s11= 0 because there is no horizontal force in that directions12= Force from SolenoidSurface area of piston= Fsp Do Lp= 108.8p815=0.2886 N/mm2Where D0=Outer diameter of piston mm, Lp=Length of Piston (mm)s22= Force from sprin g 1Outer Area-Inner Area= K1Lp4 Do2- Di2s22=5.771 7.367p4 82- 22= 42.51547.1239=0.9022 N/mm2s21= 0 because there is no horizontal force in that directionConsidering the piston and spring 2s11= 0 because there is no horizontal force in that directions12= Force from SolenoidSurface area of piston= Fsp Do Lp= 108.8p815=0.2886 N/mm2Where D0=Outer diameter of piston mm, Lp=Length of Piston (mm)s22= Force from spring 2Outer Area-Inner Area= K2Lp4 Do2- Di2s22=14.638 7.367p4 82- 22= 107.838147.1239=2.2884 N/mm2s21= 0 because there is no horizontal force in that directionThe total principal stress which is the usual cause of fracture was calculated using the total normal stresses from the springs and the shear stress from solenoid.Total shear stressesTs12=s12 from Spring 1+ s12 from Sprig 2=0.2886+0.2886= 0.5772Total normal stressesTs22=s22 from Spring 1+ s22 from Sprig 2=0.9022+2.2954= 3.1976Therefore the principal stressess11s22- s(s11+s22)+s2-s122=003.1976- s(0+3.1976)+s2-0.57722=0s2-3.19 76s-0.3331=0Principal stresses smin=-0.101 N/mm2, smax=3.2986 N/mm2B. Pressure Springs The major stress acting on the spring is shear stress acting on the coils. The force and consequentially the shear stress on the springs vary as the piston deflection (i.e. length of stroke) increases and decreases. The various forces and shear stresses were calculated and the graph plotted (see appendix 12). But the calculation of the maximum shear stress, which occurs at the full deflection is shown below4Fig 2.4 Force acting on spring4Shear stress tmax= 8FDWpd3Where F=Force on spring ND=Mean outer diameter of spring mmd=diameter of spring coil mmW = Wahl Correction Factor which accounts for shear stress resulting from the springs curvatureW=4C-14C-4+0.615CC=DdConsidering Spring 1Fmax= K1Length of stroke=5.7717.367=42.515 N/mm2D=7.62 mm and d=1.02 mm ?C=Dd= 7.621.02=7.4705W=4C-14C-4+0.615C= 47.4705-147.4705-4+0.6157.4705=1.1982tmax= 8FmaxDWpd3= 842.515 7.621.1982p1.023=931.113 N/mm2Considering S pring 2Fmax= K1Length of stroke=14.6387.367=108.17 N/mm2D=7.55 mm and d=1.25 mm ?C=Dd= 7.551.25=6.04W=4C-14C-4+0.615C= 46.04-146.04-4+0.6156.04=1.2506tmax= 8FmaxDWpd3= 8108.17 7.551.2506p1.253=1331.119 N/mm2CHAPTER 3INITIAL MANUFACTURING DESIGN ANALYSIS3.1 DimensionsThe dimensions of all the main components piston, springs, cylinder and valves had been obtained from the calculations and graphical analysis made above. However, the detailed dimensions of all components namely pump body (left and right side), cylinder and liners, piston, springs and valves are shown in the CAD drawing in appendix 13.3.2 TolerancesTolerance for Stroke LengthThe statistical tolerance of the stoke length was calculated using integral method, which is much more effective than an additional tolerance. Given the tolerance of the flow rate as 0.1litres/min, the tolerance of the frequency was assumed to be 5 cycles/sec under normal dispersion condition. The tolerance of the stroke length was calculated as f ollowsStandard deviation s=Tolerance3 Cp where Cp=process capability indexIn general manufacturing industry, a process capability index (Cp) of 1.33is considered acceptable. Therefore CpFlow rateQ=1 0.1 lit/min= 1.667 104 1.667 103mm3/sec Q=3.33 1033 1.33=8.356 102Frequency F= 45 5 cycles/sec (Assuming a Normal distributed variable) f=103 1.33=2.506Therefore the flow rate and frequency could be written asQ N 1.667 104 , 8.356 102 mm3/secf N 45 , 2.506 cycles/secQ = Volumetime=Volume frequency= d2 L4 fL= 4Qd2fUsing differential tolerance2= i=1nxi2 xi2L2= LQ2Q2+ Lf2f2+ Ld22d2L2= 4 1d2 f2Q2+ Qd2 f22f2+ Qd3 f2d2 2Tolerance=3 CpThe standard deviations and tolerances of the stoke length were calculated using the above equations, while varying the diameter from 1 to 20 mm, and the results were plotted out (see appendix 14).Tolerance for Piston Principal Stress Assuming a normally distributed around the maximum principal stress acting on the piston, the standard deviation and the toler ance of the maximum principal stress was calculated using the load distribution obtained in appendix 11.3=3.2918-0.5772=2.7146Tolerance=Cp3=1.332.7146=3.6104 N/mm2Upper and lower limit=3.61042= 1.8052 N/mm2Tolerance for Springs Shear StressAlso assuming a normally distributed around the maximum shear stress acting on the springs, the standard deviation and the tolerance of the maximum shear stress was calculated using the load distribution obtained in appendix 12.For spring 13=931.113-0=931.113Tolerance=Cp3=1.33931.113=1238.38 N/mm2Upper and lower limit=1238.382= 619.19 N/mm2For spring 23=1331.119-0=1331.119Tolerance=Cp3=1.331331.119=1770.39 N/mm2Upper and lower limit=1770.392= 885.195 N/mm23.3 Fits The components that are fitted into the cylinder, namely cylinder liner, piston springs 1 and 2 are almost of equal diameter. But because of the consideration of the fits and limits to give some allowance a transition fit was chosen from Data opinion poll 4500A British Standard selecte d ISO Fits-Hole Basis. Since it fell in between the nominal size of 0 6 mm, the transition fit selected was H700.015 for the hole and k60-0.009 for the shaft5.3.4 Material SelectionPiston and CylinderThe piston and the cylinder are to be made of stainless steel grade 431. This is due to the prevention of fracture which could be caused by principal stress. From the maximum principal stress obtained for the piston above (3.2986 N/mm2 = 3.2986 MPa), it is sure that the material which has a yield strength of 655 MPa will be able to prevent failure. Also the other reason for choosing this material is because of its high resistance to corrosion6. Since the piston and cylinder interacts with the fluid, which increases the tendency for corrosion to occur, it is quite safe to use a highly corrosion resistance material like this. It is also in truth easily machined in annealed condition. The properties of the stainless steel grade 431are shown in appendix 156.Springs The springs are to be ma de of stainless steel grade 316. This is also due to the strength of the grade in preventing fracture, breakage and buckling of the spring due to the shear stress acting on it. From the maximum shear stress calculated above (1331.119 N/mm2 = 1.331 GPa), it is sure that this grade of stainless steel with an elastic modulus of 193 GPa will be able to withstand the compression. The material is also highly corrosion resistant and relatively easy to machine. The other properties of the stainless steel grade 316 are shown in appendix 156.ValvesThe valves are to be made of polytetrafluoroethylene PTFE, which is a thermoplastic. This was chosen because the material has to be light and can easily float. Also, it has very low coefficient of friction, which reduces the fluid drag force and wears on the piston and spring.3.5 Surface FinishThe surface finishing chosen for the manufacturing of the parts was to be 0.8 m Ra. This is to reduce friction and rate of wear, because there are lots of p arts moving against each other. The grinding process requires a very great accuracy because it is a relatively delicate manufacturing process.3.6 Geometric ToleranceIn obtaining the tolerance of the components, since algebraic addition of tolerances is very unrealistic and will not be efficient, the tolerances of components that fit into each other were taken from the Data Sheet 4500A British Standard selected ISO Fits-Hole Basis5. These are show belowS/NoPartsDimensions (mm)Tolerances (mm)Drawings1Cylinder11.00+ 0.0152Cylinder liner8.00 0.0093Piston2.00 0.0064Spring 117.00 0.00153.7 Process SelectionThe manufacturing processes of the various parts of the pump will be very important aspects of the design.The parts to be manufactured are pump body, cylinder liners and piston. It will take a great deal of accuracy in carrying out the process, because the geometry of the parts is very delicate. Any wrong dimension will affect the output or operation of the pump.There are three steps i n manufacturing the components mentioned above. Firstly, all the components would be manufactured by casting, which would probably not give the accurate dimensions. Then a turning/boring process can then be carried out, using a CNC or lathe machines, to achieve better dimension. The last process is the surface finish, which gives a smoother and precise dimension.It is relatively easier to manufacture the components by this method because of the intricacies of the geometry and dimensions, and also the materials chosen are easily machined. The manufacturing process of the springs would not be considered in this report because they are provided by suppliers.CHAPTER 4DESIGN optimization4.1 Component Manufacturing Risk AssessmentComponent make upPump Body (Left Right Side)Calculation of qmDrawing number001mp = 1 1.6 = 1.6gp = 1.7 1 1 1 1.1 1.1 = 2.057Ajustable tol= Design tolmpgp= + 0.0151.6 2.057=+0.00455tp = 1.71 = 1.7sp = 1 1.3 = 1.3qm = 1.7 1.3 = 2.21Manufacturing variabili ty risk, qm = 2.21Material431 Stainless SteelManufacturing ProcessTurning/BoringCharacteristic explanationHoles at fondness to edgeCharacteristic Dimension8 mmDesign Tolerance+ 0.015Surface Roughness0.8m RaComponent NamePistonCalculation of qmDrawing number005mp = 1 1.6 = 1.6gp = 1 1 1 1 1 1.1 = 1.1Ajustable tol= Design tolmpgp= 0.0061.6 1.1=0.0034tp = 1.71 = 1.7sp = 1 1 = 1qm = 1.7 1 = 1.7Manufacturing variability risk, qm =1.7Material431 Stainless SteelManufacturing ProcessTurning/BoringCharacteristic DescriptionHoles at centre to edgeCharacteristic Dimension2 mmDesign Tolerance 0.002, -0.008Surface Roughness0.8m Ra The values of the component manufacturing risk analysis obtained above are considerably with a low risk. This shows that the processes chosen for the manufacturing of the components are acceptable.4.2 Failure Mode and Effects Analysis (FMEA)The failure mode and effects analysis (FMEA) is an analytical technique performed to ensure that all possible failure mo des of the piston pump have being identified and address. Below are the predicted failure modes of each components of the piston pump, the caused, effects and the suggested solutionsIt can be seen from the FMEA above that the spring breakage has the greatest severity, but the wear on all the components has the greatest risk priority number. This is because wear would be experience by the customer over time of use which made the risk priority number very high. Therefore, while desig